Transmission for four-wheel drive vehicles

Abstract: An improved geartrain includes a transmission unit (20) operatively attached to a longitudinally placed engine. The transmission unit incorporates a multi-speed transmission (42,44) and a lockable planetary differential (68) and further includes an input shaft (46), an output shaft (32,38), and connecting drive and driven gears. Two alternative torque paths are provided to a quill shaft (60) from the output shaft. One of these paths is a high-range path (64) and the other of these is a low-range path (66). The two torque paths are directed through a synchronizer clutch (62) between low-range and high-range. One end of the quill shaft (60) is connected to the low-high synchronizer clutch (62) and the other end is connected to the lockable differential (68,70). The differential end of the quill shaft is connected to the pinion carrier (280) of the differential. The torque delivered by the quill shaft is split by the differential when in its unlocked mode with a portion going to the rear wheels and a portion going to the front wheels. A front axle synchronizer clutch (72) assembly is provided to allow the front axle to be disengaged for two-wheel drive operation, in which case the differential is locked.

Description: BACKGROUND OF THE INVENTION The present invention relates generally to transmissions for use in four-wheel drive vehicles. More particularly, the present invention is directed to a transmission having a multi-speed geartrain and a power transfer arrangement contained in a single unit for establishing various two-wheel and four-wheel drive modes. As is known, the majority of four-wheel drive vehicles are equipped with a transfer case mounted to a multi-speed transmission for directing power from the engine to all four wheels. To accommodate different road surfaces and conditions, many transfer cases are equipped with a mode shift mechanism which p mits the vehicle operator to selectively de-couple one set of wheels for establishing a two-wheel drive mode in addition to the four-wheel drive mode. In addition, some transfer cases are also equipped with a gear reduction unit for providing a high-range (i.e., direct drive) and a low-range (i.e., reduced ratio drive) four-wheel drive mode. Finally, some transfer cases are equipped with an interaxle differential for permitting torque proportioning and speed differentiation between the front and rear drivelines of the four-wheel drive vehicle. In an effort to minimize the overall size of the drivetrain used in four-wheel drive vehicles, it has been proposed to utilize a transmission of a transax a-type normally used to drive the front wheels of a front wheel drive vehicle as a four-wheel drive geartrain. In particular, British Patent No. GB 2,035,930 A to Jones et al, teaches of rotating the front wheel drivetrain (i.e., engine and transaxle) 90° such that the transaxle outputs can be interconnected to the front and rear drivelines of the motor vehicle. While such an arrangement may provide an economical drivetrain layout for a full-time four-wheel drive vehicle, the practical applications for such an arrangement are severely limited. In particular, such an arrangement does not permit the vehicle operator to selectively shift "on-the-fly" between high and low four-wheel drive modes, nor can one of the outputs be selectively disconnected from its driveline for establishing a two-wheel drive mode. Accordingly, the need exists for a transmission which can be used in most conventional four-wheel drive applications and which permits the vehicle operator to selectively shift between the available drive modes and speed ranges. SUMMARY OF THE INVENTION It is therefore an object of the present invention to provide a transmission for use with four-wheel drive vehicles having a multi-speed geartrain and a power transfer mechanism integrated into a single unit. A further object of the present invention is to provide the power transfer mechanism of the transmission with a synchronized mode shift mechanism for permitting the vehicle operator to selectively shift "on-the-fly" between a four-wheel drive mode and a two-wheel drive mode. It is another object of the present invention to provide the power transfer mechanism of the transmission with a synchronized range shift mechanism for permitting the vehicle operator to shift "on-the-fly" between a four-wheel low-range drive mode and a four-wheel high-range drive mode. It is yet a further object of the present invention to provide the power transfer mechanism of the transmission with an interaxle differential for establishing full-time four-wheel high-range and low-range drive modes. As a related object, means are provided for selectively locking the interaxle differential to establish part-time four-wheel high-range and low-range drive modes. Accordingly, the improved transmission of the present invention includes a multi-speed geartrain including a input shaft, a mainshaft, and a plurality of constant-mesh gearsets arranged for selectively coupling the mainshaft to the input shaft for driven rotation at various speed ratios. The mainshaft can be selectively coupled to a power transfer mechanism for establishing two alternative power transmission routes. In particular, a synchronized range shift mechanism is provided for establishing a high-range transmission route and a low-range transmission route from the mainshaft to a quill shaft which, in turn, drives an interaxle differential. The torque delivered by the quill shaft is split by the interaxle differential between the front and rear drivelines to establish a "full-time" four-wheel high-range and low-range drive modes. Additionally, a lock-out mechanism is provided for permitting the interaxle differential to be locked for establishing "part-time" four-wheel high-range and low-range drive modes. Finally, a mode shift mechanism is provided for permitting one of the drivelines to be selectively disconnected from the transmission for establishing a two-wheel drive mode. BRIEF DESCRIPTION OF THE DRAWINGS Further objects, features and advantages of the present invention will become apparent to those skilled in the art from studying the following description and the accompanying drawings in which: FIG. 1 is a schematic view of a four-wheel drive motor vehicle equipped with a transmission according to the present invention; FIG. 2 is a sectional view of the transmission constructed according to a preferred embodiment of the present invention; FIG. 3 is an enlarged partial view of FIG. 2 showing the geartrain components of the transmission in greater detail; FIG. 4 illustrates a power transmission route through the power transfer mechanism of the transmission for establishing a full-time four-wheel high-range drive mode; FIG. 5 is a view, similar to that of FIG. 4, showing a power transmission route through the power transfer mechanism of the transmission for establishing a full-time four-wheel low-range drive mode; FIG. 6 illustrates a power transmission route through the power transfer mechanism of the transmission for establishing a part-time four-wheel high-range drive mode; FIG. 7 is a view, similar to FIG. 6, illustrating the power transmission route through the power transfer mechanism of the transmission for establishing a two-wheel high-range drive mode; and FIG. 8 is a table listing the combination of speed ranges and drive modes available with the transmission of the present invention. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT The present Invention relates to a multi-speed transmission for use In four-wheel drive vehicles. In general, the transmission integrates a multi-speed geartrain and a power transfer mechanism into a common housing and is adapted for use with a longitudinally-aligned engine. More particularly, the geartrain provides five forward speed ratios and a reverse drive which are delivered to the power transfer mechanism for distribution to the vehicle's front and rear drivelines. The power transfer mechanism includes a synchronized range shift mechanism for permitting the vehicle to be shifted "on-the-fly" between a high (i.e., direct drive) speed range and a low (i.e., reduced ratio) speed range. Additionally, the power transfer mechanism includes an interaxle differential interconnecting the output of the range shift mechanism to the front and rear drivelines for establishing full-time (i.e., "differentiated") four-wheel high-range and low-range drive modes. A differential lock-out mechanism is also provided for selectively locking the interaxle differential to establish part-time (i.e., "non-differentiated") four-wheel high-range and low-range drive modes. Finally, the power transfer mechanism includes a synchronized mode shift mechanism operable for selectively connecting and disconnecting the transmission from the one of the drivelines for shifting "on-the-fly" between a two-wheel drive mode and a four-wheel drive mode. As will be understood, the novel transmission to be disclosed hereafter is shown in a merely exemplary vehicular application to which modifications can be made. Referring to FIG. 1, a vehicle 10 is schematically shown which is suited for use with the present invention. Vehicle 10 has a front driveline 12 and a rear driveline 14 drivable from a drivetrain 16. Drivetrain 16 includes an engine 18 and a transmission 20. Engine 18 is mounted in an in-line or longitudinal orientation along the long axis of vehicle 10 and its output is coupled to the input of transmission 20. Front and rear wheels 22 and 24 are part of front and rear drivelines 12 and 14, respectively, and are connected at opposite ends of front and rear axle assemblies 26 and 28, respectively. Front axle assembly 26 includes a front differential 30 that is coupled to a front output 32 (FIG. 2) of transmission 20 via a front prop shaft 34. Likewise, rear axle assembly 28 includes a rear differential 36 that is coupled to a rear output 38 of transmission 20 via a rear prop shaft 40. As is conventional, one or more universal joints (not shown) may be installed on opposite ends of prop shafts 34 and 40, as required. With particular reference to FIG. 2, transmission 20 is shown to include two primary sections, namely, a multi-speed geartrain 42 and a power transfer mechanism 44. Geartrain 42 includes an input shaft 46 driven by the output of engine 18 through engagement of a conventional manually-operable clutch (not shown), a mainshaft 48, and a series of constant-mesh gearsets 50, 52, 54, 56 and 58. Each gearset can be selectively engaged for coupling mainshaft 48 to input shaft 46 for rotation at a predetermined gear or speed ratio. Thus, power from engine 18 is selectively delivered from input shaft 46 to mainshaft 48 through the gearsets. Power is thereafter delivered from mainshaft 48 to front and rear outputs 32 and 38 via power transfer mechanism 44. In particular, power transfer mechanism 44 of transmission 20 includes a quill shaft 60, a synchronized range shift mechanism 62 including a high-range constant-mesh gearset 64 and a low-range constant-mesh gearset 66, and an interaxle differential 68 interconnecting quill shaft 60 to front output 32 and rear output 38. Interaxle differential 68 permits speed differentiation between front output 32 and rear output 38 while delivering drive torque thereto from geartrain 42 at a predetermined torque split or ratio. Accordingly, full-time high-range and low-range four-wheel drive modes are provided by transmission 20. While interaxle differential 68 is specifically shown as a planetary-type gearset, it will be appreciated that any suitable equivalent differential device known in the transmission art can be used in substitution therefor. Power transfer mechanism 44 is also shown to include a lock-out mechanism 70 for selectively locking interaxle differential 68 to inhibit speed differentiation between front output 32 and rear output 38. Lock-out mechanism 70 is operable in an "open" mode to permit differentiation across interaxle differential 68 and establish the full-time four-wheel high-range and low-range drive modes. Lock-out mechanism 70 is also operable in a "Locked" mode to inhibit differentiation across interaxle differential 68 for establishing part-time four-wheel high-range and low-range drive mode. Finally, power transfer mechanism 44 of transmission 20 includes a synchronized mode select mechanism 72 for selectively connecting and disconnecting front output 32 and front prop shaft 34 for establishing a two-wheel drive mode in addition to the four-wheel drive mode. As best seen from FIGS. 2 and 3, first gearset 50 is shown to include a first input gear 74 that is secured to input shaft 46 and a first drive gear 76 that is rotatably mounted on mainshaft 48 via a suitable bearing assembly 78. First input gear 74 and first drive gear 76 are in constant-mesh so as to define a first power transmission path from input shaft 46 to mainshaft 48 at a first speed ratio. Second gearset 52 includes a second input gear 80 that is secured to input shaft 46 and a second drive gear 82 that is rotatably mounted on the mainshaft 48 via a suitable bearing assembly 84. Second Input gear 80 and second drive gear 82 are in constant-mesh so as to define a second power transmission path from input shaft 46 to mainshaft 48 at a second speed ratio. Input gears 74 and 80 are shown formed integrally on input shaft 46. Conversely, input gears 74 and 80 could have been fixed (i.e., splined) to input shaft 46 for rotation therewith. Third gearset 54 includes a third input gear 86 that is rotatably mounted on input shaft 46 via a suitable bearing assembly 88 and a third drive gear 90 that is secured to mainshaft 48. Third input gear 86 and third drive gear 90 are in constant-mesh so as to define a third power transmission path from input shaft 46 to mainshaft 48 at a third speed ratio. Fourth gearset 56 includes a fourth input gear 92 that is rotatably mounted on input shaft 46 via a suitable bearing assembly 94 and a fourth drive gear 96 that is secured to mainshaft 48. Fourth input gear 92 and fourth drive gear 96 are in constant-mesh so as to define a fourth power transmission path at a fourth speed ratio. Finally, fifth gearset 58 includes a fifth input gear 98 that Is rotatably mounted on input shaft 46 via a suitable bearing assembly 100 and a fifth drive gear 102 that is secured to mainshaft 48. Fifth input gear 98 and fifth drive gear 102 are in constant-mesh so as to define a fifth power transmission path at a fifth speed ratio. Drive gears 90 and 96 are shown formed on a stubshaft 104 that is splined to mainshaft 48 for rotation therewith. Drive gear 102 is shown splined to mainshaft 48. Obviously, one, two, or all three of drive gears 90, 96, and 102 could have been formed integrally on mainshaft 48. As noted, geartrain 42 is arranged to deliver drive torque at a selected speed ratio through one of the five forward power transmission paths from input shaft 46 to mainshaft 48. To provide means for delivering such drive torque from mainshaft 48 to transmission outputs 32 and 38, power transfer mechanism 44 includes synchronized range shift mechanism 62 which transfers drive torque through one of high-range gearset 64 and low-range gearset 66 to quill shaft 60. High-range gearset 64 includes third drive gear 90 which, as noted, is fixed to mainshaft 48, and a high-range drive gear 108 that is rotatably mounted on quill shaft 60 by a suitable bearing assembly 110. Third drive gear 90 and high-range drive gear 108 are in constant-mesh so as to define a high-range power transmission route from geartrain 42 to power transfer mechanism 44 having a high-range speed ratio of substantially 1:1. Thus, when high-range gearset 64 is selectively engaged, quill shaft 60 is directly or near-directly driven by mainshaft 48. Low-range gearset 66 includes a low-range input gear 112 that is secured to mainshaft 48 and a low-range drive gear 114 that is rotatably mounted on quill shaft 60 by a suitable bearing assembly 116. Low-range input gear 112 and low-range drive gear 114 are in constant-mesh so as to define a low-range power transmission route from geartrain section 42 to power distribution section 44 having a low-range speed ratio. Thus, when low-range gearset 66 is selectively engaged, drive gear 114 is driven at a slower rotational speed than mainshaft 48. Accordingly, the low-range speed ratio is less than the high-range speed ratio. Transmission 20 includes a housing assembly 118 having interconnected housing sections 120 and 122. Housing section 120 includes a bell housing portion 124 and an intermediate wall portion 126 which together define a clutch chamber 128. As is conventional, a releasable clutch (not shown) is housed in bell housing portion 124 and is coupled to first end 130 of input shaft 46. Housing section 120 also includes a gear housing portion 132 that defines an enlarged chamber 134 in which both geartrain 42 and power transfer mechanism 44 are mounted and which defines a common lubricating oil sump. Housing section 122 is mated with the open end of gear housing portion 132 by fasteners, such as bolts 136. First end 130 of input shaft 46 extends through an aperture 144 formed through intermediate wall portion 126 of housing section 120 and is rotatably supported therein by a suitable bearing assembly 146. A second end 138 of input shaft 46 extends through an aperture 150 formed in a support plate 152 that is shown fixed between housing sections 120 and 122, with second end 138 of input shaft 46 being supported for rotation therein by a suitable bearing assembly 154. As previously noted, mainshaft 48 is operatively arranged to transfer engine torque from input shaft 46 to quill shaft 60. Mainshaft 48 is rotatably supported at its first end 140 by a suitable bearing assembly 156. Likewise, mainshaft 48 is supported at its second end 142 by a suitable bearing assembly 158 that is retained in an aperture 160 formed in support plate 152. Quill shaft 60 is operatively arranged to selectively transfer engine torque from mainshaft 48 through one of high-range gearset 64 and low-range gearset 66 to interaxle differential 68. Quill shaft 60 Is rotatably supported at Its output end 162 by a suitable bearing assembly 164 which is fixed in an aperture 166 formed in support plate 152. Likewise, quill shaft 60 is supported at its input end 168 by a suitable bearing assembly 170 that is retained in an aperture 172 formed in a radial support plate 174. Support plate 174 is fixed within gear housing portion 132 by suitable fasteners, such as bolt 176. With particular reference to FIG. 3, an enlarged portion of the sectional view shown in FIG. 2 is provided for illustrating geartrain 42 of transmission 20 in more detail. In general, each gearset 50, 52, 54, 56 and 58 is associated with a selectively engageable synchronizing clutch for shifting between the various gear ratios. To this end, a first synchronizer clutch assembly 184 is shown operably installed between first gearset 50 and second gearset 52 on mainshaft 48 and includes a hub 186 fixed for rotation with mainshaft 48 and a clutch sleeve 188 that is supported for rotation with and bi-directional axial movement on hub 186. First synchronizer clutch assembly 184 is preferably of the double-cone type, as shown. It is to be understood that synchronizer clutch assembly 184 (as well as each of the synchronizer clutch assemblies subsequently discussed for effecting gear changes between input shaft 46 and mainshaft 48) may be of any suitable type conventionally used in multi-speed manual transmissions. When clutch sleeve 188 is moved from the centered neutral position shown to the position denoted by construction line "A", it couples first drive gear 76 to mainshaft 48, thus establishing the first power transmission path at the first speed ratio. Such movement of clutch sleeve 188 to its "A" position from its centered position causes speed synchronization between clutch sleeve 188 (and mainshaft 48) and first drive gear 76, thus allowing internal splines 192 on clutch sleeve 188 to move into driven engagement with clutch teeth 194 formed on first drive gear 76. Conversely, when clutch sleeve 188 is moved from the central neutral position to the position denoted by construction line "B", it couples second drive gear 82 to mainshaft 48, thus establishing the second power transmission path and the second speed ratio. Movement of clutch sleeve 188 to its "B" position from its centered position causes speed synchronization between clutch sleeve 188 (and mainshaft 48) and second drive gear 82, thus allowing internal splines 192 of clutch sleeve 188 to move into engagement with clutch teeth 196 formed on second drive gear 82. Clutch sleeve 188 is axially moveable relative to hub 186 in a well known manner by means of a first shift fork, partially shown at reference numeral 190, that is attached to a suitable gearshift mechanism (not shown). As is apparent, geartrain 42 of transmission 20 is a five-speed arrangement having three synchronized clutch assemblies that are manually shifted via the vehicle operator manipulating a gear shift lever and the clutch in a well-known manner. A second synchronizer clutch assembly 200, preferably of the single-cone type, is shown operably installed on input shaft 46 between third gearset 54 and fourth gearset 56. Synchronizer clutch assembly 200 includes a hub 202 fixed (i.e., splined) for rotation with mainshaft 48 and a clutch sleeve 204 that is supported for rotation with and bi-directional axial movement on hub 202. When clutch sleeve 204 is moved from its centered neutral position shown to its third gear position, as denoted by line "C", it couples third input gear 86 (and third gearset 54) to input shaft 46, thus establishing the third power transmission path at the third speed ratio. Conversely, when clutch sleeve 204 is moved to its fourth gear position, as denoted by line "D", it couples fourth input gear 92 (and fourth gearset 56) to input shaft 46 for establishing the fourth power transmission path at the fourth speed ratio. Clutch sleeve 204 of second clutch assembly 200 has internal splines 206 meshed with external splines 208 formed on hub 202 for causing concurrent rotation therewith. Clutch sleeve 204 is axially movable between its "C" and "D" positions by means of a second shift fork, partially shown at 210, that is also coupled to the gearshift mechanism. Clutch teeth 211 are formed on a gear hub 212 that is fixed to third input gear 86 and are selectively engageable with splines 206 on clutch sleeve 204 when, following synchronization, clutch sleeve 204 is moved to its "C" position for coupling third input gear 86 to input shaft 46. Similarly, clutch teeth 216 are formed on a gear hub 218 that is fixed to fourth input gear 92 and are selectively engageable with splines 206 on clutch sleeve 204 when, following synchronization, clutch sleeve 204 is moved to its "D" position for coupling fourth input gear 92 to input shaft 42. A third synchronizer clutch assembly 220, also preferably of the single-cone type, includes a hub 222 that is splined for rotation on second end 138 of input shaft 46. A third clutch sleeve 224 has internal splines 226 meshed for rotation with and axial movement on external spline teeth 228 formed on hub 222. Splines 226 on clutch sleeve 224 selectively engage clutch teeth 230 formed on a gear hub 232 that is fixed to fifth input gear 98 when clutch sleeve 224 is moved, following synchronization, to its "E" position for interconnecting fifth gearset 58 to input shaft 46 so as to establish the fifth power transmission path at the fifth speed ratio. Clutch sleeve 224 is moved to its "F" position in response to shifting of geartrain 42 into its Reverse drive mode for effectively stopping rotation of input shaft 42 to inhibit inertia-related gear clash. To this end, transmission 20 includes a reverse brake mechanism 235 having a brake cone 236 that is frictionally mounted on a conical surface of a brake plate 238 which, in turn, is non-rotatably fixed to housing section 122. Movement of clutch sleeve 224 toward its "F" position causes speed synchronization between input shaft 46 and housing section 122, thereby braking rotation of input shaft 46. Clutch sleeve 274 is axially moveable between its "E" and "F" positions by means of a third shift fork, partially shown at 240, that is also connected to the gearshift mechanism. Geartrain 42 includes a reverse assembly 242 having a reverse idler gear 244 slidably and rotatably mounted on a shaft 246. Shaft 246 has a first end secured to support plate 174 and its opposite end retained in a bore formed in intermediate wall portion 126. Reverse idler gear 244 is shown in its non-engaged position where further or excess forward movement is limited by engagement with wall portion 126. In the illustrated non-engaged position, reverse idler gear 244 is located between first input gear 74 and a reverse input gear 250 that is fixed to input shaft 46. When clutch sleeve 224 of third synchronizer clutch assembly 222 is moved to its braked position ("F"), reverse idler gear 244 is moved from its non-engaged position to a Reverse position (R), as shown in phantom in FIG. 2, whereat it meshes with reverse input gear 250 and a reverse output gear (not shown) that, in turn, is meshed with a reverse drive gear 252 fixed on or integral with first clutch sleeve 188. Rotation of reverse drive gear 252 causes concurrent rotation of clutch sleeve 188 and hub 186 which, in turn, rotates mainshaft 48. Obviously, first clutch sleeve 188 is in its neutral centered position when such a reverse shift Is attempted. Excess movement of reverse idler gear 244 beyond its Reverse position (R) is limited by engagement with support plate 174. With reference now to FIGS. 4 through 7, power transfer mechanism 44 of transmission 20 is shown for the purpose of illustrating the power transmission routes to outputs 32 and 38 from geartrain 42. In general, FIGS. 4 through 7 illustrate the power transmission routes through power transfer mechanism 44 of transmission 20 for establishing full-time and part-time four-wheel drive modes, and a two-wheel drive mode, all at both the high and low speed ranges. Specifically, these drawings illustrate the power transmission routes taken from mainshaft 48 through quill shaft 60 and interaxle differential 68 to front output 32 and rear output 38. In the construction shown, rear output 38 is a rear output shaft 256 that is adapted to be coupled to rear prop shaft 40. Likewise, front output 32 is a front output shaft 258 about which quill shaft 60 is journally supported and which is selectively engageable with front prop shaft 34 via mode select mechanism 72. With continued reference to FIGS. 4 through 7, synchronized range shift mechanism 62 is shown to include a bi-directional dual-cone synchronizer clutch assembly 260 having a hub 262 fixed to quill shaft 60 for rotation therewith and which is located between high-range drive gear 108 and low-range drive gear 114. Hub 262 includes external splines 264 which are adapted to slidingly receive internal splines 266 formed on a range collar 268. Thus, range collar 268 is rotatable with and axially moveable relative to hub 262 in a well known manner by means of a range fork, partially shown at 270. Preferably, range fork 270 is supported for movement on or with a range shaft (not shown). More preferably, movement of range fork 270 is controlled by a shift apparatus having a rotatable sector plate which can be selectively rotated for moving range fork 270 between its high and low speed range positions. The shift apparatus can be a manually-operated mechanical system coupling the rotatable sector plate to a shift lever, or an electronically-controlled arrangement wherein an actuator rotates the sector plate in response to the vehicle operator energizing a suitable switch. An exemplary sector-type shift apparatus which could be modified by skilled artisans for application to this invention is disclosed in commonly owned U.S. Patent No. 5,159,847 entitled "SECTOR PLATE FOR TRANSFER CASE" the disclosure of which is hereby incorporated by reference. Axial movement of range collar 268 from its central non-actuated Neutral mode position ("N") to its High-range position ("HI") results in speed synchronization between quill shaft 60 and high-range drive gear 108 which is driven by mainshaft 48 through third drive gear 90. Following synchronization, internal splines 266 on range collar 268 can move into meshed engagement with clutch teeth 272 formed on a clutch ring 273 that is fixed for rotation with high-range drive gear 108. Conversely, axial movement of range clutch collar 268 from its Neutral position ("N") toward its low-range position ("LO") results in speed synchronization between quill shaft 60 and low-range drive gear 114, thus allowing internal splines 266 of range collar 268 to move into meshed engagement with clutch teeth 274 formed on a clutch ring 276 fixed for rotation with low-range drive gear 114. Accordingly, range shift mechanism 62 permits the vehicle operator to shift "on-the-fly" between the high and low speed ranges. However, In most applications, the ability to shift from high to low on-the-fly will be limited to low vehicular speed conditions to avoid sudden and undesirable jerking. Such shift limitations can be predicated on the synchronizing torque capacity of synchronizer clutch 260 or via any other known shift inhibiting device. In its Neutral mode position ("N"), no drive torque is transferred from geartrain 42 to power transfer mechanism 44 since quill shaft 60 is de-coupled from mainshaft 46. As noted, movement of range clutch collar 268 can be controlled either manually (via a suitable mechanical shift apparatus) or electrically (via a suitable actuator controlled by a pushbutton) based on the particular vehicular application. With respect to the remainder of the power transmission route through power transfer mechanism 44, quill shaft 60 is coupled to differential 68 for transferring drive torque thereto. Based on the preferred construction shown, differential 68 is a planetary gearset having an input planet carrier 280, an output sun gear 282 coupled (i.e., splined) to front output shaft 258, and an output ring gear 284 operable coupled to rear output shaft 256. Planet carrier 280 includes a pair of laterally-spaced carrier rings 279 and 281, and a plurality of pinion gears 286 journally supported on pins 283 between the carrier rings which constantly mesh with both sun gear 282 and ring gear 284. In particular, forward end 162 of quill shaft 60 is splined to a tubular extension 302 of carrier ring 279 for directly driving planet carrier 280. Sun gear 282 is splined to a first end 288 of front output shaft 258. Ring gear 284 is integrally mated with a cylindrical drum 290 which, in turn, is coupled to an annular collar 292. In particular, collar 292 has a radial segment 294 connected to drum 290 and a hub segment 296 splined to rear output shaft 256. Collar 292 is supported in housing section 122 by a suitable bearing assembly 298. In FIGS. 4 and 5, interaxle differential 68 is shown in an "open" operating state such that speed differentiation between ring gear 284 and sun gear 282 is permitted to accommodate interaxle rotational speed differences between front driveline 12 and rear driveline 14. Moreover, the torque split ratio delivered between front output shaft 258 and rear output shaft 256 is based on the planetary gear geometry of the intermeshed components of interaxle differential 68, as is known in the differential art. To provide means for locking differential 68 in a "locked" state for inhibiting speed differentiation, lock-out mechanism 70 is shown to include a lock-out collar 300 splined for rotation with and axial sliding movement on tubular extension 302 of carrier ring 279 by means of collar internal splines 304 engaging external splines 306 formed on the outer peripheral of extension 302. Lock-out collar 300 is axially movable by means of engagement with a shift fork, partially shown at 310, which, in turn, is also coupled to the shift apparatus. Lock-out collar 300 has external clutch teeth 312 that are meshingly engageable with internal clutch teeth 314 formed on a clutch ring 316 that is fixed to drum 290. In FIGS. 4 and 5, lock-out collar 300 is shown positioned in its Open position ("O") whereat clutch teeth 312 are disengaged from clutch teeth 314 such that interaxle differential 68 operates in its "Open" state for establishing full-time four-wheel drive. That is, differential 68 splits the torque delivered by quill shaft 60 at a predetermined ratio between front output shaft 258 and rear output shaft 256 while permitting speed differentiation therebetween. In FIGS. 2, 6 and 7, lock-out collar 300 is shown position in its Locked position ("L") whereat clutch teeth 312 are meshed with clutch teeth 314 such that interaxle differential 68 operates in its "Locked" state for establishing part-time four-wheel drive. In this position, ring gear 284 (and rear output shaft 256) is directly coupled for rotation with planet carrier 280 and quill shaft 60, thereby inhibiting speed differentiation between front output shaft 258 and rear output shaft 256. While the arrows in FIG. 4 illustrates the power flow path through transmission 20 for establishing the full-time high-range four-wheel drive mode and FIG. 5 illustrates the power flow path establishing the full-time low-range four-wheel drive mode, each of these modes may be selectively changed from the illustrated full-time four-wheel mode to a corresponding part-time four-wheel mode. This change is made by actuating lock-out mechanism 70 to move lock-out collar 300 from the Open ("O") position to its Locked ("L") position for locking-out differential 68. In this differential locked-out condition, part-time four-wheel drive is established because interaxle differentiation is prohibited. Thus, in both high-range and low-range the four-wheel drive mode may be selected to be either full-time or part-time. Moreover, while lock-out mechanism 70 is illustrated as a dog clutch type arrangement, it is contemplated that a synchronizing-type clutch assembly could be used for facilitating shifting from full-time to part-time four-wheel operation "on-the-fly" without stopping the vehicle. Alternatively, lock-out mechanism 70 could be replaced with a slip limiting device such as, for example, a viscous coupling for automatically and progressively inhibiting differentiation across differential 68 when such interaxle speed differentiation becomes excessive. As a further alternative, an electrically-controlled friction clutch can be used to apply a braking torque between the two outputs of interaxle differential 68 for limiting slip therebetween. In such a system, the braking torque could be controlled in various manners (i.e., ON/OFF, modulated, etc.) in response to the magnitude of the interaxle slip. As an additional feature, transmission 20 may also be shifted "on-the-fly" between the four-wheel drive (full-time or part-time) modes and two-wheel drive modes. To this end, mode select mechanism 72 provides a means for selectively coupling and uncoupling front output shaft 258 to front prop shaft 34. In particular, mode select mechanism 72 includes a first hub 322 fixed for rotation with front output shaft 258, a second hub 324 fixed for rotation with front prop shaft 34, and a synchronizer clutch assembly 326 therebetween. Synchronizer clutch assembly 326 is a dual-cone type and includes a mode collar 328 having internal splines 330 retained for rotation with and axial sliding movement on external splines 332 formed on first hub 322. A blocker ring 334 and a friction cone 336 are interposed between mode clutch collar 328 and second hub 324 in a well known manner such that movement of mode collar 328 from its 2WD position (FIG. 7) toward its 4WD position (FIGS. 4, 5 and 6) causes speed synchronization to occur between front output shaft 258 and front prop shaft 34. Once synchronization is complete, splines 330 on mode collar 328 meshingly engage clutch teeth 338 on second hub 324, thereby directly coupling front output shaft 258 to front prop shaft 34. Movement of mode collar 328 is effectuated by a mode fork, partially shown at 340, which is likewise connected to the shift apparatus. When operation of transmission 20 in the two-wheel drive mode (either high-range or low-range) is desired, differential lock-out mechanism 70 is actuated in coordination with movement of mode collar 328 to its 2WD position to cause lock-out collar 300 to move to its Locked ("L") position. Since geartrain 42 of transmission 20 is shown as a manually-operated five-speed arrangement, it will be clear that any suitable gearshift mechanism can be utilized for controlling the coordinated movement of shift forks 190, 210 and 240 so as to permit the vehicle operator to select a desired speed ratio. In addition, it is contemplated that a movement coordinating mechanism, such as the above-noted sector plate, can be utilized to control the coordinated movement of range collar 268, mode collar 328, and lock-out collar 300 for permitting the vehicle operator to select a desired drive mode. As set forth in the table shown at FIG. 8, transmission 20 is capable of permitting the vehicle operator to select any one of a two-wheel high-range drive mode, a two-wheel low-range drive mode, a full-time high-range four-wheel drive mode, a full-time low-range four-wheel drive mode, a part-time high-range four-wheel drive mode, a part-time low-range four-wheel drive mode, and a neutral mode. The shift apparatus associated with power transfer mechanism 44 can be a manually-operated mechanical system controlled by a gearshift lever or can be a electronically-controlled system of actuators arranged to move the various shift collars. While transmission 20 has been described above in its preferred embodiment as integrating both a planetary-type interaxle differential 68 and mode select mechanism 72 for disengaging front output shaft 258 from front driveline 12, it must be understood that either one of these features or neither one of these features may be included in transmission 20.

Claims: A transmission for use in a four-wheel drive vehicle having an engine (20) and front and rear drivelines (34, 40), said transmission comprising: a housing (118); a first shaft (46) supported for rotation in said housing and driven by the engine; a second shaft (48) supported for rotation in said housing; a plurality of constant-mesh gearsets (50, 52, 54, 56, 58) arranged for selectively coupling said second shaft to said first shaft for driven rotation at various speed ratios; a third shaft (60) supported for rotation in said housing; first gear means (64) supported for rotation on said third shaft and rotatably driven by said second shaft at a direct speed ratio relative thereto; second gear means (66) supported for rotation on said third shaft and rotatably driven by said second shaft at a reduced speed ratio relative thereto; clutch means (44) for selectively coupling one of said first and second gear means to said third shaft for causing driven rotation thereof at one of said direct and reduced speed ratios; and an interaxle differential (68) mounted in said housing and having an input member (283) rotatably driven by said third shaft, a first output member (282) coupled to the front driveline, and a second output (284) member interconnected to the rear driveline, said interaxle differential operable for permitting speed differentiation between said first and second output members while delivering drive torque to the front and rear drivelines from said third shaft. The transmission of Claim 1 wherein said clutch means (44) is movable between a high-range position coupling said clutch means to said first gear means (64) for coupling said third shaft to said second shaft for rotation at said direct speed ratio such that a full-time high-range four-wheel drive mode is established, and a low-range position coupling said clutch means to said second gear means (66) for coupling said third shaft to said second shaft for rotation at said reduced speed ratio such that a full-time low-range four-wheel drive mode is established. The transmission of Claim 2 wherein said clutch means is a clutch assembly (44) having a hub (262) fixed for rotation with said third shaft (60) and a range collar (268) supported for rotation with an axial sliding motion on said hub, whereby movement of said range collar in a first direction (fig.4) to said high-range position couples said first gear means to said third shaft for establishing a high-range power transmission route from said second shaft to said third shaft, and movement of said range collar in an opposite direction (fig 5) to said low-range position couples said second gear means to said third shaft for establishing a low-range power transmission route from said second shaft to said third shaft. The transmission of Claim 2 further comprising a lock-out mechanism (70) operable in an open mode (fig 4, 5) for permitting speed differentiation between said first and second output member such that when said clutch means is in said high-range position said full-time high-range four-wheel drive mode is established and when said clutch means is in said low-range position said full-time low-range four-wheel drive mode is established, said lock-out mechanism also operable in a locked mode (fig 6) for preventing speed differentiation between said first and second output members such that when said clutch means is in said high-range position a part-time high-range four-wheel drive mode is established and when said clutch means is in said low-range position a part-time low-range four-wheel drive mode is established. The transmission of Claim 4 wherein said lock-out mechanism (70) includes second clutch means (300) movable between a locked position directly coupling said input member (302) to one of said output members (290) for shifting said interaxle differential into said locked mode and an open position de-coupling said input member from said one output member for permitting differential action across said interaxle differential. The transmission of Claim 1 wherein said differential is a planetary-type gear assembly (68) including a carrier (283) rotatably driven by said third shaft (60) and having a plurality of pinion gears (286) supported thereon, wherein said first output member is a sun gear (282) fixed to a first output shaft (258/32) that is coupled to one of the front and rear drivelines, and wherein said second output member is a ring gear (284) coupled to a second output shaft (38) which is interconnected to the other of the front and rear drivelines, said pinion gears being in meshed engagement with said sun gear and said ring gear. The transmission of Claim 6 further comprising a lock-out mechanism (70) operable in an open mode for permitting speed differentiation between said sun gear and ring gear, and in a locked mode for coupling said carrier (283) for common rotation with one of said sun gear (282) and said ring gear (284) for preventing speed differentiation between said front and rear output shafts. The transmission of Claim 1 further comprising a mode shift mechanism (72) operable in a 2WD mode (fig.7) for selectively uncoupling one of said first and output members from its corresponding driveline for establishing a two-wheel drive mode, and said mode shift mechanism operable in a 4WD mode for coupling said output member to its corresponding driveline for establishing a four-wheel drive mode. The transmission of Claim 8 wherein said mode shift mechanism (72) is a synchronized clutch assembly having means for synchronizing the speed of said output member (32) relative to its corresponding driveline (34) prior to coupling said output member thereto, thereby permitting said transmission to be shifted on-the-fly from said two-wheel drive mode to said four-wheel drive mode.

Inventors: FORSYTH JOHN R, FORSYTH, JOHN R.
Applicants: NEW VENTURE GEAR INC, NEW VENTURE GEAR, INC.
Application year: 1996
Main classification: B60K17
IPC codes: B60K17, B60K23