Planetary steering differential

Continuously driven vehicle steering differentials have heretofore been too complex and have undesirably incorporated a second cross shaft for counter-rotating a pair of oppositely located planetary elements. The planetary steering differential (10) overcomes these deficiencies by including first, second and third interconnected planetary mechanisms (12,14,16) for rotating a pair of output members (48,52) in the same direction and at the same speed in response to rotation solely of a first input member (57) and holding a second input member (63) stationary, and for rotating the output members (48,52) in opposite directions at the same speed in response to rotation solely of the second input member (63) and holding the first input member (57) stationary. The first input member (57) is driven by a transmission (58), and the second output member (63) is driven by a reversible steering motor (66).

DESCRIPTION 
1. Technical Field 
This invention relates generally to a differential for a vehicle cross 
drive or the like, and more particularly to a compact and simplified 
planetary steering differential. 
2. Background Art 
A large number of earthmoving tractors have been developed which have 
incorporated either clutch-and-brake steering mechanisms or geared 
steering mechanisms. In both of these classifications, however, the drive 
to the opposite tracks is discontinuous when making turns. This results in 
poorer operation of the vehicle than is desired during turns in marginal 
ground conditions. 
In some seasons of the year, for example, logging industry tractors with 
conventional clutch-and-brake steering mechanisms cannot be maneuvered to 
provide useful work because when one of the steering clutches is 
disengaged the power to that side of the tractor is interrupted. The other 
side of the tractor then has to pull the total drawbar load to maintain 
momentum. Unfortunately, soil strength under only one side is often 
inadequate to carry the total load so the powered track shears the soil 
and loses traction. Such a loss is a particular disadvantage when it comes 
to making relatively small steering corrections. On the other hand, 
clutch-and-brake steering mechanisms are widely used because they perform 
well in straight-ahead work applications and are simple in construction. 
In geared steering mechanisms, rather than disconnecting one track and/or 
bringing that track to rest, one track is driven at a lower speed than the 
other by having additional gear sets in the drive to each track. But these 
mechanisms are complex and costly in construction because duplicate 
planetary sets, gears, brakes and/or clutches are typically provided at 
both sides. Furthermore, since these mechanisms are discontinuous a lower 
mean track speed is provided during a turn. 
Another major group of steering mechanisms includes differential mechanisms 
in which drive is transmitted continuously to both tracks. The simplest 
form thereof is a braked differential, but these are rarely used because 
of relatively large power losses at the steering brake. Many of the 
disadvantages of the braked differential are obviated by controlled 
differentials. In such mechanisms engine power is not wasted in the 
steering brakes since power is merely transferred from the inner track to 
the outer track. But these also have disadvantages. One major disadvantage 
is that when the brakes are off these mechanisms act as simple 
differentials so that they depend on the reaction between the track and 
the ground to be equal for straightahead operation. Since this is often 
not true the vehicle tends to drift so that frequent corrections are 
required. Moreover, if one track loses traction the drive to the opposite 
track is reduced. 
A good many of the above mentioned problems can be overcome by the use of 
double differentials, triple differentials and equivalent mechanisms which 
usually can counter rotate the opposite output members for spot turns. 
Typically, two differentials or their equivalent are arranged in parallel 
with their output shafts interconnected by gearing. Usually, a main drive 
power path is provided to one of the differentials and a steering drive 
power path is provided to the other one of the differentials. In some 
instances the steering input shaft has been driven by a hydrostatic pump 
and motor system, with the ability to hold the steering input shaft 
stationary to prevent differential action and to assure straight ahead 
operation of the tractor, military tank or the like in an effective 
manner. One major deficiency thereof is that dual cross shafts and 
associated gearing are required so that the construction is not only 
complex and costly, but also an unnecessarily large housing is required 
for containment of the components. Another related deficiency is that most 
of these differential mechanisms have a dual power path steering input, 
which usually involves rotating one gear at one side of the cross drive at 
the same speed but in the opposite direction to a counterpart gear on the 
other side. 
Thus, what is desired is a compact steering differential having continuous 
drive capability to the opposite output members, a relatively simple 
construction arranged along a single axis, a single steering input power 
path, and preferably a built-in reduction or speed step-down capability 
for better matching the transmission input speed without the need for an 
additional drive train. 
The present invention is directed to overcoming one or more of the problems 
as set forth above. 
DISCLOSURE OF THE INVENTION 
In one aspect of the invention a planetary steering differential is 
provided having first and second input members, first and second output 
members, and planetary means for rotating the first and second output 
members in the same direction and at the same speed when only the first 
input member is rotated and the second input member is held stationary, 
and for rotating the first and second output members in opposite 
directions at the same speed when only the second input member is rotated 
and the first input member is held stationary. 
When the first and second input members of the instant differential are 
powerably driven at the same time by suitable transmission means and 
steering means respectively, the average speed of the output members is 
proportional to the speed of the transmission means, and the difference in 
the speed of the output members is proportional to the speed of the motor 
means. 
In accordance with another aspect of the invention a planetary steering 
differential is provided having three planetary mechanisms, each having 
first, second and third planetary elements, and one of the elements of the 
second and third planetary mechanism serving as a first input and another 
as a second input. Advantageously, transmission means are provided for 
rotating the first input and thereby an output element of each of the 
first and third planetary mechanisms at equal rates of speed for 
straight-ahead vehicle operation in response to holding the second input, 
and steering means are provided solely through a single input power path 
for rotating the second input and thereby the same output elements at 
different rates of speed for turning vehicle operation. 
In accordance with a further aspect of the invention a planetary steering 
differential includes a first planetary mechanism having a first ring 
gear, a first sun gear, and a first planet carrier, and a second planetary 
mechanism having a second ring gear, a second sun gear, and a second 
planet carrier, the first ring gear being connected to the second planet 
carrier, the first and second sun gears being connected for joint 
rotation, and the first sun gear and the second planet carrier serving as 
output members in response to powered rotation of one of the first planet 
carrier and the second ring gear. 
Basically, however, a family of three interconnected planetary mechanisms 
is presented wherein one planetary element is powerably driven by a 
transmission and one other planetary element is powerably driven by a 
variable speed and reversible steering motor. Operation of the steering 
motor in one direction of rotation causes the speed of the left output 
member to increase and the right output member to simultaneously decrease 
the same amount for a right turn, while rotation in the opposite direction 
causes an opposite reaction for a left turn. The first planetary mechanism 
in each embodiment in the family has a grounded or stationary planetary 
element and is provided for speed control. This greatly simplifies the 
construction of the differential, beneficially provides continuous power 
to the opposite sides of the vehicle during a turn, and allows 
straight-ahead operation without drift. It is especially valuable in a 
vehicle requiring either frequent small steering corrections or repetitive 
spot steering type operation under adverse footing conditions.

BEST MODE FOR CARRYING OUT THE INVENTION 
Referring initially to the diagrammatic drawing of FIG. 1, a planetary 
steering differential 10 is shown in simplified block-like form which has 
first, second and third interconnected planetary mechanisms 12, 14 and 16 
respectively aligned along a central vehicle cross drive axis 18 oriented 
normal to the usual forward and reverse direction of vehicle travel. 
The first planetary mechanism 12 includes ring gear, sun gear and planet 
carrier elements 20, 22 and 24 of the typical type wherein a plurality of 
similar planet elements 26 are rotatably mounted on the carrier element 
and are in intermeshing toothed engagement with the ring and sun elements. 
In the instant embodiment the first planetary mechanism serves as a 
grounded member train with the ring element being fixedly secured to a 
differential housing 28, and provides a speed reduction. An end view of 
the first planetary mechanism, looking from the right side of FIG. 1, is 
illustrated in FIG. 4. 
The second planetary mechanism 14 includes ring gear, sun gear and planet 
carrier elements 30, 32, 34 and a plurality of planet elements 36 are 
rotatably mounted on the carrier element. As is illustrated in FIG. 1, the 
second sun element 32 is directly connected to first sun element 22 by a 
shaft 38. A diagrammatic end view of the second planetary mechanism is 
illustrated in FIG. 3. 
Similarly, at the left side when viewing FIG. 1 is the third planetary 
mechanism 16 including ring gear, sun gear and planet carrier elements 40, 
42, and 44, and a plurality of planet elements 46 on the carrier element. 
An end view thereof is shown in FIG. 2. The third planet carrier element 
44 is connected for joint rotation with the second ring element 30, as is 
also the third sun element 42 with the second sun element 32. 
The first carrier element 24 on the right side when viewing FIG. 1 serves 
essentially as the first output member 48, and a conventional service 
brake assembly 50 can be associated therewith to selectively brake the 
right side of the vehicle. On the left side the third carrier element 44 
serves as the second output member 52 and a similar conventional service 
brake assembly 54 can selectively connect that element or output member to 
the differential housing 28 to brake the left side. 
Transmission means 56 is desirably provided for powerably rotating at least 
one element of the second and third planetary mechanism 14, 16 via a 
single reversible power path or first input member 57. Preferably the 
transmission means includes a conventional multi or variable speed and 
reversible transmission 58 driven by an engine 60 as is illustrated in 
FIG. 1. In the instant example, the transmission is essentially connected 
to the second carrier element 34 so that that element serves as the 
primary or first input. 
Advantageously, infinitely variable ratio steering means 62 is provided for 
powerably rotating at least one other element of the second and third 
planetary mechanisms 14,16 via a single reversible power path or second 
input member 63. Preferably, such means includes a hydraulic or 
hydrostatic pump 64 driven by the engine 60, and a corresponding motor 66 
which is in fluid communication with the pump and driven by the pump. In 
the instant example the hydraulic motor 66 is essentially connected to the 
third ring element 40 which thereby serves as the secondary input to the 
steering differential 10. 
Referring now to FIG. 8, a diagrammatic cross sectional elevation view of 
one form of the planetary differential 10 described in FIG. 1 is shown 
installed in a vehicle 68 and looking from the rear of the vehicle. 
Centrally thereof is a transmission case 70 releasably secured within the 
differential housing 28, and defining a central chamber 72 in which the 
multi-speed transmission 58 is contained. Although the conventional 
transmission is not shown, the output thereof drives a bevel gear 74 
releasably secured to a hollow cross shaft 76. The cross shaft is 
rotatably supported within the transmission case by first and second 
bearing assemblies 78 and 80, and extends outwardly from the left side of 
the case when viewing the drawing to provide a spline connection 82 
thereat. This spline connection couples the hollow shaft to the second 
carrier element 34 of the second planetary mechanism 14 to provide the 
primary input to the differential. 
More specifically, the second sun element 32 has an extended body portion 
84 defining an internal spline 86, and the center shaft 38 is coupled to 
the spline and extends rightwardly within the hollow cross shaft 76 to 
provide the drive to the first planetary mechanism 12. The sun element 22 
is shown as being integral with the center shaft. The first ring element 
20 is connected to the service brake assembly 50, and this assembly is 
releasably connected to the differential housing 28 at a cylindrical 
opening 88 by a plurality of fasteners or bolts 89. In turn, the right 
side service brake assembly rotatably supports an annular hub 90 by a 
bearing assembly 92, and the hub supports the right output member or axle 
shaft 48 at a releasable spline connection 94. Since the right axle shaft 
is also releasably connected to the first carrier element 24 at a spline 
connection 96 it may be withdrawn to the right for servicing independently 
of the service brake assembly. Advantageously, the entire service brake 
assembly 50, the hub 90 and the first planetary mechanism 12 can be 
removed for servicing substantially as a modular unit by screwthreaded 
release of the bolts 89. The shaft 38 can thereafter be removed 
rightwardly through the opening 88. 
The left side service brake assembly 54 has substantially the same 
construction and convenient serviceability features as the right side. In 
addition, after the cross shaft 38 is released from the right side via the 
opening 88, and after the left axle shaft 52 and the second planetary 
mechanism 14 are moved to the left, the transmission case 70 and 
transmission 58 contained therein can be withdrawn rearwardly as a unit 
from within the differential housing 28 through a rearwardly facing 
opening in the rear wall thereof, not shown. 
The bidirectional, variable speed hydraulic motor 66 can be serviced from 
the top of the differential housing 28. It can be removed independently of 
an output gear 100 because of a releasable spline connection 102 
therebetween. The output gear 100 is in toothed engagement with a larger 
gear 104 connected to the third ring element 40 and rotatably supported on 
the service brake assembly 54 by a bearing assembly 106. 
ALTERNATE EMBODIMENT 
While the construction illustrated in FIG. 8 is very desirable as described 
above, it is also possible to rearrange the components of the steering 
differential 10 in a central location so that the transmission 58' and all 
three cf the planetary mechanisms 12', 14', and 16' can be removed as a 
modular unit as may be visualized with reference to FIG. 9. In FIG. 9 
elements common to FIGS. 1 and 8 have been shown with similar reference 
numbers with an added prime indicator for convenient identification. In 
the alternate embodiment the transmission case 70' is releasably connected 
to a reduced size differential case 108, and the differential case 
nestably contains the three planetary mechanisms 12', 14' and 16' and 
rotatably supports the input bevel gear 74' through an opposed pair of 
roller bearings 110. A first transfer gear 112 is connected to the input 
bevel gear for joint rotation and is in driving contact with a second 
transfer gear 113. The second transfer gear is connected to the second 
carrier element 34' and is essentially rotatably supported within the 
differential case 108 as by a second opposed pair of roller bearings 114, 
and it is this second transfer gear that provides the primary input to the 
second planetary mechanism 14'. The motor 66' illustrated in broken 
outline drives the gear 104' to provide the steering input to the third 
ring element 40' through an intermediate member 116. 
In the FIG. 9 construction it is only necessary to withdraw the left and 
right axle shafts 52', 48' to the left and right respectively in order to 
subsequently permit the differential case 108 and transmission case 70' to 
be withdrawn through the rear of the differential housing 28'. 
INDUSTRIAL APPLICABILITY 
In operation, the planetary steering differential 10 is preferably situated 
at the rear portion of the vehicle 68 such as an earthmoving tractor in 
order to place the transverse axis 18 thereof substantially in line with, 
or near, the axis of the sprockets that drive the left and right endless 
tracks. Although these latter members are not illustrated in the drawings, 
it can be appreciated by reference to FIG. 1 that if the left and right 
output members 52,48 are rotated in the same direction and at the same 
speed then the vehicle will travel straight-ahead or in reverse in a 
longitudinal direction. Either mode of operation is achieved by selecting 
the gear speed and direction of the output of the transmission 58 so that 
the second carrier element 34 is driven in the desired manner. 
Simultaneously, the hydraulic motor 66 is selectively positioned in a 
holding mode of operation to hold the third ring element 40 stationary and 
to assure that both output members 48,52 will rotate at the same speeds 
without drift. 
Specifically, and with reference to FIG. 3, the second planet carrier 
element 34 can be rotated in a counterclockwise direction when viewing 
along the axis 18 from the right side in FIG. 1 by the transmission 58 as 
is indicated by the letter A. By establishing the number of teeth in the 
rotating elements of the first, second and third planetary mechanisms 
12,14 and 16 at preselected values certain advantages will accrue. This 
can best be appreciated by selectively establishing the ratio of the 
number of teeth of the ring gear element divided by the number of teeth of 
the sun gear element, hereinafter called the e value, such that the e 
value of the second planetary mechanism 14 minus 1 is equal to the e value 
of both the first and third planetary mechanisms 12 and 16. For example, 
the number of teeth in the first planetary mechanism ring and sun elements 
20,22 respectively can be 81 and 27, which results in an e value of 3. The 
number of teeth in the second planetary mechanism ring and sun elements 
30,32 can be 80 and 20, which results in an e value of 4. And the number 
of teeth in the third planetary mechanism ring and sun elements 40,42 can 
be 81 and 27, which provides an e value of 3. 
With the aforementioned 3-4-3 e value relationship, and with the planetary 
mechanisms 12,14 and 16 interconnected as previously described, the ring 
element 30 will be forced to rotate in a counterclockwise direction as 
indicated by the letter B in FIG. 3, but at about 5/8 of the speed rate of 
the input. Also, the interconnected sun elements 22, 32, 42 will rotate 
together in a counterclockwise direction at 21/2 times the speed rate of 
the input as is indicated by the letters C in FIGS. 4, 3 and 2 
respectively. The first carrier output element 24 and the third carrier 
output element 44 will then rotate in a counterclockwise direction at a 
rate 5/8 of the input speed as indicated by the letters D in FIGS. 4 and 
2. Note that both the first and third ring elements 20,40 are held to 
achieve this straight-ahead mode of operation. 
If a steering correction is desired while traveling, the engine driven pump 
64 is selectively coupled to the hydraulic motor 66 to drive it in either 
direction of rotation simultaneously with rotation of the second carrier 
element 34 by the transmission 58. Assuming that a turn to the left is 
desired while traveling forward, then the motor will be powered to drive 
the third ring element 40 in a clockwise direction as is illustrated by 
the letter E in FIG. 5. This causes a reaction on the planet gear elements 
46 such that the speed of the third carrier element 44 is reduced as 
indicated by the letter F. On the other hand the sun elements 42, 32 and 
22 experience a speed increase as indicated by the letter G in FIGS. 5, 6 
and 7 with the result that the first carrier element 24 is forced to speed 
up as indicated by the letter H. Since the speed of the left output member 
or drive axle 52 decreases and the right output member or drive axle 48 
increases a steering correction to the left results. 
By reversing the direction of rotation of the hydraulic motor 66 a steering 
correction to the right can be made. Advantageously, power is continuously 
supplied to both sides of the vehicle 68 throughout a turn with the 
planetary steering differential 10, and the average speed of the opposite 
output members 48,52 remains unchanged and equal to a preselected 
proportion of the output speed of transmission 58. 
In any mode of operation, the service brake assemblies 50,54 can be 
operated together to brake the vehicle. Such operation is desirably 
independent of the steering function. 
In order for the vehicle 68 to make a spot turn the transmission 58 is 
selectively controlled to hold the second carrier element 34 stationary. 
Then the motor can be operated in either direction to force the first and 
third carrier elements 24 and 44 to rotate in opposite directions at the 
same speed. Under these conditions the demands on the hydraulic motor can 
be expected to reach a maximum horsepower rating. However, with the 
subject differential 10 the size or capacity of the motor need not be 
equal to the full horsepower rating of the engine 60, but rather only a 
preselected portion thereof. For example, a motor horsepower rating of 
about 60% that of the engine is considered practical in one instance. 
It is to be appreciated that the motor speed can be selectively controlled 
within limits to be large enough relative to the transmission input speed 
to cause any desired speed ratio of the opposite output members to give 
desired steering correction. Moreover, the output members can even be 
counterrotated when the transmission input member 57 is being rotated at a 
preselected speed for a more rapid steering correction. 
The differential 10 of FIG. 1 also provides a desirable speed reduction of 
about 1.6, which can be considered in the range of a moderate reduction 
level. A greater level of speed reduction, for example a reduction level 
of about 3.0, can be advantageous in some cases because it could mean that 
less speed reduction would be required within the final drives and result 
in a savings in construction complexity and space requirements thereat. 
SECOND ALTERNATE EMBODIMENT 
As shown in FIG. 10, a second alternate embodiment planetary steering 
differential 10' includes first, second and third interconnected planetary 
mechanisms 12', 14' and 16' respectively, so that it bears a family 
resemblance to the embodiment of FIG. 1. Each of the first, second and 
third planetary mechanisms have ring, sun, carrier and planet elements 
identified by reference numerals similar to those employed during the 
description of the first embodiment, only with the addition of a prime 
indicator. However, the connections are different and the e values are 
different to provide certain advantages. The first planet carrier element 
24' is connected for joint rotation with the second planet carrier element 
34', the first sun gear element 22' is connected to the third sun gear 
element 42' for joint rotation, and the second ring gear element 30' is 
connected for joint rotation with the third planet carrier element 44'. 
Both the transmission 58' and the motor 66' are connected to the same 
planetary mechanism 14' via pinion gears and bevel gears 118,120 and 
122,124 respectively, and with the bevel gears 120 and 124 being 
respectively connected for joint rotation with the second ring element 30' 
and the second sun element 32'. This construction can have certain 
advantages by better accommodating the differential 10' and particularly 
the location of the transmission and motor input elements 118, 122 to the 
space requirements of the vehicle. Moreover, the first carrier element 24' 
provides the right power output and the third ring element 40' provides 
the left power output, and this is preferred generally over a sun gear 
power output. The respective e values for the first, second and third 
planetary mechanisms 12', 14', 16' can be 2, 2 and 3 to provide gearing 
elements of a desirable size and a speed reduction at the output elements 
24',40' of about 1.5:1 with respect to the input element 30'. 
THIRD ALTERNATE EMBODIMENT 
A third alternate embodiment planetary steering differential 10" is shown 
in FIG. 11 having elements similar to those employed above identified with 
similar reference numerals, only with the addition of double prime 
indicators. In this construction the three sun gear elements 22", 32" and 
42" are connected together for joint rotation, and the two ring gear 
elements 30" and 40" are connected together for joint rotation so that 
only two separate pieces need be manufactured to effect a considerable 
cost savings. Yet, the desirable carrier element output members are 
retained, and a desirable range of e values can be selected to provide the 
required ratios. For example, e values of 4, 2 and 4 can be chosen for the 
respective mechanisms 12", 14" and 16" to provide a reduction level of 
1.67:1 in the differential 10". 
FOURTH ALTERNATE EMBODIMENT 
A fourth alternate embodiment planetary steering differential 10"' is shown 
in FIG. 12 which has the desirable feature of a increased speed reduction. 
In this instance the first planetary mechanism 12"' has a different 
grounded element, namely the first planet carrier element 24"' is 
continually grounded to the stationary housing 28"' and the ring element 
20"' serves as the right output member. The sun gears 22"', 32"' and 42"' 
are connected together for joint rotation, and the left output member is 
connected for joint rotation with the third planet carrier 44"' and also 
the second ring gear element 30"'. The transmission 58"' is coupled to the 
third ring gear element 40"' in this construction, and the steering motor 
66"' is coupled to the second carrier element 34"'. With e values of 3, 3 
and 2 for the first, second and third planetary mechanisms 12"', 14"' and 
16"' respectively, a relatively high speed reduction capability of 3.0:1 
is effected. It also features manufacturing economies similar to those 
mentioned above with respect to FIG. 11. 
FIFTH ALTERNATE EMBODIMENT 
FIG. 13 illustrates a fifth alternate embodiment planetary steering 
differential 10""which features a negative overdrive. 
The first carrier element 24""is continually grounded to the housing 
28""and the right output is by way of the first sun gear element 22"". The 
first ring element 20"", the second carrier element 34"" and the third 
ring element 40"" are connected together for joint rotation, and the 
second and third sun gear elements 32"" and 42"" are connected together to 
provide the left output. Transmission 58"" drives the second ring element 
30"" and steering motor 66"" drives the third planet carrier element 44"". 
With the respective e values being 4, 3 and 4, a speed reduction of -0.67 
can be effected. In other words, if the transmission drives the ring 
element 30"" at 1 rpm in a first direction of rotation with such 
construction then the output members 48"" and 52"" are driven at 11/2 rpm 
in a direction opposite to the first direction of rotation. 
Thus it can be recognized that the planetary steering differentials 10, 
10', 10", 10"' and 10"" present a family of differentials having a single 
transmission input power path and a particularly desirable single steering 
motor input power path to planetary means 12, 14, 16. This avoids the need 
for an additional cross shaft and the dual steering motor input power 
paths required of a great many of the prior art planetary steering 
differentials. Basically, three planetary mechanisms achieve rotation of 
the opposite output members in the same direction and at the same speed in 
response to holding the single steering input member and driving the 
transmission input member for straight ahead operation, or achieve 
rotation of the output members in opposite directions at the same speed in 
response to holding the transmission input member and driving the single 
steering input member for effecting a spot turn. This is accomplished in a 
compact and economical package, which package may be conveniently tailored 
in its speed reduction capability and construction to best match a 
particular vehicle's requirements. 
Other aspects, objects and advantages of this invention can be obtained 
from a study of the drawings, the disclosure and the appended claims.