Transfer case with dual planetary gear reduction unit
BACKGROUND OF THE INVENTION 
     This invention relates to a transfer case for four-wheel drive vehicles and more particularly to an arrangement wherein a dual planetary gear reduction unit delivers torque to its output shaft at a two-to-one ratio. 
     The U.S. Pat. No. 4,677,873 issued July 7, 1987 to Eastman et al. entitled Transfer Case With Inter-Axle Dual-Planetary Differential discloses a dual-planetary arrangement operative as an interaxle differential. The U.S. Pat. No. 3,060,765 issued Oct. 30, 1952 to Rinsoz discloses a differential gear including a toothed driving crown wheel rigid with the casing. A planetary gear arrangement is housed inside wherein two planet pinions mesh with sun wheels keyed to the ends of two aligned driven shaft elements. 
     SUMMARY OF THE INVENTION 
     The transfer case assembly of the present invention involves a dual planetary gear reduction unit uniquely designed to deliver output torque at a two-to-one ratio. The assembly comprises a outboard carrier ring grounded to the transfer case housing and an inboard grounded carrier ring formed with a central bore concentrically encircling a shiftable clutch collar in a radially spaced manner. The clutch collar has internal splines slidably engaging external splines on the output shaft. A planetary annulus gear, formed with an inboard axial extension concentrically surrounding the inboard carrier ring, is driven by dual planetary gear sets journally supported between the carrier rings. An annulus plate, concentrically disposed about the output shaft, has peripheral gear teeth in meshing engagement with internal teeth formed on the annulus gear extension. The annulus plate includes a central hub portion extending axially outboard therefrom and formed with an internally splined bore. 
     The annulus plate hub portion is concentrically interposed in spaced relation between the shiftable clutch collar and the inboard carrier ring central bore defining an annular clearance therebetween. A bushing is mounted on the annulus plate hub portion in the annular clearance so as to journally support the inboard carrier ring thereon. 
     The transfer case clutch collar may be shifted into a high or direct drive range position wherein it concentrically interconnects the input shaft integral sun gear directly to the output shaft. Upon sliding the clutch collar to its neutral position the input shaft integral sun gear is uncoupled from the output shaft. Upon further shifting the clutch collar to the low-range position it concentrically interconnects the annulus plate hub portion internally splined bore with external splines on the clutch collar thereby delivering low-range power to the output shaft. 
     The dual planetary has its carrier grounded to the transfer case housing with the carrier supporting inner and outer sets of intermeshing planetary gears. The inner set of planetary gears mesh with the sun gear teeth while the outer planet gears mesh with the annulus internal teeth. The sun gear teeth are formed on the input shaft and have a predetermined number of teeth exactly one half the number of annulus gear internal teeth. As a result the annulus gear and the annulus plate rotate in the same direction as the input and output shaft. With the transfer case collar shifted into its low-range drive mode a two-to-one drive torque ratio transfers power from the input shaft integral sun gear to the dual planetary gears and thence, via the annulus gear and the annulus plate, through the clutch collar to the output shaft. 
     Other objects and advantages of the invention will be made apparent from the accompanying drawings and the detailed description of the preferred embodiment of the invention. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a fragmentary cross-sectional view of a portion of a transfer case showing a dual planetary gear assembly adapted to engage a power takeoff unit input adapter gear; 
     FIG. 2 is a fragmentary vertical cross-sectional view taken on line 2--2 of FIG. 1 with parts broken away; and 
     FIG. 3 is a fragmentary cross-sectional view taken on line 3--3 of FIG. 2 showing the dual planetary carrier grounding structure located intermediate adjacent sets of planetary pinion gears. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Referring now to the drawings, a portion of a transfer case constructed in accordance with the present invention is generally indicated at 10 in FIG. 1. A description of one type of a vehicle transfer case may be found in the above mentioned U.S. Pat. No. 4,677,873 assigned to the assignee of the instant application the disclosure of which is incorporated by reference herein. As shown in FIG. 1 a transfer case housing 12 includes an end wall 14 with a side wall 16 extending inboard therefrom in an axial direction. 
     The end wall 14 has an annular hub portion 18 formed therein having an axial bore 19 supporting an input shaft 20 thereon by means of a double row annular contact ball bearing assembly 22. The input shaft 20 terminates in an outwardly stepped integral axially extending end flange portion 23 joined to the input shaft 20 by an intermediate stepped radially extending flange portion 24. An output shaft 26 is aligned on the longitudinal principal axis of the input shaft 20 and has a pilot portion 28 journally received in input shaft bore 30 by needle bearings 32. An end plug 33 closes the bore 30. The output shaft 26 has a clutch collar 34 slidably supported thereon by means of collar internal splines 36 engaged with output shaft external splines 38. 
     The clutch collar 34 has external clutch teeth 40 shown slidably engaged with internal clutch teeth 42 provided in axial counter-bore 44 formed in the inboard end of the input shaft 20. Torque is transferred directly from the input shaft 20 through the engaged teeth 40, 42 and the collar internal splines 36 engaged to the external splines 38 of the output shaft 26, thereby establishing a first or direct high-ratio drive range shown by shift collar position &#34;H&#34;. 
     As seen in FIGS. 1 and 2, the input shaft axial flange portion 23 has formed thereon helical sun gear teeth 46 of a helical planetary gear assembly generally indicated at 50. A planetary carrier is provided comprising outer 56 and inner 58 axially spaced ring members. The carrier ring members 56 and 58 support first inner and second outer sets of circumferentially spaced planet pinion gears. In the disclosed form the first inner set consists of four inner planet pinion gears 60 and the second outer set consists of four outer planet pinion gear 62. It will be noted that the inner and outer pinion gears 60 and 62 are each equal in diameter and formed with the same number of identical gear teeth, i.e., twenty-one gear teeth on each pinion gear 60 and on each pinion gear 62 in the disclosed embodiment. 
     The inner pins 64 have their pin axes located at a first radius &#34;R1&#34;, a predetermined lesser radial distance outwardly from the axis 63 of the output shaft 26 than the radius &#34;R2&#34; of the outer pins 66. Thus, the sun gear teeth 46 are in meshing engagement with the teeth of each inner planet gear 60 while internal teeth 68 of annulus gear 70 are in meshing engagement with the teeth of each outer planet gear 62. Further, the first and second sets of pinion gears 60 and 62 are disposed in paired meshing engagement such that each first set inner pinion gear 60 drives its associated paired second set outer pinion gear 62. In the preferred embodiment there are four pinion gears 60 in the first inner set and four pinion gears 62 in the second outer set each defining four pair of intermeshed pinion gears with each pair uniformly spaced at ninety degree intervals. 
     It will be noted that the inner pin 64 is smaller in diameter than the outer pin 66 while the helical teeth of the planet pinion gears 60 and 62 are identical in size and pitch diameter. Thus, the central bore 72 of each inner planet pinion gear 60 is smaller than the central bore 74 of the outer pinion gear 62. The difference in bore and pin sizes between the planet pinion gear sets allows ready identification during assembly. 
     As seen in FIG. 1 the annulus gear 70 concentrically surrounds the two sets of respective inner and outer pinion gears 60 and 62. It will be noted that the annulus gear internal gear teeth 68 are in meshing engagement with the teeth of each of the four outer pinion gears 62, while the sun gear teeth 46 are in meshing engagement with each of the four inner pinion gears 60. The annulus gear 70 incudes an axially directed inboard extension portion 78 formed with an internal annular counterbored race 80 receiving in a spaced concentric manner outer periphery 82 of the inboard carrier ring 58. Annular washer 84 is interposed between the inboard carrier ring 58 and race vertical shoulder 86 of the annulus gear while annular washer 88 is interposed between the inboard carrier ring 58 and opposed face 89 of an annulus plate, generally indicated at 90. Annular outer and inner washers 92 and 94 respectively, are provided on the pinion pins 64 and 66. 
     The annulus plate 90 is formed with an integral central cylindrical hub portion 96. The annulus plate hub portion 96 has an axial concentric internally splined bore 98 adapted for engagement with the clutch collar external splines 40 upon sliding the collar 34 to the right to its phantom low range position indicated by the dashed line &#34;L&#34; in FIG. 1. It will be seen that the inboard carrier ring 58 has its inner concentrically formed central opening 100 journally supported on an annular hydro-dynamic bushing 102. The bushing 102 is sized for press-fit reception on concentric annular exterior surface 103 of the hub portion 96. 
     The annulus plate 90 has it outer periphery formed with external splines 104 engaged with internal splines 105 provided on the annulus gear inboard extension portion 78. An expanding snap ring 106 axially retains the annulus plate 90 splined to the annulus gear. The outer periphery of the annulus gear extension portion 78 is formed with external power takeoff helical gear teeth 110. The medial transverse plane of the annulus power takeoff gear teeth 110 is defined by dashed construction line 112 shown in FIG. 3. It will be noted that the transverse vertical plane of the construction line 112 also coincides with the medial plane of the annular hydro-dynamic bushing 102. 
     The annulus gear external power takeoff gear teeth 110 extend outwardly adjacent to a rectangular access opening 120 (FIG. 1) in the housing side wall 16 so as to be aligned radially therewith. The opening 120 is defined by a peripheral frame 122 defining a planar mounting interface surface 123. 
     A conventional power takeoff gear box unit or assembly, generally indicated at 124 in FIG. 1, is adapted to be mounted on the interface surface 123 such as by machine bolts (not shown). A peripheral seal 126 is provided between the interface surface 123 and the mating gear box surface to seal housing opening 120 during operation. The gear box assembly 124 includes a longitudinally extending power takeoff shaft 128 journally supported in its case 130 with the shaft 128 oriented parallel with the transfer case axially aligned input 20 and output 26 shafts. The takeoff shaft 128 has a helical toothed takeoff gear 132 formed with internal splines slidably mounted on external splines 134 of the shaft 128 for rotation therewith. The gear 132 is sized such that its helical gear teeth 136 extend radially through the opening 120 and are positioned to slidably mesh with the annulus gear peripheral teeth 110. Thus, the helical gear 132 is shifted axially rightwardly or inboard by fork mechanism 138 to its dashed line position for driving engagement with the annulus gear extension external helical gear teeth 110. 
     With reference to FIGS. 2 and 3 it will be seen that the carrier outer and inner rings 56 and 58 are fixed to the transfer case housing 12 by means of spaced machine screws 140. Each machine screw extends through one of a plurality of boss portions 142 formed integral with inner carrier ring 58. Each carrier boss portion 142 has an axial bore 144 aligned with a hole 146 in the outer carrier ring and a blind bore 148 formed in a mating boss 149 in the housing wall portion 14. Thus, the dual planetary carrier rings 56 and 58 are prevented from rotating thereby providing a dual planetary with a grounded carrier. 
     As viewed in FIG. 1 a unique advantage of the present design results from locating the external power takeoff teeth 110, formed on the extension portion 78 of the annular gear 70, in radial alignment with the bushing 102 as shown by the transverse plane defined by construction line 112. It can be seen that the load imposed by the power takeoff gear 132 on the annulus gear extension external teeth 110 are initially transferred to the annulus plate 90 through the engaged splines 104 and 105. This power takeoff gear loading is transferred from the annulus plate hub portion 96 to the bushing 102 and thence via an oil film on bushing 102 exterior surface 150 to the inboard carrier ring 58. As a result no eccentric loads from the power takeoff gear 132 are transferred to the planet pinion gear pins 64 and 66 thereby obviating the application of any unbalanced or &#34;tipping&#34; forces on the carrier assembly. The design thus prevents excessive gear teeth wear or stress conditions, such as tight meshing, for example, developing between the planetary gears 60, 62 during operation of the transfer case in its power takeoff mode. 
     The rotation of the input 20 and output 26 shafts are in the same direction and is conventionally counterclockwise, as viewed in FIG. 2 for output shaft 26 by the arrow 151, during forward travel of the vehicle. Thus, it will be noted that during forward travel of the vehicle the sun gear teeth 46 are rotated in a counterclockwise direction. As a consequence of the carrier rings 56 and 58 being grounded to the case housing 12 the grounded inner planet gears 60 are rotated in a clockwise direction as indicated by arrow 153 in FIG. 2 and the grounded outer planet gears 62 are rotated in a counterclockwise direction as indicated by arrow 154 in FIG. 2. The result is the ungrounded annulus gear 70 is rotated by the outer pinion gears 62 in a counterclockwise direction 5 indicated by arrow 155 in FIG. 2. 
     In the disclosed dual planetary embodiment a total number of &#34;48&#34; sun gear teeth 46 are provided in conjunction with a total number of &#34;96&#34; annulus gear teeth. By virtue of applicants&#39; unique arrangement of a dual planetary with grounded pinions and the annulus gear formed with twice the number of teeth as the sun gear an exact two-to-one output gear ratio is achieved with the transfer case collar shifted to its low-range drive mode &#34;L&#34;. 
     Thus the invention enables an exact two-to-one output gear ratio to be achieved wherein power is transferred from the input shaft 20 to the output shaft 26 upon the clutch collar 34 being shifted to its third or low-range drive ratio position &#34;L&#34; wherein the drive torque transfers power from the input shaft 20 to the output shaft 26 via the annulus plate 90. This occurs because in low-range drive the clutch collar external splines 40 are engaged with the annulus plate internal splines 98. The annulus gear 70 thus always rotates in the same direction as the input shaft 20 because of the grounded carrier arrangement and the constant meshing of the outer pinions 62 with the ungrounded annulus gear internal teeth 68 together with the constant meshing of the inner pinions with the sun gear teeth 46. 
     Although only one embodiment of the subject invention has been illustrated and described, modifications and variations thereof will readily come to mind of a person skilled in the art which modifications and variations do not fall outside the scope of the invention as defined hereinafter.