Abstract:
A hydromechanical transmission includes a hydrostatic transmission driven by an engine and drivingly connected to a planetary gear set. A plurality of clutches are associatable with the ring gear or elements of the planetary gear set for establishing a corresponding plurality of operating modes in which additional overall transmission speed can be achieved while reducing the power transmitted by the hydrostatic transmission. The gears of the planetary gear set can be removed and replaced with gears having a different number of teeth so as to change the overall ratio of the hydromechanical transmission without changing the power rating of the hydrostatic transmission or the size of the housing. Thus, multiple overall corner horsepower capabilities can be provided with the same hydrostatic transmission and housing package size.

Description:
BACKGROUND OF THE INVENTION 
     The present invention relates to the field of propulsion systems for vehicles. More particularly, this invention relates to a hybrid hydromechanical vehicle transmission that includes both a hydraulic transmission and a planetary gear set. The hybrid transmission has multiple modes of operation and provides greater control of corner horsepower. The hybrid hydromechanical transmission of this invention is flexible in maximum output torque, maximum output speed, and maximum corner power without the need to alter the physical layout of the gears or change the size of hydrostatic transmission. 
     When similar vehicles are produced with different engine power levels, they frequently have requirements for a different “corner horsepower” value. The corner power is defined by the product of the maximum output torque and maximum output speed. See the dashed lines in FIG.  10 . The engine power and the required vehicle corner power normally vary somewhat in proportion, as limitations of tractive effort or vehicle speed also vary with engine power. When using gear transmissions, this may be accommodated without changing the transmission ratios as vehicle corner power is determined by the product of engine power and overall gear transmission ratio. 
     When using a hydraulic transmission, the vehicle corner power is defined by the product of maximum hydraulic unit torque and speed, and is usually not affected by engine power. Hydromechanical transmissions in general are discussed in U.S. Pat. No. 4,341,131 and 4,306,467. The complete disclosure of those patents is incorporated by reference herein. 
     There is a need for a hydromechanical transmission which provides greater flexibility and represents an improvement over the prior art. Thus, a primary objective of the present invention is the provision of an improved hydromechanical transmission. 
     A further objective of this invention is the provision of a hydromechanical transmission that uses the same hydraulic units and the same gear layout to achieve different hydromechanical corner horsepower outputs. 
     A further objective of this invention is the provision of a hydromechanical transmission that can provide at least two substantially different levels of overall corner horsepower while maintaining a constant package size for installation in a vehicle. 
     A further objective of this invention is the provision of a hydraulic unit and gear configuration that is flexible in maximum output torque, maximum output speed, and maximum corner power. 
     A further objective of this invention is the provision of a hydromechanical transmission that has a plurality of clutches that are used to establish a corresponding plurality of modes. 
     A further objective of this invention is the provision of a hydromechanical transmission with a compound gear set having a rotatable carrier plate assembly that has two powered output shafts extending therefrom for front and rear output. 
     A further objective of this invention is the provision of a hydromechanical transmission that has two centerlines and front and rear outputs. 
     A further objective of this invention is the provision of a hydromechanical transmission that has a space efficient layout and a simple, cost-effective gear design. 
     A further objective of this invention is the provision of a method of changing the overall corner horsepower of a hydromechanical transmission without changing the power rating of its hydrostatic transmission or the size of the planetary gear set housing. 
     These and other objectives will be apparent to one skilled in the art from the drawings, as well as from the description and claims that follow. 
     SUMMARY OF THE INVENTION 
     An improved hydromechanical transmission includes a hydrostatic transmission driven by an engine and drivingly connected to a compound planetary gear set. The hydrostatic transmission includes a variable displacement unit connected in a closed loop circuit to a fixed displacement unit. The planetary gear set includes a ring gear rotatably mounted in a housing, a carrier plate assembly with output shafts protruding therefrom, a plurality of planet gears rotatably mounted on the carrier plate assembly so that at least some engage the ring gear, and sun gears meshed with the planet gears. 
     A first drive gear is drivingly attached to the input shaft of the variable unit and a second drive gear is drivingly attached to the output shaft of the fixed unit. A plurality of clutches are associated with elements of the planetary gear set to establish a plurality of operating modes in which additional overall transmission speed can be achieved while reducing the power consumed by the hydrostatic transmission. 
     Two-mode and three-mode configurations of the hydromechanical transmission of this invention are disclosed below, as well as their operation and the method of switching between them. The gears of the planetary gear set can be removed and replaced with gears having a different number of teeth so as to change the overall ratio of the hydromechanical transmission without changing the power rating of the hydrostatic transmission or the size of the housing. Thus, multiple overall corner horsepower capabilities can be provided with the same hydrostatic transmission and housing package size. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a schematic representation of the two-mode configuration of the hydromechanical transmission of the present invention in its first mode. 
     FIG. 2 is schematic representation of the two-mode configuration of the hydromechanical transmission of the present invention in its second mode. 
     FIG. 3 is a schematic representation of the three-mode configuration of the hydromechanical transmission of this invention in its first mode. 
     FIG. 4 is a schematic representation of the three-mode configuration of the hydromechanical transmission of this invention in its second mode. 
     FIG. 5 is a schematic representation of the three-mode configuration of the hydromechanical transmission of this invention in its third mode. 
     FIG. 6 is a schematic representation showing the compound planetary gear set layout of this invention in greater detail. 
     FIG. 7 is a schematic representation focusing on the planetary gear set layout of this invention. 
     FIG. 8 is a graph of overall transmission output speed versus hydrostatic power and fixed unit speed for the two-mode configuration of this invention. 
     FIG. 9 is a graph similar to FIG. 8 but illustrates the three-mode configuration. 
     FIG. 10 is a graph illustrating output torque versus output speed. 
     FIG. 11 is a diagram illustrating how the invention can be used to supply power to an optional second axle for a 4-wheel drive. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     In the figures and the description that follows, the hydromechanical transmission of this invention can be configured in at least two different ways, generally designated by the reference numerals  10  and  10 A respectively. In the two-mode configuration of the invention shown in FIGS. 1 and 2, the hydromechanical transmission  10  of this invention includes a housing  12  in which a compound planetary gear set  14  is mounted. The hydraulic portion of the transmission  10  includes a hydrostatic transmission  16 . The hydrostatic transmission (HST)  16  includes a variable displacement hydraulic unit  17 V (pump) fluidly connected in a closed loop circuit to a hydraulic unit  17 F. Preferably hydraulic unit  17 F is a fixed displacement unit sometimes referred to as a motor. The variable unit  17 V has an input shaft  18  connected to a source of rotational power, such as a conventional engine (not shown). The input shaft  18  drivingly connects to a first drive gear  20 . The fixed displacement  17 F has an output shaft  22  that drivingly connects to a second drive gear  24 . The input shaft  18  and the output shaft  22  may share the same centerline, although other arrangements are also possible without significantly detracting from the invention. The hydrostatic unit  16  and the planetary gear set  14  can also be mounted in the same housing  12  yet have different centerlines. 
     FIGS. 6 and 7 show the planetary gear set  14  in greater detail. As best seen in FIG. 7, the planetary gear set  14  has a ring gear  26  and carrier plate assembly  28  rotatably mounted in the housing  12 . The carrier plate assembly  28  has oppositely directed, centrally located power output shafts  30 ,  32 . Together, the output shafts  30 ,  32  define an axis of rotation  33 . The axis of rotation of the gear set  14  is offset from the centerline of the input shaft  18  and output shaft  22  of the hydraulic units. As is conventional, the carrier plate assembly  28  includes opposing plates that have outwardly directed opposing surfaces from which the output shafts  30 ,  32  extend. 
     A plurality of radially and angularly spaced pins  34  extend from one of the outwardly directed opposing surfaces. Each of the pins  34  is generally parallel to the axis of rotation  33 . A plurality of planetary gears  36 A,  36 B,  36 C rotatably mount on the pins  34 . As is conventional, the planetary gears  36 A,  36 B,  36 C are generally cylindrical and have an outer diameter with a plurality of gear teeth thereon. Together the gears  36 A,  36 B,  36 C engage and support a sun gear  38  between them. The sun gear  38  is rotatably mounted on the output shaft  32 . 
     Another plurality of radially and angularly spaced pins  40  extend from the other of the outwardly directed opposing surfaces. The pins  40  are generally parallel to the axis of rotation  33 . A second plurality of planetary gears  42 A,  42 B,  42 C rotatably mount on pins  40 . The planetary gears  42 A,  42 B,  42 C are generally cylindrical and have an outer diameter with a plurality of gear teeth thereon. The gear teeth on each gear  42 A,  42 B,  42 C simultaneously engage the gear teeth on an adjacent respective planetary gear  36 A,  36 B,  36 C and the gear teeth on the inside diameter of the ring gear  26 . Furthermore, the gear teeth on the planetary gears  42 A,  42 B,  42 C mesh with a sun gear  44  that is supported between them. The sun gear  44  slidably mounts on the output shaft  30 . FIG. 6 illustrates the meshing of the gears of the planetary gear set. As is known in the art of planetary gear sets, the carrier plate assembly  28  has voids and clearance holes where needed to allow for the corresponding gears  36 A,  36 B,  36 C and  42 A,  42 B,  42 C to mesh respectively with each other and to reduce the weight of the assembly. The dashed lines connecting gears  36 A,  36 B,  36 C and  42 A,  42 B,  42 C in FIG. 7 indicate that these gears mesh with each other. 
     Two clutches are operatively associated with the planetary gear set  14  in the two-mode transmission configuration shown in FIGS. 1 and 2. A first clutch  46  selectively connects the ring gear  26  to the housing  12 , thereby fixing the ring gear  26  to the housing  12  and preventing the ring gear  26  from rotating relative to the housing  12 . Thus, the first clutch  46  acts as a brake for the ring gear  26 . 
     A second clutch  48  selectively connects the sun gear  44  to the first drive gear  20  through a first intermediate gear  50 . A similar second intermediate gear  52 , preferably identical in terms of number of teeth and pitch diameter, interconnects the second drive gear  24  with the sun gear  38 . When the second clutch  48  is engaged, the sun gear  44  is rotated at the speed of gear  50 . Ideally the clutches  46 ,  48  are synchronous clutches or at least near synchronous clutches. Thus clutches  46 ,  48  can be engaged or disengaged without changing the output speed ratio. Preferably, the clutches  46 ,  48  are disposed on the axis of rotation  33 . 
     In operation, the two-mode configuration of this invention allows for selection between two power paths, as best seen in view of FIGS. 1 and 2. In FIG. 1, the clutch  46  is engaged by the operator and the ring gear  26  is prevented from rotating relative to the housing  12 . The engagement of the clutch  46  is indicated by the dark angled line drawn through the clutch. Power in mode  1  is transmitted from the engine to the variable unit  17 V, which converts the rotational energy to fluid energy and thereby causes the fixed hydraulic unit  17 F to rotate its output shaft  22 . The second drive gear  24  attached to the output shaft  22  rotates, driving the sun gear  38  through the intermediate gear  52 . The sun gear  38  rotates the first planet gears  36 A,  36 B,  36 C, which act as idlers that counter-rotate to drive the second planet gears  42 A,  42 B,  42 C in the same direction of rotation as sun gear  38 . Because the ring gear  26  cannot rotate in response to the rotation of the second planet gears  42 A,  42 B,  42 C, the carrier plate assembly  28  rotates instead. This causes the power output shafts  30 ,  32  to rotate. 
     The two rotating power output shafts  30 ,  32  can be used for two-wheel drive of a vehicle or can be used for front and rear axle drives in a four-wheel vehicle. See FIG.  11 . As shown in the graph of FIG. 8, mode  1  is available in both forward and reverse directions of vehicle travel. The variable displacement pump  17 V is merely stroked in an opposite direction so that fluid is pumped in a different direction around the closed loop circuit. 
     FIG. 2 shows the power path for mode  2 , in which the first clutch  46  is disengaged and the second clutch  48  is engaged. The engine rotates the input shaft  18  of the variable unit  17 V and the attached first drive gear  20 . The first drive gear  20  rotates the first intermediate gear  50 , which is now drivingly connected to the second sun gear  44  by the clutch  48 . The sun gear  44  then rotates at a speed that is proportional to the engine speed. 
     Meanwhile, the hydrostatic transmission  16  is also driven by the engine and rotates the second drive gear  24  at a given speed that depends on the commanded displacement of the variable unit  17 V. The intermediate gear  52  rotates the sun gear  38 , which drives the planet gears  36 A,  36 B,  36 C. Because the planet gears  36 A,  36 B,  36 C are meshed with planet gears  42 A,  42 B,  42 C, as is the ring gear  26 . The carrier plate assembly  28  is continuously driven at a speed that is dependent on the displacement of the variable unit  17 V. 
     Thus, the carrier assembly and the output shafts  30 ,  32  attached thereto rotate at a speed that is a function of both the input speed and displacement ratio of the hydrostatic transmission. The operator can increase the speed of the overall hydromechanical transmission  10  at the power output shafts  30 ,  32  without consuming additional hydrostatic transmission power by shifting the hydromechanical transmission  10  from mode  1  to mode  2 . See the graph of FIG.  8 . At low output speeds, the fixed unit  17 F actually functions as a pump and the variable unit  17 V acts as a motor. Thus, the hydrostatic (HST) power dips into the negative area of FIG.  8 . At higher output speeds, the fixed unit passes through a zero displacement position (neutral) and reverses direction such that the HST power becomes positive again. Greater overall speed is attainable by the hydromechanical transmission  10  in mode  2 , without consuming as much HST power. 
     In mode  1 , the maximum amplitude of the HST power curve is limited by the full stroke displacement of the variable unit  17 V at the maximum input or engine speed in rpm. The fixed unit speed shown as F-unit speed in FIG. 8 is proportional to the speed of the sun gear  38 . The HST power is proportional to the displacement of the variable unit  17 V and the hydrostatic pressure. 
     FIGS. 3-5 and  9  illustrate another configuration of this invention that provides a three-mode hydromechanical transmission  10 A. A new gear set  114  is formed in the same housing  12  by physically replacing some of the gears ( 26 ,  32 A,  32 C,  42 A,  42 B,  42 C,  38 ,  44 ) in the planetary gear set  14  with gears ( 126 ,  132 A,  132 B,  132 C,  142 A,  142 B,  142 C,  138 ,  144 ) to alter the ratios of the gears. The number of teeth or pitch diameter of at least some of the latter gears are different than the former gears so that the gear ratios within the planetary gear set change. A third clutch  154  interconnects the ring gear  126  with the first drive gear  120  through the first intermediate gear  150 . Preferably, the third clutch  154  is disposed on the axis of rotation  33 . When the operator engages clutch  154  while the other two clutches  46 ,  48  are disengaged, a third mode of the transmission  10 A results, as shown in FIG.  9 . FIGS. 3 and 4 show the transmission  10 A with the third clutch  154  disengaged and are quite similar in structure and operation to the previously described two mode or clutch configuration of FIGS. 1 and 2. FIGS. 3-5 depict the power paths and clutch statuses of the three modes. 
     As best seen in view of FIG.  9  and FIGS. 3-5, the operation of the three-mode transmission is similar in some respects to the operation of the two mode configuration. However, as can be seen by comparing FIGS. 8 and 9, mode  2  of the transmission does not result in as high of a transmission output speed as mode  2  in the two mode configuration. Instead, mode  2  is compressed due to the alteration of gear ratios in the planetary gear set. The third clutch  154  is engaged when the F-unit speed reaches its maximum negative amplitude and the output speed of the overall transmission  10 A increases. Modes  1 ,  2  and  3  require progressively less hydrostatic transmission power. Greater output speed is achieved with the same gear layout and hydrostatic transmission. The additional output speed (and horsepower) is accomplished by merely altering the gear ratios and adding another clutch  154  between the ring gear  126  and the first intermediate gear  150 , which is connected to the first input gear  120  and input shaft  18 . 
     Of course, the invention is not limited to particular input power values, output torque, output speed, or gear ratios, but the example described above illustrates a combination possible with the invention. Pertinent values for the 2 mode and 3 mode configurations of the transmission described above are shown below: 
     2 mode: 
     Input (engine) rpm=7000 
     Input Power from engine=15 HP 
     R=Drive/Intermediate=4.0 
     K 1 =R 1 /S 1 =4.75 
     K 2 =R 1 /S 2 =−2.75 
     Output Torque T at two output shafts=325 ft-lb 
     Output Speed N=1750 rpm at two output shafts 
     Overall Corner Power=TN(12/63025)=108 HP 
     3 mode: 
     Input (engine) rpm=7000 
     Input Power=28 HP 
     R=Drive/Intermediate Ratio=5.8 
     K 1 =S 1 /R 1 =3.5 
     K 2 =R 1 /S 2 =−1.5 
     Output Torque T=325 ft-lb 
     Output Speed N=2400 rpm 
     Overall Corner Power=TN(12/63025)=148 HP 
     Where: 
     S 1  is the number of teeth or the pitch circle diameter of the first sun gear  38  or  138 ; 
     S 2  is the number of teeth or the pitch circle diameter of the second sun gear  44  or  144 ; 
     R is the effective ratio between the first drive gear  20  or  120  and the first intermediate gear  50  or  150  (when the first and second drive gears are the same size and the first and second intermediate gears are the same size, as shown in the figures, this is also the effective ratio between the second drive gear  24  or  124  and the second intermediate gear  52  or  152 ); 
     R 1  is the number of teeth or the pitch circle diameter of the ring gear  26  or  126 ; 
     K 1  is the ratio of the first sun gear S 1  to the ring gear R 1 ; and K 2  is the ratio of the ring gear R 1  to the second sun gear S 2 . 
     The signs of K 1  and K 2  indicate whether the gears rotate in the same direction (+) or counter-rotate (−). 
     Thus, it can be seen that the present invention at least accomplishes its stated objectives. 
     In the drawings and specification, there have been set forth preferred embodiments and examples relating to the invention, and although specific terms are employed, these are used in a generic and descriptive sense only and not for purposes of limitation. Changes in the form and the proportion of parts as well as in the substitution of equivalents are contemplated as circumstances may suggest or render expedient without departing from the spirit or scope of the invention as further defined in the following claims.