Patent Abstract:
A power-branched transmission, particularly for agricultural vehicles such as tractors or similar, includes a stepped planetary gear which is disposed between an input shaft and an output shaft and is used for dividing the power supplied at the input shaft onto a mechanical power branch and a hydraulic power branch. The hydraulic power branch is formed by two hydraulically interconnected, identical hydrostatic axial piston engines which can be selectively operated as a pump or an engine, can be swiveled within a predefined pivoting angle, and can be connected to the input shaft or the stepped planetary gear in a different manner via two respective clutches so as to cover different operating ranges or running steps. In order to obtain better efficiency in such a power-branched transmission, the two hydrostatic axial piston engines are configured as wide-angle hydrostats that are provided with a minimum pivoting angle range of 45°.

Full Description:
BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The present invention relates to the field of vehicle transmissions. It refers to a power-branched transmission, in particular for agricultural vehicles, such as tractors or the like, and to a method for the operation of such a transmission. 
   2. Description of Related Art 
   Power-branched transmissions, particularly for use in vehicles employed for agricultural or construction purposes, such as, for example tractors, have been known for a long time. In such power-branched transmissions, the power prevailing at an input shaft or drive shaft and normally output by an internal combustion engine is apportioned to a first mechanical power branch with a fixed ratio and to a second power branch of continuously variable ratio and is subsequently combined again in order to be available at an output shaft or take-off shaft. The second power branch is mostly designed as a hydrostatic branch, in which two hydrostatic axial piston machines (hydrostats) of the oblique axis or swashplate type, which are connected hydraulically to one another, operate selectively as a pump or as a motor. The ratio can in this case be varied by a variation in the pivot angle of the cylinder block or of the swashplate. Apportioning the power to the two power branches and combining the branched powers normally take place by means of a planetary gear. Power-branched transmissions of the type described are disclosed in various embodiments in DE-A1-27 57 300, DE-C2-29 04 572, DE-A1-29 50 619, DE-A1-37 07 382, DE-A1-37 26 080, DE-A1-39 12 369, DE-A1-39 12 386, DE-A1-43 43 401, DE-A1-43 43 402, EP-B1-0 249 001 and EP-A2-1 273 828. 
   So that a power-branched transmission can be used successfully in practice, it should generally be distinguished by the following properties:
         The transmission should have high efficiency over the entire speed range. This should be the case particularly at the high driving speeds which occur in road traction over a lengthy period of time.   The transmission should have a compact construction, in order to allow installation in the most diverse possible vehicles, if possible without structural restrictions.   The transmission should make it possible to transfer high powers.   The transmission should have as simple a construction as possible in order to limit the power losses and increase operating reliability.   The transmission should allow fully comprehensive electronic control in connection with engine management and, even in the event of a failure of specific control elements, make sufficient emergency driving programs available.       

   In DE-A1-43 43 402 initially mentioned, a power-branched transmission designated as an SHL transmission (continuously variable hydrostatic power-branched transmission) has already been described, distinguished by two hydraulically coupled identical hydrostats in the oblique-axis type of construction which can be coupled in different ways to a planetary differential gear via pairs of clutches or selective shift elements K 1 /K 2  or K 3 /K 4 . The known SHL transmission has been used and tested in town buses under the type designation SHL-Z. The two hydrostats used have a pivoting range of only 0-25°. For forward drive, in this case, there are 3 driving steps or driving ranges: in the first driving range, at the starting point, the hydrostatic fraction of the transferred power is 100% and then approaches zero linearly with the speed. In the second driving range, it goes from zero to a maximum of about 27% and then back to zero again. In the third driving range, it goes from zero to a maximum value of 13% at the highest forward speed. 
   The known SHL transmission has the disadvantages not only of the subdivision of the forward drive into three driving ranges, which leads to an increased outlay in shift and control terms but, above all, of the hydrostatic power transfer fraction which deviates markedly from zero at maximum speed. This leads, on long-distance trips in which the high speeds are maintained virtually constantly over a lengthy period of time, to unnecessary efficiency losses which have an adverse effect on consumption and on exhaust gas emission. 
   The object of the invention, therefore, is to provide a continuously variable hydrostatic power-branched transmission which avoids the disadvantages of known transmissions and which is distinguished, in particular, by high and improved efficiency in rapid forward drive, and also to specify a method for the operation of such a transmission. 
   SUMMARY OF THE INVENTION 
   The essence of the invention is, in the transmission configuration described initially, to design the two hydrostatic axial piston machines as wide-angle hydrostats with a pivot angle range of at least 45° and to carry out the adjustment of the angles of the hydrostatic piston machines, the hydraulic connection between the two hydrostatic axial piston machines and the activation of the clutches in such a way that forward drive is subdivided into two successive driving ranges, and in such a way that the fraction of the power transferred over the hydraulic branch in each case approaches zero at the end of each of the two driving ranges. This measure achieves, at high speeds, a vanishing hydrostatic fraction of the power transfer which is manifested at the same time in markedly improved efficiency. The operating values are particularly favorable when, according to a preferred refinement, the two hydrostatic axial piston machines have a pivot angle range of at least 50°. 
   A preferred refinement of the transmission according to the invention is characterized in that the stepped planetary gear comprises double planet wheels mounted rotatably on a planet web and having a smaller gearwheel and a larger gearwheel which mesh with a larger sun wheel and with a smaller sun wheel and which run with the larger gearwheel in a ring wheel, in that the larger sun wheel is coupled to the input shaft, in that the first hydrostatic axial piston machine can be coupled to the ring wheel via a first clutch and to the input shaft via a second clutch, in that the second hydrostatic axial piston machine can be coupled to the planet web via a third clutch and to the smaller sun wheel via a fourth clutch, in that the output shaft is coupled to the planet web, in that the larger sun wheel is seated fixedly in terms of rotation on the input shaft, in that a first spur wheel is flanged to the planet web, and in that a second spur wheel, which meshes with the first spur wheel, is arranged fixedly in terms of rotation on the output shaft. 
   In particular, the power-branched transmission is characterized in that a third spur wheel is flanged to the ring wheel and the coupling of the first hydrostatic axial piston machine by means of the first clutch takes place via a fourth spur wheel which meshes with the third spur wheel, in that a fifth spur wheel is arranged fixedly in terms of rotation on the input shaft and the coupling of the first hydrostatic axial piston machine by means of the second clutch takes place via a sixth spur wheel and a reversing wheel which meshes with the fifth spur wheel and with the sixth spur wheel, in that the coupling of the second hydrostatic axial piston machine by means of the third clutch takes place via a seventh spur wheel which meshes with the first spur wheel, and in that the smaller sun wheel is connected fixedly in terms of rotation to an eighth spur wheel via a hollow shaft surrounding the input shaft, and the coupling of the second hydrostatic axial piston machine by means of the fourth clutch takes place via a ninth spur wheel which meshes with the eighth spur wheel. 
   The power-branched transmission becomes particularly compact when the input shaft is connected fixedly in terms of rotation to a coaxial take-off shaft which passes through the power-branched transmission. 
   The hydrostatic axial piston machines are preferably equipped in each case with a driven shaft, the clutches are designed as hydraulically actuable multiple-disk clutches and arranged on the driven shafts, and the clutches are actuated via axial hydraulic ducts running in the driven shafts. 
   Preferably the input shaft, the stepped planetary gear, the two hydrostatic axial piston machines and the output shaft are accommodated in a space-saving way in a common housing, the two hydrostatic axial piston machines being hydraulically connectable to one another via high-pressure ducts running in the housing. The housing comprises a housing lower part and a housing upper part, on the housing upper part is arranged a high-pressure block in which the high-pressure ducts are accommodated, the two hydrostatic axial piston machines in each case comprise a cylinder block with a plurality of cylinder bores and with pistons mounted displaceably therein, which cylinder block is mounted in a pivot housing rotatably about a horizontal axis, the pivot housings are mounted in each case with an upper bearing journal in the high-pressure block pivotably about a vertical pivot axis, and the cylinder bores are connected to the high-pressure ducts in the high-pressure block via connecting ducts running in the pivot housing into the upper bearing journals. 
   In particular, in each of the two hydrostatic axial piston machines, the cylinder bores of the cylinder block which lie above a horizontal mid-plane can be connected to an upper connecting duct via upper orifices in the pivot housing and the cylinder bores of the cylinder block which lie below the horizontal mid-plane can be connected to a lower connecting duct via lower orifices in the pivot housing, the upper connecting ducts being connected to first high-pressure ducts and the lower connecting ducts to second high-pressure ducts in the high-pressure block, and the first and the second high-pressure ducts being selectively connectable to one another by means of valves accommodated in the high-pressure block. 
   The power-branched transmission is particularly compact and operationally reliable when the first and second high-pressure ducts in the high-pressure block are produced by casting and when the valves are designed as hydraulically actuable valves and are accommodated in bores which are introduced into the high-pressure block transversely to the high-pressure ducts. The hydraulically actuable valves are, in particular, activated in pairs via first electromagnetic valves. 
   Good emergency driving properties arise when, for redundancy reasons, a second electromagnetic valve is in each case connected in parallel to the first electromagnetic valves or an electromagnetic equivalent coil is assigned to the latter. 
   A pivot bolt is preferably arranged in each case on the pivot housings of the hydrostatic axial piston machines at a predetermined radial distance from the upper bearing journal, and hydraulic cylinders are provided on the housing upper part, and engage on the pivot bolts in order to pivot the pivot housings. The hydraulic cylinders are activated via third electro-magnetic valves, and, for redundancy reasons, the third electromagnetic valves are assigned in each case an electromagnetic equivalent coil. 
   The clutches are activated via electromagnetic valves, the electromagnetic valves are accommodated in valve plates flanged to the housing, and the hydraulic connection between the electromagnetic valves and the clutches takes place via ducts running in the housing, one of the clutches being provided for coupling the second hydrostatic axial piston machine to the output shaft, and, for redundancy reasons, a further electro-magnetic valve is connected in parallel to the electromagnetic valve assigned to this clutch. 
   A preferred refinement of the method according to the invention is distinguished in that the stepped planetary gear comprises double planet wheels mounted rotatably on a planet web and having a smaller gearwheel and a larger gearwheel which mesh with a larger sun wheel and with a smaller sun wheel and which run with the larger gearwheel in a ring wheel, in that the larger sun wheel is coupled to the input shaft and the output shaft is coupled to the planet web, in that, in the first driving range, the first hydrostatic axial piston machine is coupled to the ring wheel via a first clutch and the second hydrostatic axial piston machine is coupled to the planet web via a second clutch, and the first hydrostatic axial piston machine is operated as a pump and the second hydrostatic axial piston machine as a motor, and in that, in the second driving range, the first hydrostatic axial piston machine is coupled to the ring wheel via the first clutch and the second hydrostatic axial piston machine is coupled to the smaller sun wheel via a third clutch, and the first hydrostatic axial piston machine is operated as a motor and the second hydrostatic axial piston machine as a pump. 
   In particular, to run through the first driving range, the first hydrostatic axial piston machine, starting from the pivot angle 0°, runs through the entire pivot angle range up to the maximum pivot angle, and the second hydrostatic axial piston machine, starting from the maximum pivot angle, runs through the entire pivot angle range up to the pivot angle 0°, and, to run through the second driving range, the first hydrostatic axial piston machine, starting from the maximum pivot angle, runs through the entire pivot angle range up to the pivot angle 0°, and the second hydrostatic axial piston machine, starting from the pivot angle 0°, runs through the entire pivot angle range up to the maximum pivot angle. 
   It is particularly beneficial if the first hydrostatic axial piston machine can be coupled to the input shaft via a fourth clutch, and if, for a temporary increase in traction, the first hydrostatic axial piston machine is coupled simultaneously to the ring wheel via the first clutch and to the input shaft via the fourth clutch. 
   Preferably, the clutches are designed as hydraulically actuated multiple-disk clutches and the clutches, when actuated, are acted upon by a shift pressure which depends on the high pressure prevailing in the hydraulic connection between the hydrostatic axial piston machines. 
   When the adjustment of the angles of the hydrostatic axial piston machines, the hydraulic connection between the two hydrostatic axial piston machines and the activation of the clutches take place via electromagnetic valves, and, for the electromagnetic valves, equivalent means are provided which, in the event of a failure of one or more of the electro-magnetic valves, can be used in order to maintain essential functions of the power-branched transmission, an emergency driving program can be implemented in that, in the event of a failure of one or more of the electromagnetic valves, the associated equivalent means are used, in particular additional parallel-connected electromagnetic valves and/or equivalent coils for the electromagnetic valves being used as equivalent means. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The invention will be explained in more detail below by means of exemplary embodiments, in conjunction with the drawing in which: 
       FIG. 1  shows the transmission diagram of a power-branched transmission for a tractor according to a preferred exemplary embodiment of the invention; 
       FIG. 2  shows an exploded illustration of the embodiment of a power-branched transmission according to the transmission diagram of  FIG. 1 ; 
       FIG. 3  shows a longitudinal section through the arrangement of the stepped planetary gear and of the two parallel hydrostatic axial piston machines cooperating therewith, from  FIG. 2 , along a plane passing through the three axes; 
       FIG. 4  shows a longitudinal section through one of the two hydrostatic axial piston machines, with attached double clutch, from  FIG. 2 ; 
       FIG. 5  shows a front view ( FIG. 5   a ) and a “transparent” side view ( FIG. 5   b ) of the pivot housing with the internal ducts of the axial piston machine from  FIG. 4 ; 
       FIG. 6  shows a “transparent” illustration of the interior of the high-pressure block for the valve-controlled hydraulic connection of the two axial piston machines from  FIG. 2  in a side view ( FIG. 6   a ) and in a top view ( FIG. 6   b ), and in a first valve position; 
       FIG. 7  shows a “transparent” illustration of the interior of the high-pressure block for the valve-controlled hydraulic connection of the two axial piston machines from  FIG. 2  in a side view ( FIG. 6   a ) and in a top view ( FIG. 6   b ), in a first valve position; 
       FIG. 8  shows a three-dimensional model illustration of the power-branched transmission from  FIG. 2 , the pivot angle of the two axial piston machines and the shift state of the two double clutches being revealed; 
       FIG. 9  shows, in the type of illustration from  FIG. 8 , the control of the clutches and of the pivot angle of the axial piston machines in the two forward driving steps ( FIG. 9(   a   1 )-( a   3 ) and  FIG. 9(   b   1 )-( b   3 )) and in one reverse step ( FIG. 9   c ); 
       FIG. 10  shows a graph of the efficiency and of the percentage fraction of hydraulic power in the transferred power as a function of speed in the two forward driving steps of the transmission according to  FIG. 1 and 2 ; and 
       FIG. 11  shows a simplified hydraulic diagram of the power-branched transmission according to  FIG. 2 . 
   

   DETAILED DESCRIPTION OF THE INVENTION 
     FIG. 1  illustrates the transmission diagram of a power-branched transmission for a tractor according to a preferred exemplary embodiment of the invention. The power-branched transmission  10  transfers the power of an internal combustion engine  20  which is symbolized in  FIG. 1  by a piston seated on a crankshaft. The power-branched transmission  10  is connected to an input shaft (drive shaft)  12  via a cardan shaft  11  and to the internal combustion engine  20  via a torsion damper  19 . It outputs the transferred power via an output shaft (driven shaft)  18  and an axial power divider.  39  with a longitudinal differential LD and with a longitudinal differential lock LDS to an axle connection  40  to the front axle and an axle connection  41  to the rear axle. Coupling between the output shaft  18  and the axial power divider  39  takes place via two intermeshing gearwheels z 14  and z 15 . 
   A take-off shaft  17  extends through the power-branched transmission  10  and is a direct continuation of the input shaft  12 . The take-off shaft  17  drives via gearwheels z 18 , z 19  and z 20  a first pump  24  for the working hydraulics and a second pump  24 ′ for steering. Power can be taken off from the take-off shaft  17  from outside via a clutch  16 . A third pump  24 ″ for emergency steering is driven by the output shaft  18  via gearwheels z 21 , z 22 . Further pumps are the feed pump  42  and the lubricating-oil pump  42 ′ which, seated on a common axle, are driven by the input shaft  12  via. the backgear consisting of gearwheels z 16 , z 11 , z 12  and z 10 . 
   The core of the power-branched transmission  10  is formed by a stepped planetary gear  15  with a large sun wheel z 1  and with a small sun wheel z 1 ′, with double planet wheels z 2  and z 2 ′, the ring wheel z 3 , and with a planet web  49  connected fixedly in terms of rotation to a gearwheel z 8  (see also  FIG. 3 and 8 ), and by two hydrostatic axial piston machines H 1  and H 2 , the driven shafts  13  and  14  of which can be coupled in each case via a pair of clutches K 3 , K 4  and K 1 , K 2  in different ways to the input shaft  12 , the output shaft  18  and the stepped planetary gear  15 . The hydrostatic axial piston machines H 1  and H 2 , which operate selectively as a pump and as a motor, are connected hydraulically to one another via high-pressure lines  21 ,  22  which can be cross-switched by means of a multiway valve  23 . The first axial piston machine H 1  can be coupled with its driven shaft  13  to the ring wheel z 3  by means of the clutch K 3  via a backgear consisting of the gearwheel z 5  and of a gearwheel z 4  connected fixedly in terms of rotation to the ring wheel z 3 . It may, however, also be coupled to the input shaft  12  by means of the clutch K 4  via the gearwheel z 11 , the intermediate wheel z 12  and the gearwheel z 10  arranged fixedly in terms of rotation on the input shaft  12 . The second axial piston machine H 2  can be coupled with its driven shaft  14  to the planet web  49  and consequently to the output shaft  18 , on the one hand, by means of the clutch K 1  via the hollow shaft  26  and the gearwheel z 9  which is arranged fixedly in terms of rotation on the latter and which meshes with the gearwheel z 8 . It may, on the other hand, be coupled to the smaller sun wheel z 1 ′ of the stepped planetary gear  15  by means of the clutch K 2  via the pair of gearwheels z 7 , z 6  and the hollow shaft  25 . 
   The power prevailing at the input shaft  12  is apportioned in the power-branched transmission  10  by means of the stepped planetary gear  15  to two power branches, to be precise a mechanical power branch and a hydraulic power branch, and is combined again later at the output shaft  18 . The mechanical power branch runs from the input shaft  12  via the larger sun wheel z 1  connected fixedly in terms of rotation to the input shaft  12 , the double planet wheels z 2 , the planet web  49  and the gearwheel z 8 . The hydraulic power branch runs via the two hydraulically connected axial piston machines H 1  and H 2  and is designed differently according to the shifting of the clutches K 1 , . . . , K 4 . 
   To explain the functioning of the power-branched transmission  10  from  FIG. 1 , this is reproduced once again in  FIG. 8  in model form in a three-dimensional illustration. The output side between the gearwheel z 8  and the axial connections  40 ,  41  is in this case illustrated in simplified form, as compared with  FIG. 1 . The same applies to the input side between the internal combustion engine  20  and the input shaft  12 . The clutches K 1 , . . . , K 4  are designed (as in  FIG. 1 ) as multiple-disk clutches, and the hydrostatic axial piston machines H 1 , H 2  are of the oblique-axis type, in which the cylinder block together with the pistons located in it can be pivoted out of the axis of the driven shafts  13 ,  14  to one side over a pivot angle, the maximum value of which is at least 45°, preferably 50° and above (what are known as “wide-angle hydrostats”). By means of the power-branched transmission  10  from  FIG. 8  and the wide-angle hydrostats H 1 , H 2 , operation can be implemented in which forward drive can be covered overall by only two driving ranges or driving steps, at the upper end of which in each case the hydrostatic fraction of the transferred power approaches zero. 
   The shifting of the clutches K 1 , . . . , K 4  and the pivoting position of the hydrostats H 1 , H 2  for the various operating states of the transmission are illustrated in  FIG. 9 ,  FIG. 9(   a   1 ) to  9 ( a   3 ) showing the first forward driving step,  FIG. 9(   b   1 ) to  9 ( b   3 ) the second forward driving step and  FIG. 9(   c ) reverse drive. During starting ( FIG. 9(   a   1 )), as in the entire first forward driving step, the clutches K 3  and K 1  are actuated, so that the first hydrostat H 1  is coupled to the ring wheel z 3  of the stepped planetary gear  15  and the second hydrostat H 2  is coupled to the planet web or the gearwheel z 8  or the output shaft  18  (the driven side of the actuated clutch is in each case colored dark in  FIG. 9) . The first hydrostat H 1 , which operates as a pump in the first forward driving step, is first unpivoted (pivot angle 0°), whereas the second hydrostat H 2  operating as a motor is pivoted out fully (maximum pivot angle). On account of the zero position of the first hydrostat H 1 , no pressure medium is pumped to the second hydrostat H 2  and therefore no power is transferred hydraulically either. The starting operation is initiated in that the first hydrostat H 1  is gradually pivoted, volume increasingly being pumped to the second hydrostat H 2 , and the second hydrostat beginning to rotate with a high torque and increasing speed. When the first hydrostat H 1  is pivoted out fully ( FIG. 9(   a   2 )), the first phase of the first driving step is concluded. In the second phase, with the first hydrostat H 1  pivoted out fully, the second hydrostat H 2  is gradually moved back from the maximum pivot angle to the pivot angle 0° ( FIG. 9(   a   3 )), the rotational speed being increased ever further with a decreasing torque. At the end of the first driving step, the second hydrostat H 2  no longer receives torque, and the rotational speed of the first hydrostat H 1  approaches zero. The hydrostatically transferred power approaches zero, and the entire power is transferred mechanically. 
   At the transition from the first driving step to the second driving step ( FIG. 9(   a   3 )→ FIG. 9(   b   1 )) the clutch K 1  is opened and the clutch K 2  is closed. Since the second hydrostat H 2  receives no torque at the pivot angle 0°, the changeover takes place virtually without shift torque. The second hydrostat H 2  is then coupled to the smaller sun wheel z 1 ′ of the stepped planetary gear  15 . With the changeover of the clutches K 1  and K 2 , the multiway valve  23  ( FIG. 1)  is also changed over, so that the hydraulic connections between the two hydrostats H 1  and H 2  are interchanged. In the second driving step, the first hydrostat H 1  operates as a motor and the second hydrostat H 2  as a pump. As in the first driving step, the hydrostat operating as a pump (now the second hydrostat H 2 ) in a first phase, starting from the pivot angle 0°, is gradually pivoted out to the maximum pivot angle ( FIG. 9(   b   2 )), while the hydrostat operating as a motor (now the first hydrostat H 1 ) remains fully pivoted out. In a subsequent second phase ( FIG. 9(   b   2 )→ FIG. 9(   b   3 )), the first hydrostat H 1  is then pivoted back into the zero position. At the end of the second driving step, the hydraulically transferred power again approaches zero; the entire power is transferred via the mechanical power branch. 
   The graph obtained for a power branched transmission in a tractor according to  FIG. 1  or  FIG. 8 , of the efficiency η in % and of the percentage fraction of the hydrostatically transferred power HP is reproduced in  FIG. 10  as a function of the vehicle speed v. Curve A shows the profile of the efficiency η, and curve B shows the profile of the fraction of the hydrostatically transferred power. On account of the wide-angled hydrostats used in the transmission, the entire driving range extending from 0 to 63 km/h can be subdivided into only two driving steps, the first driving step extending from 0 to about 18 km/h and the second driving step from about 18 km/h to 63 km/h. In the first driving step, the fraction of the hydrostatically transferred power goes from an initial 100% linearly down to 0. In the second driving step, the fraction of the hydrostatically transferred power rises from 0 to a maximum of almost 30% (at about 30 km/h) and then falls (at about 53 km/h) to 0 and stays there until the upper end of the driving step. The result of this is that efficiency does not fall again until the end of the second driving step but, instead, even increases. This results, for high driving speeds maintained when driving long distances for a lengthy period of time, in a particularly good efficiency of the transmission which leads to markedly lowered operating costs. 
   In reverse drive ( FIG. 9(   c )), starting from the situation from  FIG. 9(   a   1 ), there is a changeover from the clutch K 3  to the clutch K 4 . The multiway valve  23  in the hydraulic connection between the hydrostats H 1  and H 2  is likewise changed over. The first hydrostat operating as a pump is then driven directly by the input shaft  12  and, starting from 0°, is gradually pivoted out. The fully pivoted-out second hydrostat H 2  then receives rotational speed with high torque. 
   A power-branched transmission implemented according to the transmission diagram from  FIG. 1  is reproduced in an exploded illustration in  FIG. 2 . The power-branched transmission  10  is accommodated in a multipart housing which is composed of a trough-shaped housing lower part  27 , a shallow housing upper part  28 , a front housing cover  29  and a rear housing cover  29 ′. In the lowest part of the housing, the axial power divider  39  is arranged, which has an output forward and rearward for the front axle and the rear axle. Directly above the axle power divider  39 , the stepped planetary gear  15  is fastened, axially parallel, to the side walls of the housing lower part  27  by means of an upper bearing bridge  38 . On the rear part of the upper bearing bridge  38 , two circular upper bearing orifices  36  are provided for receiving the upper bearing journals ( 46  in  FIG. 4 ) of the pivot housings ( 44  in  FIG. 4 ) of the hydrostats H 1  and H 2 . Corresponding lower bearing orifices  37  for receiving the lower bearing journals ( 47  in  FIG. 4 ) of the hydrostats H 1  and H 2  are arranged on a lower bearing bridge  96  which serves at the same time for mounting the two hydrostats H 1  and H 2 . The hydrostats H 1  and H 2  are placed, axially parallel to the axle power divider  39 , on both sides below the stepped planetary gear  15 . They project with the front ends of their driven shafts  13  and  14  through the front wall of the housing lower part  27  and are connected there, by means of the front housing cover  29  equipped with corresponding connection devices, to a hydraulic control located in the housing upper part  28 . The hydraulic control, comprising two valve plates  92 ,  93  with electromagnetic valves (V 11 , . . . , V 15  in  FIG. 11 ), activates the clutches K 1 , . . . , K 4  seated on the. driven shafts  13 ,  14  via the hydraulic ducts  74 , . . . ,  77  ( FIG. 3 ) running in the driven shafts  13 ,  14 . 
   The rear housing cover  29 ′ contains the pump  24 ″ for emergency steering, which is driven by the output shaft  18 . Flanged on the outside to the rear housing cover  29 ′ is a drive unit which comprises the take-off shaft  17  and the two pumps  24  and  24 ′ for the working hydraulics and the steering respectively. 
   The housing upper part  28  contains, in addition to the hydraulic control for the clutches K 1 , . . . , K 4 , further control and connection elements  31 , . . . ,  33  and V 9 , V 10  for the hydrostats H 1  and H 2 . The functioning and configuration of these control and connection elements depend on the internal construction of the hydrostats H 1 , H 2  used. This internal construction is illustrated by the example of the hydrostat H 1  in  FIG. 4 . The hydrostat H 1  is an oblique-axis hydrostat with a driven shaft  13  rotating about a fixed axis  72  and with a cylinder block  70  which rotates about a pivotable axis  73  and which is mounted in a pivot housing  44 . The pivot housing  44  with the cylinder block  70  can be pivoted about the pivot axis  45  by means of a pivot bolt  48 . 
   The hydrostatic axial piston machine or hydrostat Hl of  FIG. 4  comprises an elongate driven shaft  13 , the cylinder block  70 , a plurality of pistons  67  and a synchronizing shaft  63  for synchronizing the rotations of the driven shaft  13  and cylinder block  70 . At one end, which faces the cylinder block  70 , the driven shaft  13  is thickened and ends in a flange  52  concentric to the axis  72  of the driven shaft  13 . Nine circular cylindrical bearing receptacles into which spherical bearings  58  for the pivotable mounting of the pistons  67  are inserted, are milled, distributed uniformly about the axis  72  on a partial circle, into the end face of the flange  52 . 
   Provided in the center of the flange  52  is a funnel-shaped orifice  53  which merges, further inside the driven shaft  13 , into a central bore  55  of stepped diameter. Three axially parallel bores  54  arranged in each case so as to each be rotated at 120° are introduced into the driven shaft  13  around the bore  55  so as to overlap partially with the bore  55  and are part of a first tripod joint  62 . Comparable bores are present, opposite them, in the cylinder block  70  and are part of a second tripod joint  64 . The two tripod joints  62  and  64  allow a rotationally fixed coupling of the synchronizing shaft  63  to the driven shaft  13  and the cylinder block  70  in the case of a simultaneous pivotability of the cylinder block  70  in relation to the flange  52  or the driven shaft  13 . For this purpose, the synchronizing shaft  63  is equipped at each of the two ends with three radially oriented cylindrical journals which are arranged so as to be rotated through 120° and which, in the case of the first tripod joint  62  extend from the central bore  55  through the laterally open overlap region into the adjacent bores  54 . A comparable engagement of the journals also takes place in the second tripod joint  64 . To reduce the play, rings  57 , crowned on the outside, are drawn onto the journals in each case. 
   When the cylinder block  70  is pivoted with respect to the flange  52 , the distance to be bridged between the cylinder block  70  and the flange  52  by the synchronizing shaft  63  changes. So that this distance change can be compensated, the synchronizing shaft  63  is mounted displaceably in the axial direction in the region of the first tripod joint  62 . The synchronizing shaft  63  is seated pivotably with its end facing the cylinder block  70  on a first pressure pin  65  which is inserted into the cylinder block  70  and projects with a portion of its length out of the cylinder block  70 . So that the synchronizing shaft  63  does not come out of engagement with the cylinder block in the second tripod joint  64 , it is pressed in the axial direction, with prestress, against the second pressure pin  61 . A compression spring  59  accommodated in the bore  55  serves for generating the prestress and presses onto the synchronizing shaft  63  via an axially displaceable pressure piston  60  and a second pressure pin  61 . The pressure piston  60 , pressure pins  61 ,  65  and synchronizing shaft  63  have in each case a central oil duct. 
   The (cylindrical) cylinder block  70  has nine axially parallel cylinder bores  68  which are distributed uniformly about its axis  73  on a partial circle and which are in each case at an angular distance of 40° from one another. The cylinder bores  68  are designed, from the side facing the flange  52 , as blind bores. The pistons  67 , which are mounted pivotably in the flange  52 , penetrate from this side into the cylinder bores  68 . For this purpose, each piston  67  has an elongate, downwardly tapering piston shank  67 ′ merging at the lower end into a spherical head  66  with which it is mounted pivotably in the associated spherical bearing  58 . The cylinder block  70  can be pivoted by means of the pivot housing  44  about the pivot axis  45 . The maximum pivot angle amounts to at least 45° and is preferably greater than or equal to 50°. 
   If, in the case of a constant pivot angle ≠0°, the driven shaft  13  and consequently, via the synchronizing shaft  63 , also the cylinder block  70 , are rotated about their respective axes  72  and  73 , each of the nine pistons  67  executes for each revolution a complete spoke cycle. The hydrodynamic axial piston machine H 1  can in this case operate as a hydraulic pump when drive takes place via the driven shaft  13 , and a hydraulic medium is sucked in by the pistons  67  moving out of the cylinder bore  68  and is pressed out by the pistons moving into the cylinder bore  68 . The volumetric pumping capacity for each revolution is in this case the higher, the greater the pivot angle α is. It may, however, also operate as a hydraulic motor when the cylinders are acted upon in each case by a hydraulic medium under pressure, and when the rotational movement occurring is picked up at the driven shaft  13 . In this case, the torque is the higher the greater the pivot angle is. If, by contrast, high rotational speeds are to be achieved at the driven shaft  13 , the pivot angle must be made small. 
   The working space in the cylinder bores  68  which is delimited by the pistons  67  is accessible from the outer end face of the cylinder block  70  through connecting orifices  69 . Through an axial bearing  50 , the connecting orifices  69  of the cylinder bores  68  are successively connected, depending on the rotary position of the cylinder block  70 , to a plurality of upper and lower orifices  82  and  83  in the adjacent pivot housing ( FIG. 5(   a )). The upper and lower orifices  82  and  83  in the pivot housing  44  are connected to an upper and lower connecting duct  80  and  81  ( FIG. 5(   b )). The connecting ducts  80 ,  81  produced by casting run in the pivot housing  44  from the upper and lower orifices  82 ,  83  upward into the upper bearing journal  46  where they end in connecting orifices  78 ,  79  arranged one above the other and are separated by cylindrical sealing surfaces  97 . According to  FIG. 6 and 7 , the hydraulic connection between the two hydrostats H 1 , H 2  can be made via the connecting orifices  78 ,  79  and  78 ′,  79 ′ in the upper bearing journals  46  and  46 ′ of the two hydrostats H 1  and H 2 . 
   A high-pressure block  31  arranged on the housing upper part  28  serves for making (and controlling) the hydraulic connection between the hydrostats H 1  and H 2  ( FIG. 2 ,  6  and  7 ). According to  FIG. 6(   a ) and  7 ( a ), the two hydrostats H 1 , H 2  project with their upper bearing journals  46 ,  46 ′ into corresponding bores in the high-pressure block  31 . Within the high-pressure block  31 , high-pressure ducts  84 , . . . ,  87  are formed by casting, which, in the region of the upper bearing journals  46 ,  46 ′, end in two annular chambers which lie one above the other, are sealed off with respect to one another at the sealing surfaces  97  and are connected to the connecting orifices  78 ,  78 ′,  79 ,  79 ′ of the upper bearing journals  46 ,  46 ′. The high-pressure ducts  84 , . . . ,  87  lead from the upper bearing journals  46 ,  46 ′ to a valve block  88  which is arranged in the middle of the high-pressure block  31  and where they can be connected selectively to one another by means of four hydraulically actuable valves V 1 , . . . , V 4 . The valves V 1 , . . . , V 4  are accommodated in transversely running bores, in which in each case a piston is pressed with spring pressure against a sealing surface. The valves V 1 , . . . , V 4  open counter to the spring pressure when the high-pressure ducts,  84 , . . . ,  87  are acted upon by high pressure. They can be closed by means of a counterpressure with which the pistons of the valves V 1 , . . . , V 4  are acted upon from the rear via laterally flanged-on activation plates  89 ,  90 . The counterpressure is controlled by means of an electromagnetic control valve  91 . 
   The valves V 1 , . . . , V 4  in the valve block  88  are activated in pairs. In the illustration of  FIG. 6 , the valves V 1  and V 4  are open, whereas the valves V 2  and V 3  are closed. In this case, the upper connecting orifice  78  of the first hydrostat H 1  (H 1 O) is connected via the high-pressure ducts  84  and  87  and the valve V 1  to the lower connecting orifice  79 ′ of the second hydrostat (H 2 u). The lower connecting orifice  79  of the first hydrostat H 1  (H 1 u) is likewise connected via the high-pressure ducts  85  and  86  and the valve V 4  to the upper connecting orifice  78 ′ of the second hydrostat H 2  (H 2 o). This valve switching (V 1 , V 4  open, V 2 , V 3  closed) illustrated in  FIG. 6 , is provided for the first driving step of the transmission, in which the first hydrostat H 1  operates as a pump and the second hydrostat H 2  as a motor. In the second driving step, according to  FIG. 7 , the conditions are reversed: the valves V 1  and V 4  are closed, whereas the valves V 2  and V 3  are open. In this case, the two lower connecting orifices  79  and  79 ′ and the two upper connecting orifices  78  and  78 ′ are in each case connected to one another. 
   Supply lines are led outward in the high-pressure block  31  from the high-pressure ducts  86 ,  87 , so that the pressures prevailing in the ducts can be measured and monitored via pressure transducers. Other supply lines make it possible to supply hydraulic medium into the circuit existing between the hydrostats H 1 , H 2 . Arranged behind the high-pressure block  31 , on the housing upper part  28 , are two oblique hydraulic cylinders  32 ,  33  which are activated by electromagnetic valves V 9  and V 10  and which engage on the pivot bolts  48  ( FIG. 4 ) which project into the housing upper part  28  and which are arranged at a radial distance from the pivot axis  45  on the pivot housing  44  of the hydrostats H 1 , H 2 . 
   The resulting hydraulic diagram of the power-branched transmission  10  from  FIG. 1-7  is reproduced in simplified form in  FIG. 11 . The necessary lubricating-oil and feed pressure is generated by a lubricating-oil pump  42 ′ and a following feed pump  42 . The feed pressure is available at a first pressure accumulator  94 . It is used for actuating the clutches K 1 , . . . , K 4 , control taking place via the valves V 11 , . . . , V 15  which are accommodated in the valve plates  92 ,  93  and which are designed as electromagnetic multiway valves. The clutch K 1  can be actuated, for redundancy reasons by two identical valves V 14  and V 15  which are interconnected by means of a shuttle valve. The pressure accumulator  94  for the feed pressure is connected via nonreturn valves having antiparallel-connected pressure limiters to the high-pressure ducts  84 , . . . ,  87  in the high-pressure block  31  which can be interconnected in the way already described by means of the valves V 1 , . . . , V 4 . The valves V 1 , . . . , V 4  are activated in pairs via electromagnetic valves V 5  and V 6 , to which further valves V 7  and V 8  are connected in parallel as redundant equivalent valves by means of shuttle valves. 
   A second pressure accumulator  95  is connected via a shuttle valve to the two pressure ducts  86  and  87 . The pressure for actuating the valves V 1 , . . . , V 4  is extracted from this pressure accumulator  96 . The two hydraulic cylinders  32 ,  33  for pivoting the hydrostats H 1  and H 2  are also actuated by means of the same pressure. To control the hydraulic cylinders  32 ,  33 , the electromagnetic valves V 9  and V 10  ( FIG. 6 ,  7 ) are used which, for redundancy reasons, have additional equivalent coils  34 ,  35 . 
   The overall control and monitoring of the transmission as a function of the engine data and of the torque and driving speed requirements and also the changeover to an emergency driving program in the event of a failure of specific control elements are assumed by an electronic transmission control unit  43  ( FIG. 2 ) which is placed in the immediate vicinity of the measurement transducers (for pressure, rotational speed and valve position, etc.) and control valves on the housing upper part  28 . Integrating the control and monitoring functions of the transmission, including the switchable high-pressure ducts  84 , . . . ,  87  for the hydraulic connection of the hydrostats H 1 , H 2  into the housing upper part  28  results in a highly compact transmission construction, at the same time with high operating reliability. Owing to the built-in redundancy, in the event of a failure of specific control elements, an emergency driving program can be implemented which in most cases allows further travel without restriction, and in other cases ensures at least restricted driving home or to the nearest garage. If, for example, the main coils of the valves V 9  and/or V 10  for controlling the hydraulic cylinders  32  and  33  fail, further travel without restrictions can be ensured, using the equivalent coils  34  and/or  35  (or using complete equivalent valves). The same also applies to the situation where the valve V 12  for activating the clutch K 1  fails, because the equivalent valve V 15  then can assume its role. If the activation for the clutch K 2  fails, driving can still take place in the first driving step (and in reverse). If the activation of the clutch K 3  fails, a restricted forward drive without the second driving step (and unrestricted reverse drive) can be implemented by the engagement of the clutch K 4  and the simultaneous changeover of the valves V 1 , . . . , V 4 . If the activation of the clutch K 4  fails, forward drive is not restricted. Restricted reverse drive is then achieved by the activation of the clutch K 3 . If one of the valves V 5  and V 6  for the high-pressure duct changeover fails, full drivability can be restored, using the corresponding equivalent valve V 7  or V 8 . 
   Finally, because of the special configuration of the power-branched transmission  10 , it is conceivable, within the framework of transmission control, to achieve a temporary increase in traction by the simultaneous closing of the clutches K 3  and K 4 , since additional mechanical force transmission thereby becomes effective.

Technology Classification (CPC): 5