Abstract:
A multiple speed transmission capable of producing five underdrive gear ratios, one direct drive ratio, four overdrive ratios, and four reverse ratios includes four interconnected planetary gearsets, multi-plate clutches and brakes controlling the gearset elements, and a coupler that changes between forward drive and reverse drive by holding and releasing alternate gearset components against rotation on the transmission case. The transmission in combination with a single-speed transfer case having either an on-demand transfer clutch, or a full-time all-wheel drive system, are formed in an integrated package.

Description:
BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   This invention relates to geared, automatic power transmissions for motor vehicles. More particularly, it pertains to the kinematic arrangement for such transmissions and a driveline that incorporates them. 
   2. Description of the Prior Art 
   A transfer case contains a drive mechanism for connecting an engine to the vehicle&#39;s wheels, a primary wheel set that is connected continually to the engine, and a secondary wheel set that is connected selectively to the engine through the transfer case. A two-speed transfer case produces both a high range and a low range of speed ratios. Each of the speed ratios produced by the transmission is operative in each of the drive ranges produced by the transfer case; therefore, the transmission-transfer case can produce in combination a multiple of the transmission gear ratios. In this way the transfer case provides greater functional flexibility to the vehicle operator, who has a wider range of speed ratios to select depending on road conditions, towing load, vehicle speed, and other drive conditions. 
   Most operators of 4×4 vehicles, those vehicles having a driveline that transmits engine power to front and rear wheels, rarely operate the driveline in the Low Range, the range in which the lowest speed ratios are produced. A principal reason for this is the need to stop the vehicle before changing the drive state of a transfer case clutch that must be engaged in order to produce the Low Range. The wider range provided by new six speed automatic transmissions will likely cause still less frequent use of the Low Range, and decrease the need for two-speed transfer cases to meet the requirements of vehicle operators. 
   Certain vehicle lines require an overall transmission/transfer case gear ratio of approximately 12:1 in combination with relatively low engine speed, i.e. 2–3 mph @1000 rpm. Other vehicle lines have much less much less stringent requirements, i.e., lower gear ratios at low engine speed. 
   Automatic shifting under full torque from the low speed range to high speed range is a goal in driveline design, but the speed ratio step, from the low range to high range, in current transfer cases is too large for good shift quality and would need to be reduced to approximately 1.8 or less. However, performance degradation due to the resulting lower gear ratios would be undesirable. The use of multi-plate clutches in the transfer case to shift from low to high, coupled with six-speed automatic transmissions would result in added parasitic drag and lower fuel economy. In certain applications where a 12:1 overall transmission/transfer case speed ratio is required, a three-speed transfer case would be necessary with currently available six-speed automatic transmissions. But this would compound the losses. 
   SUMMARY OF THE INVENTION 
   A vehicle driveline having a transmission that produces eight to ten speed ratios in combination with a single speed transfer case according to the present invention is used to provide the extended speed ratio range needed to meet the diverse requirements of a various vehicle lines, and the ability to shift from the low range to high range under full torque. The combination ensures good shift quality and robust park system function. 
   A selector switch in a vehicle having a driveline according to the present invention might have 4×2, 4×4 High and 4×4 Low modes, but it could also have added functions like 4×4 Auto, Trailer Tow, and Snow modes. These modes are achieved by starting the transmission in first, second or third gear and by using a 4×4 on-demand clutch in Off, Stand-by and Applied modes. More than one reverse drive speed ratio would be available for each of these modes. 
   A coupler, such as a synchronizer is used to change between forward drive and reverse drive. The synchronizer improves fuel efficiency by eliminating two high-capacity, hydraulic friction clutches and their associated drag losses. 
   A transmission according to this invention is intended for use in four-wheel drive applications in combination with a single speed transfer case having either an on-demand transfer clutch, or a full time all-wheel-drive system. The transmission and transfer case can be packaged in a compact space no larger than required to accommodate a five-speed transmission and two-speed transfer case. 
   Although ten forward speed ratios and four reverse speed ratios are available, in normal, on-road use the transmission would function as a six-speed with one reverse speed. It is envisioned that a vehicle using this transmission would have several useful modes of operation, such as “4×2”, “Auto”, “4×4 Hi”, “4×4 Low”, “Snow” and “Trailer Tow”. These modes would essentially select the first gear starting ratio, reverse ratio, and the state of the 4×4 on demand clutch in the transfer case. 
   A multiple-speed ratio automatic transmission for an automotive vehicle includes an input shaft; output shaft; a planetary gear system comprising first, second, third, and fourth planetary gear units, each gear unit having a sun gear, a ring gear, planet pinions meshing with the sun gear and ring gear, and a carrier rotatably supporting the planet pinions, the input shaft being connected driveably to the sun gear of the first gear unit, the output shaft being connected driveably to the carrier of the third gear unit, the sun gear of the third gear unit and the sun gear of the fourth gear unit being mutually driveably connected, the carrier of the first gear unit and ring gear of the second gear unit being mutually driveably connected, the carrier of the second gear unit and ring gear of the fourth gear unit being mutually driveably connected, the ring gear of the third gear unit and carrier of the fourth gear unit being mutually driveably connected; a coupler for alternately holding against rotation and releasing the carrier of the fourth gear unit and the sun gear of the fourth gear unit; a first brake for holding against rotation and releasing the sun gear of the second gear unit; a first clutch for driveably connecting and disconnecting the ring gear of the first gear unit and carrier of the second gear unit; a second clutch for driveably connecting and disconnecting the ring gear of the first gear unit and input shaft; a third clutch for driveably connecting and disconnecting the sun gear of the fourth gear unit and the coupler; and a fourth clutch for driveably connecting and disconnecting the sun gear of the first gear unit and ring gear of the third gear unit. 
   Various objects and advantages of this invention will become apparent to those skilled in the art from the following detailed description of the preferred embodiment, when read in light of the accompanying drawings. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is a schematic diagram of the kinematic arrangement of a transmission according to the present invention; 
       FIG. 2  is a chart that shows the pattern of engagement and release of the clutches, brakes, and couplings required to establish the forward drive and reverse drive ratios of the transmission mechanism of  FIG. 1 ; 
       FIG. 3  is a schematic diagram of the kinematic arrangement of an alternate embodiment of the transmission according to this invention; and 
       FIG. 4  is a chart that shows the pattern of engagement and release of the clutches, brakes, and couplings required to establish the forward drive and reverse drive ratios of the transmission mechanism of  FIG. 3 . 
   

   DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     FIG. 1  shows a schematic diagram of a kinematic arrangement according to the present invention capable of producing ten forward speed ratios—five underdrive gear ratios, one direct drive ratio, and four overdrive ratios—and four reverse drive speed ratios.  FIG. 2  shows the states of engagement and disengagement of various clutches and brakes corresponding to each speed ratio. The transmission arrangement of  FIG. 1  includes four planetary gear sets, six multi-plate clutches and brakes that control the gearset elements, and a synchronizer for producing forward drive and reverse drive. 
   A hydrokinetic torque converter  10  includes an impeller  12  connected to the crankshaft  14  of an internal combustion engine, a bladed turbine  16 , and a bladed stator  18 . The impeller and turbine wheels define a toroidal fluid flow circuit, whereby the impeller is hydrokinetically connected to the turbine. The stator  18  is supported rotatably on a stationary stator sleeve shaft  20 , and an overrunning brake  22  anchors the stator to the shaft  20  to prevent rotation of the stator in a direction opposite the direction of rotation of the impeller, although free-wheeling motion in the opposite direction is permitted. 
   The torque converter assembly includes a lockup clutch  24  located within the torque converter impeller housing  25 . The torque output side of lockup clutch  24  includes a damper  26  located between the impeller and the turbine shaft, which is the transmission input shaft  28 . When clutch  24  is engaged, the turbine and impeller are mechanically connected; when clutch  24  is disengaged, they are hydrokinetically connected and mechanically disconnected. The damper absorbs transitory torque fluctuations associated with engagement of a lockup clutch. Fluid to the torque converter is supplied from the output of an oil pump assembly  29 . 
   Planetary gearing includes first, second, third, and fourth planetary gear units  30 ,  32 ,  34 , and  36 . The first gear unit  30  includes a sun gear  38 , ring gear  40 , carrier  42 , and planetary pinions  44 , supported by carrier  42  and in meshing engagement with sun gear  38  and ring gear  40 . The second gear unit  32  includes a sun gear  46 , ring gear  48 , carrier  50 , and planetary pinions  52 , rotatably supported on carrier  50  and in meshing engagement with sun gear  46  and ring gear  48 . A third gear unit  34  includes a sun gear  54 , ring gear  56 , carrier  58 , and planetary pinions  60 , rotatably supported on carrier  58  and in meshing engagement with sun gear  54  and ring gear  46 . A fourth gear unit  36  includes a sun gear  62 , ring gear  64 , carrier  66 , and planetary pinions  68 , rotatably supported on carrier  66  and in meshing engagement with sun gear  62  and ring gear  64 . 
   A coupler  70 , which may be a dog clutch, but preferably is a synchronizer of the type used in a manual transmission for automotive use, includes a sleeve  72 , supported on a hub  74  for axial sliding movement leftward and rightward. The hub is secured to the transmission case  75  against rotation. Preferably, sleeve  72  has a set of spline teeth formed on its inner surface, and hub  74  has a set of spline teeth on its outer surface continually engaged with those of the sleeve  72 . Similarly, the spine teeth of sleeve  72  are aligned and engageable mutually with spline teeth on the outer surface of a disc  76 , which is driveably connected to carrier  66 . The spine teeth of sleeve  72  are aligned and engageable mutually with spline teeth on the outer surface of a second disc  78 . When sleeve  72  is moved rightward to produce a drive connection between the case  75  and disc  78 , the coupling  70  is in the forward drive state. When sleeve  72  is moved leftward to produce a drive connection between case  75  and disc  76 , the coupling  70  is in the reverse drive state. 
   Carrier  50  is continually driveably connected by member  80  to ring gear  64 ; carrier  58  is continually driveably connected to transmission output  82 ; carrier  66  driveably connects ring gear  56  and disc  76 ; shaft  84  driveably connects sun gears  54  and  62 ; and carrier  42  is driveably connected to ring gear  48 . 
   Hydraulically actuated clutches and brakes produce selected, releasable drive connections among the components of the kinematic arrangement. A brake  90  alternately connects and releases sun gear  46  and case  75 ; a clutch  92  alternately connects and releases carrier  50  and a drum  94  connected to ring gear  40 ; a clutch  96  alternately connects and releases ring gear  40  and transmission input  28 ; a brake  98  alternately connects and releases drum  94  and case  75 ; a clutch  100  alternately connects and releases disc  78  and shaft  102  which is connected to sun gears  54  and  62 ; and a clutch  104  alternately connects and releases ring gear  56  and sun gear  38  through intermediate shaft  106 . 
   During operation in first gear, a one-way coupling  108  produces a one-way drive connection between case  75  and drum  94 . During positive torque conditions, when power is transmitted from the engine through the transmission to the wheels, a one-way coupling  110  produces a one-way drive connection between coupler  70  and sun gear  62 , and it overruns during negative torque conditions when power is transmitted from the wheels through the transmission to the engine. A principal purpose of clutch  100  is to produce, during negative torque conditions, the drive connection between coupler  70  and sun gear  62  that coupling  110  provides during positive torque conditions. 
   A gear selector lever, called a “PRNDL” controlled by the vehicle operator, is used to select the operating range of the transmission by moving the selector manually among positions where the various gear ranges are produced automatically and other positions where the gear ratios are produced manually. A park gear  112 , carried on the output  82 , is releasably fixed to the case  75  when the gear selector is moved to the “P” or Park position. This holds the shaft  82  against rotation, fixes the wheels to the case  75 , and prevents the vehicle from moving. 
   Two low-pressure hydraulic pumps  114 ,  116  are mounted on the output shaft  82 . Pump  114  has an inlet connected to the transmission oil sump, and an outlet, through which lubrication fluid is supplied to the transfer case  120 . Pump  116  has an inlet in the oil sump of transfer case  120 , and an outlet, which supplies oil to the transmission lubrication circuit. This lubrication source allows the vehicle to be “flat” towed, i.e., to be towed with all its wheels rotating in contact with the ground. The pump  116  keeps the transfer case sump dry, thereby reducing drag losses. 
   A transfer case  120  has a input  122 , a primary output  124 , a secondary output  126 , a center planetary differential  128 , a bias clutch  132 , and a chain drive mechanism  133  for transferring power from an sun gear  134  of differential  128  to the output  126 . The center differential  128  includes a carrier  136 , a ring gear  138  connected to output  124 , a sun gear  134  connected to a sprocket wheel  140  carried on shaft  142 , and a set of planet pinions  144  supported rotatably on the carrier  136  and in meshing engagement with the ring gear  138  and sun gear  134 . The chain drive mechanism  133  includes another sprocket wheel  146  supported on output shaft  126 , and a drive chain  148  engaged with sprocket wheels  140  and  146 . 
   In accordance with the relative size of the gears of the differential  128  and the magnitude of slip across clutch  132 , the center differential  128  divides power carried on output shaft  82  between the primary transfer case output  124  and the secondary output  126 . Slip across clutch  132  is a function of the extent to which that clutch is fully engaged, which is controlled by the magnitude of hydraulic pressure applied to a servo that actuates clutch  132 . 
     FIG. 2  is a chart indicating the state of the clutches and brakes that are engaged and released selectively to produce each of the speed ratios. In  FIG. 2 , symbol “X” identifies an engaged friction clutch or friction brake, or a driving one-way coupling, a one-way clutch, or one-way brake. The symbol “O/R” indicates an overrunning condition for a one-way clutch or brake. The symbol “EB” indicates that the corresponding friction element is engaged during a coasting condition. The symbols “F” and “R” indicate the forward and reverse states of coupling  70 . A blank indicates that the corresponding clutch or brake is disengaged or released. 
   First gear is achieved by moving sleeve  72  of coupler  70  rightward into engagement with disc  78 , the forward drive position, and by applying brake  90 . With the control elements so located and applied, one-way clutch  110  and one-way brake  108  are also engaged and driving during a positive torque condition, i.e., when power is transmitted from the input  28  through the transmission to the output  82 . This first gear or lowest speed ratio would be used in the 4×4 “Low” range for the most severe off-road maneuvers, such as slowly surmounting obstacles, or ascending and descending steep grades. 
   An upshift to second gear is a non-synchronous shift event. Clutch  92  is engaged, brake  90  remains engaged, coupling  110  drives, but coupling  108  overruns. Second gear is roughly equivalent to the Low Range produced by most of the current transfer case systems. It would be used in 4×4 “Low” operation or as first gear ratio when trailer tow mode is selected. 
   Upshifts from second gear to third gear, and from third gear to fourth gear are synchronous shifts. The 2–3 upshift results by disengaging clutch  92 , engaging clutch  96  and maintaining brake  90  engaged. Coupling  108  overruns, but coupling  110  can provide overrun capability for coasting downshifts and parking lot, low speed driving. The third speed ratio is the start-up speed ratio most often used in normal driving conditions. The 3–4 upshift results by disengaging brake  90 , re-engaging clutch  92  and maintaining clutch  96  engaged. 
   The fourth gear to fifth gear upshift is a non-synchronous shift produced by disengaging clutch  96  and engaging clutch  104 . The fifth gear to sixth gear upshift is a non-synchronous shift, clutch  96  is engaged, and coupling  110  overruns. All upshifts above sixth gear are synchronous. Fifth gear is an underdrive gear ratio, sixth gear is a direct drive 1:1 gear ratio, and seventh gear through tenth gear are the overdrive gear ratios. The step size between the gear ratios from sixth gear through tenth gear is small. Therefore, the six gears that would be used for most normal drive conditions include third, fourth, fifth, sixth, seventh and tenth, whose gear ratios preferably are 3 rd =3.54:1, 4 th =2.28:1, 5 th =1.46:1, 6 th =1.00:1, 7 th =0.83:1, and 10 th 
=0.63:1. 
   Reverse drive requires leftward movement of the sleeve  72  of coupler  70  to the reverse drive position where disc  76  is driveably connected to the case  75 . Four reverse speed ratios can be produced using the same clutch and brake engagement and disengagement states as are used for the four lowest forward drive gears. The third reverse gear, whose gear ratio is preferably 2.73:1, would be used for normal 4×2 operation. The first reverse gear and second reverse gear would be used for off-road and trailer towing conditions. 
   The kinematic arrangement shown in  FIG. 3  produces seven forward speed ratios and avoids the lowest speed ratio produced by the arrangement of FIG.  1 . The kinematic arrangement of  FIG. 3  deletes brake  98  and coupling  108  from the arrangement of  FIG. 1 . It produces a first gear ratio that is approximately the same as the second gear ratio of the  FIG. 1  arrangement, and it produces only three reverse gear ratios in the range 4.99:1–1.76:1. In a preferred embodiment, the overdrive ratios in sixth gear and seventh gear are 0.83:1 and 0.73:1, respectively.  FIG. 4  is a chart indicating the engaged and released states of the clutches and brakes that selectively produce each of the speed ratios. 
   The transfer case  120 ′ has a primary output  124 , driveably connected to transmission output  82 , a secondary output  126 , a clutch  150 , and a chain drive mechanism  133  for transferring power from output  124  to output  126 . The chain drive mechanism  133  includes a first sprocket wheel  140 , a second sprocket wheel  146  supported on output shaft  126 , and a drive chain  148  engaged with sprocket wheels  140  and  146 . Clutch  150  divides power carried on output shaft  82  between the primary transfer case output  124  and the secondary output  126 . Slip across clutch  150  is a function of the extent to which that clutch is fully engaged, which is controlled by the magnitude of hydraulic pressure applied to a servo that actuates clutch  150 . 
   In accordance with the provisions of the patent statutes, the principle and mode of operation of this invention have been explained and illustrated in its preferred embodiment. However, it must be understood that this invention may be practiced otherwise than as specifically explained and illustrated without departing from its spirit or scope.