Abstract:
The invention relates to a manual transmission having an input shaft ( 12 ), one first and one second mechanical gear branch ( 26, 28 ) that may be coupled in a driving fashion on the input side with the input shaft and on the output side via various gears (G 1, . . . ,  G 7, R) with a common output shaft ( 40 ), and one first and one second hydrostatic machine ( 18, 18′, 20, 20′ ), each comprising a primary part ( 16 ), a secondary part ( 22, 24 ), and one first and one second pressure chamber. The primary part and the secondary part of each hydrostatic machine ( 18, 18′, 20, 20 ′) are rotatable relative to one another, wherein the secondary part ( 22 ) of the first hydrostatic machine ( 18, 18′ ) is operatively connected to the first mechanical gear branch ( 26 ) and the secondary part ( 24 ) of the second hydrostatic machine ( 20, 20 ′) is operatively connected to the second mechanical gear branch ( 28 ). At least one pressure control device is associated with the hydrostatic machines, by means of which the first pressure chamber of the first hydrostatic machine ( 18, 18′ ) may be hydraulically coupled to the first pressure chamber of the second hydrostatic machine ( 20, 20′ ) and the second pressure chamber if the first hydrostatic machine ( 18, 18′ ) may be hydraulically coupled to the second pressure chamber of the second hydrostatic machine ( 20, 20′ ) so as to equalize pressure between the two hydrostatic machines, particularly for shifting gears.

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
       [0001]    This application is a 371 National Stage of International Application No. PCT/EP2008/005608 filed Jul. 9, 2008. This application claims the benefit of German patent application No. DE 10 2007 038 175.3 filed Aug. 13, 2007. The disclosures of the above-listed applications are incorporated herein by reference. 
     
    
     FIELD 
       [0002]    This section provides a general summary of the disclosure and is not a comprehensive disclosure of its full scope or all of its features. 
         [0003]    The invention relates to a shift transmission of a motor vehicle, having an input shaft and a first mechanical transmission branch and a second mechanical transmission branch which can be coupled drive-wise to the input shaft at the input side and via different gear stages to a common output shaft at the output side. 
       BACKGROUND 
       [0004]    This section provides background information related to the present disclosure which is not necessarily prior art. 
         [0005]    Conventional transmissions which allow a shifting under load free of interruption of the driving power—so-called power shift transmissions—as a rule have a number of coupling elements and actuators to be able to carry out a gear stage change which is hardly noticeable for the driver and is therefore comfortable. Known power shift transmissions—dual clutch transmissions are usually used in passenger vehicles—include a plurality of components prone to wear and are therefore undesirably complex. The control of these power shift transmissions is moreover relatively complex and/or expensive. 
       SUMMARY 
       [0006]    This section provides a general summary of the disclosure and is not a comprehensive disclosure of its full scope or all of its features. 
         [0007]    It is the underlying object of the invention to provide a shift transmission which can be shifted under load without the driving comfort being impaired by shift procedures. The components required for this and the control of the shift transmission should be as simple and as robust as possible. The shift transmission should furthermore be designed such that a plurality of driving conditions of the vehicle can be managed without special components being required for this purpose. 
         [0008]    The shift transmission in accordance with the invention has, as initially described, a first mechanical transmission branch and a second mechanical transmission branch which can be connected drive-wise to the input shaft at the input side and via different gear stages to a common output shaft at the output side. The shift transmission furthermore includes a first hydrostatic machine and a second hydrostatic machine which each have a primary part, a secondary part and a first pressure space and a second pressure space, wherein the primary part and the secondary part of the respective hydrostatic machine are rotatable relative to one another. The secondary part of the first hydrostatic machine is operatively connected to the first mechanical transmission branch and the secondary part of the second hydrostatic machine is operatively connected to the second mechanical transmission branch. At least one pressure control device is associated with the hydrostatic machines, by means of which the first pressure space of the first hydrostatic machine can be selectively hydraulically coupled with the first pressure space of the second hydrostatic machine and, hydraulically separately therefrom, the second pressure space of the first hydrostatic machine can be selectively hydraulically coupled with the second pressure space of the second hydrostatic machine in order to bring about a pressure balance between the two hydrostatic machines—in particular for a gear stage change. 
         [0009]    The shift transmission in accordance with the invention thus includes two separate mechanical transmission branches, in particular transmission branches having stand gears or epicycle gears which are each provided for the formation of specific gear stages. For example, the odd gear stages can be formed with the first transmission branch, whereas the second transmission branch is provided for the realization of the even gear stages and of the reverse gear. 
         [0010]    The shift transmission in accordance with the invention furthermore includes a first hydrostatic machine and a second hydrostatic machine which are each associated with one of the two transmission branches. The drive-wise coupling of the input shaft to the respective transmission branch can be controlled by the hydrostatic machines, i.e. the driving torque of the input shaft can be transferred as required to the output shaft via one of the transmission branches or—in particular on a gear stage change—via both transmission branches simultaneously. For this purpose, the secondary part of the first hydrostatic machine is operatively connected to the first mechanical transmission branch—that is, for example, directly rotationally fixedly connected or indirectly connected via a transmission—whereas the secondary part of the second hydrostatic machine is operatively connected to the second mechanical transmission branch. 
         [0011]    The driving torque transferred from the input shaft to the mechanical transmission branch is a function of the fluid pressures present in the pressure spaces of the hydrostatic machines. The degree of the coupling between the respective primary parts and the secondary parts can be modified by an intervention in the hydraulic system of the hydrostatic machines. In other words, the degree of the coupling depends on the fluid throughput, i.e. on the quantity or on the volume of the fluid flowing through the respective hydrostatic machine per time unit. The fluid throughput is in turn a function of the difference between the rotational speed of the respective primary part and the rotational speed of the corresponding secondary part as well as of the quantity of the hydraulic fluid flowing through the hydrostatic machine per revolution of the secondary part relative to the primary part. 
         [0012]    On a gear stage change, the torque transfer must be transposed from one transmission branch to the other transmission branch and the rotational speeds of the secondary parts relative to the primary parts of the respective hydrostatic machine have to be varied. The control of the gear stage change takes place via the pressure control device by means of which the hydrostatic machines can be hydraulically coupled to one another in order to bring about a pressure balance between the two hydrostatic machines—for example for a gear stage change. The pressure level of the one hydrostatic machine is raised by such a pressure balance, whereas the pressure level of the other hydrostatic machine is lowered, whereby—as described above—the degree of the coupling is increased or reduced. The consequence is an—at least part—transposition of the torque transfer from one transmission branch to the other transmission branch. So that this pressure balance influences the torque transfer via the two mechanical transmission branches in the desired sense, the connection of the respective first pressure spaces to one another and the connection of the respective second pressure spaces to one another are hydraulically separate from one another. 
         [0013]    The two hydraulic motors are in particular hydraulically coupled to one another by means of the pressure control device such that the one hydrostatic machine hydraulically drives the other hydrostatic machine. A difference in rotational speed between the primary part and the secondary part of the named other hydrostatic machine (that is of the driven hydrostatic machine) can hereby be actively brought about or at least supported. The two hydrostatic machines can preferably be directly hydraulically coupled to one another, i.e. without a direct restriction of the hydraulic fluid exchanged between the hydrostatic machines, and in particular without interposed check valves or the like. 
         [0014]    Since the respective mechanical transmission branch is only connected to a secondary part of the respective hydrostatic machine, said secondary part being able to be designed with a small radial extent, the respective mechanical transmission branch has a comparatively small moment of inertia. The respective secondary part can, for example, be a rotor. Gear stage changes can hereby be carried out particularly fast and cost-effective synchronization devices with a low torque capacity can be used in the mechanical transmission branches. 
         [0015]    The hydraulic coupling of the hydrostatic machines in addition enables an almost loss-free change in the transmission path of the driving torque since only flow resistances occur in the hydraulic system of the hydraulic coupling. Complex and/or expensive and wear-prone friction clutches and their actuator systems—such as in conventional dual clutch systems, for example—are therefore omitted. In addition, with the shift transmission in accordance with the invention, the heat output occurring in the transmission on a start-up procedure due to a high speed of rotation difference between the input shaft (engine speed) and the output shaft (equal to zero when the vehicle is stationary) can be led off by the hydraulic fluid and can be supplied, where necessary, to a cooling device. The fluid effecting the mechanical coupling hereby thus simultaneously acts as a coolant, which substantially simplifies the design of the cooling of the transmission since the coolant pump can be omitted. In addition, a plurality of driving and shifting situations can be realized by a suitable coupling of the two hydrostatic machines without additional cost-raising components being necessary. The control of the shift transmission in accordance with the invention can be based on a hydraulic control which is simple to realize. 
         [0016]    In accordance with an embodiment of the shift transmission, selectively one of the hydrostatic machines can be hydraulically blocked by means of the pressure control device in order to connect the secondary part to the primary part of the respective hydrostatic machine in a substantially rotationally fixed manner, that is without significant slip. On such a block, the fluid flow flowing through the hydrostatic machine is interrupted, whereby a hydrostatic pressure is built up in the interior of the hydrostatic machine which prevents a relative movement between the primary part and the secondary part. The hydrostatic machine is then hydraulically blocked by a type of “standing liquid column” and the secondary part is connected to the primary part in an almost rotationally fixed manner. A slight slip between the respective secondary part and the primary part can in this respect occur, for example, due to leaks. Such a slight slip can even be desired under certain circumstances, in particular to prevent a mutual mechanical deformation of the components—so-called “digging in” or hammering in” under high permanent load (for example long constant travel without gear stage change). 
         [0017]    Provision can furthermore be made that selectively one of the hydrostatic machines can be hydraulically short-circuited by means of the pressure control device in order to decouple the secondary part from the primary part of the respective machine, that is to cancel the drive connection or coupling otherwise operative between the secondary part and the primary part. A hydraulic short-circuit is to be understood as a substantially direct coupling of the two pressure spaces of the respective machine. There is thereby therefore no pressure difference or only a minimal pressure difference between the two pressure spaces of the hydrostatic machine so that the secondary part is substantially—apart from flow losses of the hydraulic fluid—freely rotatable with respect to the primary part. On a speed of rotation difference between the primary part and the secondary part, the fluid is accordingly substantially directly—and thus almost free of power loss—conveyed from one pressure space of the hydrostatic machine into its other pressure space. The coupling effect between the secondary part and the corresponding primary part is accordingly sufficiently small. 
         [0018]    This situation can, for example, be desired when the corresponding mechanical transmission branch should be decoupled, i.e. when no torque should be transferred from the input shaft to the output shaft via this transmission branch—or via one of its gear stages. 
         [0019]    The pressure control device can thus be controllable such that a driving torque transferred via the input shaft is transferred in accordance with a selected gear stage only to the first mechanical transmission branch or in accordance with another selected gear stage only to the second mechanical transmission path. Provision can, however, also be made that the driving torque—in particular for a gear stage change—is transferred or distributed at least at times onto the two mechanical transmission paths. 
         [0020]    The transfer of the driving torque in equal or unequal parts to the two mechanical transmission branches can be utilized for the production of a plurality of different transmission ratios depending on which gear stages of the two mechanical transmission branches are selected. In other words, the driving torque can be variably distributed between the mechanical transmission branch by a corresponding control of the hydraulic coupling of the hydrostatic machines by means of the pressure control device—and optionally by a corresponding control of the hydrostatic machines themselves. 
         [0021]    The secondary parts of the hydrostatic machines are advantageously connected drive-wise to the respective mechanical transmission branch without interposition of friction clutches, whereby components are saved and the control of the shift transmission is simplified. 
         [0022]    In accordance with an embodiment of the shift transmission, each of the two hydrostatic machines can selectively be operated as a hydrostatic pump or as a hydrostatic motor. This means that such a hydrostatic machine can convey hydraulic fluid from one pressure space into the other pressure space on the presence of a speed of rotation difference between the primary part and the secondary part, with the conveying quantity and conveying direction substantially depending on the speeds of rotation and the sense of rotation of the primary part and of the secondary part. In this situation, the hydrostatic machine is thus operated as a hydrostatic pump, with the fluid pressure in the named one pressure space being smaller than in the named other pressure space. The named one pressure space in this respect forms a suction region, whereas the named other pressure space forms a pressure region. 
         [0023]    In the opposite case, on the presence of a difference of the fluid pressures present in the two pressure spaces, a relative movement between the primary part and the secondary part is generated by a suitable control of valves of the hydrostatic machine. In this case, the hydrostatic machine thus acts as a hydrostatic motor which generates a mechanical torque, i.e. the secondary part is, for example, drive to make a rotational movement relative to the primary part. The pressure relationships are then converse to those described for the operation as a pump, i.e. there is a higher fluid pressure in the “suction region” than in the “pressure region”. 
         [0024]    In order to be able to operate the respective hydrostatic machine selectively as a pump or as a motor, the respective hydrostatic machine can have at least one first valve which enables a connection to the first pressure space of the respective hydrostatic machine as well as at least one second valve which enables a connection to the second pressure space. The named first valve and the named second valve can in this case be actively opened or closed by means of the named pressure control device or by means of another control device. They are preferably switch valves. 
         [0025]    In the aforesaid further development having first and second valves, the already explained hydraulic blocking of one of the hydrostatic machines can also take place by corresponding closing of the at least one first valve and/or of the at least one second valve. The already explained hydraulic short-circuiting of one of the hydrostatic machines can also take place with multi-piston machines by opening the at least one first valve and additionally the at least one second valve. 
         [0026]    In accordance with a further development of the shift transmission, one of the two hydrostatic machines is operated at least at times by means of the pressure control device as a hydraulic pump, whereas the other hydrostatic machine is simultaneously operated as a hydrostatic motor which is hydraulically driven by the one hydrostatic machine. Such a configuration can in particular be advantageous for the carrying out of a gear stage change. This form of control enables a particularly efficient division of the driving torque to the two mechanical transmission branches. The division can be varied in dependence on the requirement profile, whereby an efficient and matched driving torque transfer can be provided for a plurality of driving situations. 
         [0027]    It is preferred if a control unit is provided by means of which the pressure control device and a gear stage actuator can be controlled for a gear stage change, when a gear stage of the first mechanical transmission branch is selected, such that a gear stage of the second mechanical transmission branch is selected, while the first hydrostatic machine is hydraulically blocked and the second hydrostatic machine is hydraulically short-circuited; then the first and second hydrostatic machines are hydraulically coupled to one another and the speed of rotation of the input shaft is reduced, with a pressure balance between the two hydrostatic machines taking place and the driving torque being transmitted at least partly via the second mechanical transmission branch; and then the first and second hydrostatic machines are hydraulically decoupled from one another, with the second hydrostatic machine being hydraulically blocked and the first hydrostatic machine being hydraulically short-circuited so that the driving torque is substantially completely transferred by the second mechanical transmission branch. 
         [0028]    In an advantageous further development of this embodiment, the speed of rotation of the input shaft is controllable such that it is reduced, while the first and second hydrostatic machines are hydraulically decoupled from one another. The loads on the mechanical and hydraulic components of the shift transmission which arise are thereby reduced and a “smoother” gear stage change can be carried out. 
         [0029]    The reduction in the input shaft speed of rotation is required when “switching up”, i.e. on an increase in the gear stage. Analogously, on a “shifting down”, i.e. on a lowering of the gear stage, the speed of rotation of the input shaft is increased. 
         [0030]    In a further embodiment of the shift transmission in accordance with the invention, the geometry of the hydrostatic machines is variable such that a volume throughput of the hydraulic fluid can be controlled by the respective hydrostatic machine per revolution of the secondary part relative to the primary part. In other words, for example, the volume of pistons of a hydrostatic machine is variable and can be matched to the respective requirements. The quantity of the hydraulic fluid flowing through the hydrostatic machine can thereby be changed without the speed of rotation difference between the primary part and the secondary part having to be changed. The named volume throughput per revolution is also called the injection volume. 
         [0031]    Provision can be made with a shift transmission having variable hydrostatic machines that the geometry of the hydrostatic machines can be controlled by the pressure control device such that the volume throughput per revolution of the second hydrostatic machine is smaller before the hydraulic coupling of the hydrostatic machines than the corresponding volume throughput per revolution of the first hydrostatic machine. The geometry of the hydrostatic machines is furthermore controlled in this embodiment such that, during the hydraulic coupling of the hydrostatic machine, the volume throughput per revolution of the second hydrostatic machine is increased and the volume throughput per revolution of the first hydrostatic machine is reduced until the driving torque is primarily or substantially completely transferred via the second mechanical transmission branch. The torque transfer from one transmission branch to the other transmission branch becomes more effective and “smoother” by such a procedure. 
         [0032]    For example, in a coupled state of the hydrostatic machines, the fluid throughput capacity of the second hydrostatic machine is increased—starting from a low capacity—until the first and second hydrostatic machines have the same fluid throughput capacity. The fluid throughput capacity of the second hydrostatic machine is subsequently reduced. 
         [0033]    In accordance with a further embodiment, the first pressure space of the first hydrostatic machine can be hydraulically coupled to the second pressure space of the second hydrostatic machine by means of the pressure control and the second pressure space of the first hydrostatic machine can be coupled to the first pressure space of the second hydrostatic machine. Such a “cross-over” coupling of the hydrostatic machines enables the production of additional operating states—optionally in conjunction with a variable geometry of the hydrostatic machines. For example, with a simultaneously selected first forward gear stage and reverse gear stage, the realization of a “geared neutral” function is thereby made possible (corresponding to a ratio of infinity) and thus a “hill hold” function. 
         [0034]    The geometry of the hydrostatic machines can be fixed or set—in the case of a variable geometry—such that in a condition in which two gear stages having a transmission ratio in the same sense or in the opposite sense are selected and the first pressure space of the first hydrostatic machine is hydraulically coupled to the second pressure space of the second hydrostatic machine and the second pressure space of the first hydrostatic machine is hydraulically coupled to the first pressure space of the second hydrostatic machine different positive or negative transmission ratios can be produced between the input shaft and the output shaft. 
         [0035]    For example, a “hydraulic reverse gear” can be formed by the fixing or setting of different fluid throughput capacities of the two hydrostatic machines in that two forward gear stages are selected. In addition the production of a “geared creep” is possible. For this purpose, it is necessary—as with the “geared neutral” setting—that the first forward gear stage and the reverse gear stage are selected, with fluid throughput capacities of the two hydrostatic machines of different sizes being selected for the geared creep. 
         [0036]    In accordance with an embodiment, the respective primary part and the respective secondary part of the hydrostatic machines are rotatable. In this configuration, the hydrostatic machines act as “hydrostatic clutches” between the input shaft and the transmission branches. For example, on a blocking of one of the hydrostatic machines, a rotational movement of the rotatable primary part driven by the input shaft is transferred via the secondary part to the respective transmission branch. 
         [0037]    A constructionally particularly advantageous further development provides that the primary part of the first hydrostatic machine is rotationally fixedly connected to the primary part of the second hydrostatic machine, is in particular designed in one piece with the primary part of the second hydrostatic machine. 
         [0038]    A respective differential gear can be associated with the two mechanical transmission branches. In this respect, an input of the respective differential gear is coupled to the input shaft, whereas a first output of the respective differential gear is coupled to the secondary part of the respective hydrostatic machine. A second output of the respective differential gear is coupled to the respective mechanical transmission branch. In this embodiment, the hydrostatic machines are configured as “hydrostatic brakes” which can support the driving torque. For example, on a blocking of one of the hydrostatic machines, the first output of the differential gear is blocked. The transmission branch is thereby driven by the input shaft at a rotational speed which corresponds to the ratio of the differential gear. If, however, the design of a speed of rotation difference between the primary part and the secondary part is made possible, the torque transfer and the rotational speed ratio between the mechanical transmission branch and the input shaft can be influenced. 
         [0039]    The respective differential gear is in particular formed by a planetary transmission. Provision can furthermore be made to arrange the primary parts of the hydrostatic machines stationary. This embodiment is particularly simple in a construction respect since it is not the whole hydrostatic machine which rotates, which also simplifies its control. 
         [0040]    In accordance with a further development of the shift transmission in accordance with the invention having differential gears, the input shaft and the first and second mechanical transmission branches are permanently coupled to one another, with—as explained above—the driving torque transferred via this type of coupling likewise being variable and depending on the operating state of the hydrostatic machines. 
         [0041]    It is furthermore preferred if the named primary part is a housing of the hydrostatic machine. The secondary part can be formed by a rotor. Alternatively to this, the named primary part can, if it is arranged rotatably, be a further rotor of the respective hydrostatic machine. 
         [0042]    It has proved to be particularly efficient if at least one of the two hydrostatic machines can be connected to at least one further component of a hydraulic system. For example, the transferred torque can be determined in a simple manner by a pressure measurement. In addition, in specific driving states, hydraulic fluid can be used by a connection to the hydrostatic machines for the actuation of further vehicle control components—for example of an all-wheel clutch. 
         [0043]    A connection line can be associated with the hydrostatic machines, with a controllable restrictor valve being arranged in its extent to restrict the fluid throughput of the respective hydrostatic machine. In other words, the fluid throughput can be influenced for specific driving conditions by the controllable restrictor, whereby the torque transmitted via the corresponding hydrostatic machine can be controlled. This in particular simplifies the control of the torque transfer from the input shaft to the mechanical transmission branch in a start-up situation. 
         [0044]    A common connection line and a common restrictor valve are preferably associated with the hydrostatic machines. A cooling device for the cooling of the hydraulic fluid can be arranged in the extent of the connection line, whereby the fluid flowing through the restrictor can be cooled in an efficient manner. Particularly with large speed of rotation differences between the primary part and the secondary part—as on a start-up procedure, for instance—a substantial quantity of waste heat is produced which can thus be efficiently led off. 
         [0045]    Further areas of applicability will become apparent from the description provided herein. The description and specific examples in this summary are intended for purposes of illustration only and are not intended to limit the scope of the present disclosure. 
     
    
     
       DRAWINGS 
         [0046]    The drawings describe herein are for illustrative purposes only of the selected embodiments and not all possible implementations and are not intended to limit the scope of the present disclosure. 
           [0047]      FIG. 1  is a schematic representation of an embodiment of the shift transmission in accordance with the invention; 
           [0048]      FIG. 2  is a section through a radial piston machine; 
           [0049]      FIGS. 3 to 5  illustrate different embodiments of a pressure control device of an embodiment of the shift transmission in accordance with the invention; 
           [0050]      FIGS. 6 to 8  are schematic representations of different further embodiments of the shift transmission in accordance with the invention; and 
           [0051]      FIG. 9  discloses a planetary gear arrangement which serves for the coupling of the input shaft with the hydrostatic machines and the mechanical transmission branches. 
       
    
    
     DETAILED DESCRIPTION 
       [0052]    Exemplary embodiments of the present invention will now be more fully described with reference to the accompanying drawings. 
         [0053]      FIG. 1  shows an embodiment of a shift transmission  10  in accordance with the invention. The left hand side of the shift transmission  10  facing a drive unit, not shown, of a vehicle includes an input shaft  12  which is driven to make a rotational movement by the drive unit. Rotational irregularities are conducted from the drive unit—for example an internal combustion engine—into a powertrain of the vehicle including the shift transmission  10  and result in the creation of rotational vibrations. The input shaft  12  has a torsion damper  14  to reduce the rotational vibrations. 
         [0054]    The input shaft  12  is connected at the transmission side to a first and a second hydrostatic machine  18 ,  20  which have a common housing  16 . The housing  16  is rotationally fixedly coupled to the input shaft  12 . 
         [0055]    The machines  18 ,  20  each have a rotor  22  and  24  respectively (see also  FIG. 2 ), with the rotor  22  being rotationally fixedly connected to a first mechanical transmission branch  26 , whereas the rotor  24  is rotationally fixedly connected to a second mechanical transmission branch  28 . 
         [0056]    The first transmission branch  26  includes a hollow shaft  30  which is permanently rotationally fixedly connected to the transmission gears G 1  and G 3 . Further transmission gears G 5  and G 7  can be selectively connected rotationally fixedly to the hollow shaft  30  by a synchronization device  32 . 
         [0057]    In an analog manner, the second mechanical transmission branch  28  includes a transmission shaft  34  which is in permanent rotationally fixed communication with a transmission gear G 2  and which can selectively be coupled to a transmission gear G 4  via a synchronization device  32 . In addition, a gear r is fastened to the transmission shaft  34  and is in engagement with a transmission gear R by which a reverse gear can be formed. 
         [0058]    The shift transmission  10  furthermore includes a back gear shaft  36  which has eight gears  38 . Of the eight gears  38 , the middle four gears  38  can be selectively rotationally fixedly coupled to the back gear shaft  36  by synchronization devices  32 . The remaining four gears are permanently rotationally fixedly coupled to the back gear shaft  36 . 
         [0059]    By actuation of a respective gear stage actuator (not shown), the synchronization devices  32  can be axially displaced in order to form seven forward gear stages (in accordance with the gears G 1  to G 6 ) and one reverse gear (R) in a known manner. For the formation of the first gear stage, the left hand synchronization device  32  of the back gear shaft  36  is brought into engagement with the gear  38  of the back gear shaft  36  adjacent to the right hand side so that a rotational movement of the hollow shaft  30  can be transmitted via the transmission gear G 1  to the back gear shaft  36  and finally via the transmission gear G 6  to an output shaft  40  of the shift transmission  10  and thus to further elements of the powertrain (not shown) of the vehicle. The further gear stages of the shift transmission  10  are formed in an analog manner. 
         [0060]    It will be explained in the following how a driving torque of the input shaft  12  is transferred in a suitable manner with the shift transmission  10  to the hollow shaft  30  and/or to the transmission shaft  34 . 
         [0061]    If, for example, an even gear stage (second, fourth or sixth gear) or the reverse gear is selected, the torque of the input shaft  12  has to be transferred to the transmission shaft  34 . If an odd gear stage has been selected, the transfer of the driving torque to the hollow shaft  30  is necessary. If a change of the gear stage should be carried out, a change of the transmission path of the torque must also take place. In this respect, a portion of the driving torque is transferred via both mechanical transmission branches  26 ,  28  at times, with the respective transferred portion of the driving torque changing during the gear stage change. Such a gear stage change should also be possible and should run as smoothly as possible under load so that the driving comfort is not reduced by jerky movements of the vehicle or similar negative accompanying phenomena. 
         [0062]    This is achieved by the use of the two hydrostatic machines  18 ,  20 . The rotor  24  can, for example, be blocked with respect to the housing  16  by the control of the machines  18 ,  20 , whereas the rotor  22  associated with the transmission branch  26  is decoupled from the housing  16 . In this case, the torque of the input shaft  12  is transferred completely to the transmission shaft  34  via the machine  20 . It is, however, also possible that the hydrostatic machines  18 ,  20  are controlled such that the rotors  20 ,  24  are only partly coupled to the rotational movement of the housing  16 . No friction clutches are thus required to be able to carry out and vary the torque transfer to the mechanical transmission branches  26 ,  28 . This division only takes place via the machines  18 ,  20  which are substantially identical in function. 
         [0063]    A machine type suitable for use in the shift transmission  10  is represented, for example, by hydrostatic radial piston machines. The function of a radial piston machine will be explained with reference to  FIG. 2  in the following which shows a section through a radial piston machine  20 . The radial piston machine  20  shown can be operated both as a pump and as a motor. In other words, it can be used, on the one hand, for the conveying of a hydraulic fluid; on the other hand, it can generate a relative rotational movement between the housing  16  and the rotor  24  by controlled pressure application. 
         [0064]    The radial piston machine  20  shown includes the rotor  24  which has a circular outline in the region of the machine  20 , with the center  44  of the circular shape being offset with respect to the common axis of rotation  46  of the housing  16  and of the rotor  24  or of the associated transmission shaft  34 . In other words, the rotor  24  is an eccentric element. The rotor  24  is in communication with five pistons  48  which each have a piston space  50 . On a rotation of the rotor  24  relative to the housing  16 , the volumes of the piston spaces  50  are alternately increased and decreased in size. In other words, a hydraulic fluid which first flows in through a valve  52  is subsequently expelled again through a further valve  52 ′ of the respective piston  48  by the rotational movement of the rotor  22  relative to the housing  16 . A hydraulic fluid is thus conveyed from a first pressure space (not shown) in communication with the valve  52  to a second pressure space (not shown) which is in communication with the valve  52 ′. 
         [0065]    If the radial piston machine  20  is operated as a pump, hydraulic liquid is initially sucked into the piston space  50  of a cylinder  51  a of the radial piston machine  20  in the state shown in  FIG. 2  on a rotation of the rotor  24  counter clockwise since the piston space  50  initially has a minimal volume. The pistons  48  of the cylinders  51   b  and  51   c  are also in the suction phase. If a maximum volume of the respective piston space  50  has been reached, the volume of the piston space  50  is now reduced again, that is the fluid pressure is increased, due to the effect of the rotation of the rotor  24 . From a specific rotational position of the rotor  24  or from a specific threshold of the fluid pressure onward, the valve  52 ′ is opened and the hydraulic fluid is output into the pressure space, not shown. 
         [0066]      FIG. 2  was described by way of example under the assumption that the housing  16  is not rotatably journalled. It can, however, easily be seen that the conveyed quantity of the hydraulic fluid only depends on the geometry of the piston spaces  50  and on a speed of rotation difference between the housing  16  and the rotor  24 . In other words, no hydraulic fluid is conveyed when the housing  16  and the rotor  24  rotate at the same speed. 
         [0067]    If the radial piston machine  20  is operated as a motor, a rotational movement is produced, or at least supported, by a pressure difference in the pressure spaces, not shown, with the above-named functional principle applying in an analog manner. However, the pressurized hydraulic fluid must then be fed into the respective piston space  50  by a suitable control of the respective valve  52  of the cylinders  51   a - e  on a suitable position of the rotor  24 . On pressure reduction, the volume of the piston space  50  is increased, whereby the rotor  24  has a torque applied by the piston  48 . Subsequently, the valve  52 ′ is opened to allow the hydraulic fluid to escape at a now lower pressure. 
         [0068]    It must still be noted with respect to  FIG. 2  that a radial piston machine  18  of substantially the same type can be arranged axially offset to the radial piston machine  20  shown, with the two radial piston machines  18 ,  20  in particular being able to have a common housing  16  (cf.  FIG. 1 ). Generally, other types of hydrostatic machines  18 ,  20  can also be used. In the use of the hydrostatic machine  20  described here, not only the conveying of a hydraulic fluid or the drive of a shaft is of central importance, but also a controlled coupling of the housing  16  with the rotors  20 ,  24 . This can be realized in that the flow of hydraulic fluid through the hydrostatic machine  18 ,  20  or the pressure of the hydraulic fluid is controlled. If the hydrostatic machine  20  can namely not output any hydraulic fluid through the valve  52 , the rotor  24  can no longer rotate with respect to the housing  16 . The coupling is cancelled in that the throughput of the hydraulic fluid is permitted again. The distribution of the driving torque of the input shaft  12  in accordance with  FIG. 1  transferred via the individual mechanical transmission branches  26 ,  28  is thus substantially based on a variation of the pressure of the hydraulic fluid. A schematic view of an embodiment of a pressure control  53  is shown in  FIG. 3 . 
         [0069]      FIG. 3  shows the machines  18 ,  20 . The machines  18 ,  20  are each connected to pressure lines  54  and  54 ′ and  54   a  and  54   a ′ respectively. The hydrostatic machines  18 ,  20  can be hydraulically coupled in that a connection is established between the pressure lines  54 ,  54 ′ and the pressure lines  54   a,    54   a ′. This takes place by two valves V 1 , V 2 . The valve V 1  is here a 4/3 way valve and the valve V 2  is a 4/2 way valve. 
         [0070]    The valve V 1  has three switch states. In a first switch state (lowest section of the valve V 1  in accordance with  FIG. 3 ), the pressure lines  54  and  54 ′ of the machine  18  are blocked, whereas the pressure lines  54   a  and  54   a ′ of the machine  20  are connected to one another. In the second switch state of the valve V 1  (shown in  FIG. 3 ), the pressure line  54 ′ is connected to the pressure line  54 ′ and the pressure line  54  is connected to the pressure line  54   a.  The third switch state is the converse of the first state, i.e. the pressure lines  54   a  and  54   a ′ are blocked, while the pressure lines  54  and  54 ′ are connected to one another (topmost section of the valve V 1  in accordance with  FIG. 3 ). 
         [0071]    The valve V 2  has two switch states, with the second switch state of the valve V 2  in particular being of importance in the aforesaid second switch stage of the valve V 1 . A “cross-over” connection or coupling inversion of the hydrostatic machines  18 ,  20  can then be established by the valve V 2 . In this case, the pressure line  54  is in communication with the pressure line  54   a ′, whereas the pressure line  54 ′ is in communication with the pressure line  54   a.  The first switch state of the valve V 2  does not produce this effect, but rather only serves for the “normal” coupling of the hydrostatic machines  18 ,  20 . 
         [0072]    In other words, a block or an idling of one of the hydrostatic machines  18 ,  20  can be effected by the valves V 1 , V 2 , with—as already described above—the respective mechanical transmission branch  26 ,  28  being decoupled from the input shaft  12  on an idling of the hydrostatic machines  18 ,  20 , i.e. on a short-circuit of the pressure lines  54 ,  54 ′ or  54   a,    54   a ′ respectively associated with the corresponding hydrostatic machine  18 ,  20 . On a block of the pressure lines  54 ,  54 ′ or  54   a,    54   a ′ respectively, in contrast, a substantially slip-free coupling of the drive shaft  12  to the corresponding mechanical transmission branch  26 ,  28  is brought about. A pressure balance—and thus a torque transfer—can be established between the hydrostatic machines  18 ,  20  by a hydraulic coupling by the second switch position of the valve V 1 , which is significant within the framework of a gear stage change, for example, as will be described in the following. 
         [0073]    The hydraulic system described above for the hydraulic coupling of the hydrostatic machines  18 ,  20  is in communication via a supply line  56  and an outflow line  58  as well as a check valve  59  with a hydraulic control unit (HCU)  60 . Check valves  62  in the pressure lines  54 ,  54 ′,  54   a,    54   a ′ ensure that no hydraulic fluid can flow back into the supply line  56  or no hydraulic fluid can flow back out of the outflow line  58  into the aforesaid part of the hydraulic coupling system. The supply line  56  and the outflow line  58  have rotary leadthroughs  64 . The rotary leadthroughs  64  are necessary since the machines  18 ,  20 , the pressure lines  54 ,  54 ′ and  54   a ,  54   a ′ respectively associated with them and the valves V 1 , V 2  rotate (rotation region Ro above the dashed line), whereas the remaining components, still to be described in the following in part, of the control  53  are arranged stationary (stationary region S beneath the dashed line). 
         [0074]    Control lines  66  can be pressurized by the hydraulic control unit  64  to control the valves V 1  and V 2 , on the one hand, and also a valve V 5 , on the other hand—whose function will be explained in the following—by means of a control pressure. 
         [0075]    The hydraulic control unit  60  is supplied with pressurized hydraulic fluid through a pump  68  in communication with a motor M, with the motor M being electrically controlled by a transmission control unit (TCU)  70 . The pump  68  takes the hydraulic fluid via a hydraulic fluid filter  71  from a sump  72  which is also in communication with the hydraulic control unit  60 . 
         [0076]    If, for example, the first gear stage is selected and if the driving torque of the drive unit of the vehicle should therefore be transmitted completely via the first mechanical transmission path  26 , the rotor  22  of the first hydrostatic machine  18  rotationally fixedly connected to the hollow shaft  30  must be blocked with respect to the housing  16  rotationally fixedly connected to the input shaft  12  (cf.  FIG. 1 ). For this purpose, the valve V 1  shown in  FIG. 3  must be in the explained first switch state. The hydrostatic machine  18  is then blocked due to the blocking of the pressure lines  54 ,  54 ′ so that the rotor  22  rotates together with the housing  16 . The hydrostatic machine  20  is, in contrast, in a short-circuited state so that its two pressure spaces are substantially in direct communication with one another. With a speed of rotation difference between the rotor  24  and the housing  16 , only hydraulic fluid is thus circulated and is conveyed substantially loss-free from one pressure space into the other, which corresponds to an idling of the hydrostatic machine  20 . 
         [0077]    Starting from this state, the operation of the shift transmission  10  should now be described with reference to  FIGS. 1 to 3  by way of example with respect to a change from the first gear stage into the second gear stage. 
         [0078]    Since the second hydrostatic machine  20  is short-circuited, the new gear stage can be selected by means of the associated synchronization device  32 , i.e. the transmission gear G 2  of the second mechanical transmission branch  28  is rotationally fixedly coupled to the transmission shaft  34 . Due to the transmission ratio of the second gear step—which is lower in comparison with the transmission ratio of the first gear stage—there is a speed of rotation difference between the rotational speed of the input shaft and the rotational speed of the second mechanical transmission branch  28 , with the hydrostatic machine  20  acting as a hydrostatic pump. No driving torque is yet transferred to the mechanical transmission branch  28  at this moment in time due to the short-circuit of the lines  54   a  and  54   a′.    
         [0079]    Then a takeover of a portion of the driving torque by the second transmission branch  28  is initiated in that the valve V 2  is brought into the second switch state shown in  FIG. 3 . A hydraulic coupling of the two hydrostatic machines  18 ,  20  is thereby established. The hydraulic fluid conveyed by the hydrostatic machine  20  acting as a pump is now supplied to the machine  18  operated as a motor by an active control of the valves  52 ,  52 ′. There is initially not yet any speed of rotation difference between the housing  16  and the rotor  22  of the machine  18 . 
         [0080]    The fluid conveyed by the large pump capacity of the hydrostatic machine  20  now, however, drives the hydrostatic machine  18 —with a corresponding actuation of the valves  52 ,  52 ′—whereby a lowering of the rotational speed of the input shaft and thus of the drive unit of the vehicle is supported. The lowering of the rotational speed of the drive unit is also carried out actively simultaneously. The speed of rotation difference between the housing  16  and the rotor  24  of the hydrostatic machine  20  is reduced by the lowering of the rotational speed of the input shaft  12  since the rotational speeds of the mechanical transmission branches  26 ,  28  are constant during the total shift procedure due to the substantially unchanging vehicle speed. This has the consequence of a lowering of the conveying capacity of the hydrostatic machine  20 . In contrast to this, the speed of rotation difference between the housing and the rotor  22  of the hydrostatic machine  18  increases, whereby the drive performance of the hydrostatic machine  18  likewise falls 
         [0081]    The falling of the capacities of the hydrostatic machines  18 ,  20 , on the one hand, results in an increase of the torque transferred via the second transmission branch  28 ; on the other hand, the torque transmitted via the first transmission branch  26  reduces. This procedure continues until a pressure balance is established between the hydrostatic machines  18 ,  20  and a balanced state is adopted in which the driving torque is transferred via the first mechanical transmission branch  26 , in one part, and via the second mechanical transmission branch  28 , in the other part. If the hydrostatic machines  18 ,  20  are substantially identical, i.e. have substantially the same piston space geometries, an equal division of the torque transferred via the individual transmission branches  26 ,  28  is adopted in the balanced state. 
         [0082]    Subsequently, the machines  18 ,  20  are again hydraulically decoupled from one another in that the valve V 1  is brought into the explained third switch state, whereby the hydrostatic machine  18  is short-circuited and the hydrostatic machine  20  is hydraulically blocked. To avoid warping of the mechanical components of the shift transmission  10 , the switching of the valve V 1  is accompanied by an active rotational speed reduction of the input shaft  12  until the input shaft  12  and the second transmission branch  28  have the same rotational speed. The driving torque is now transferred substantially completely by the second mechanical transmission branch  28  with the blocking of the pressure lines  54   a,    54   a ′. The change from the first gear stage into the second gear stage is thus concluded. 
         [0083]    Gear stage changes between other gear stages take place in an analog manner. A gear stage change from a higher gear stage into a lower gear stage takes place substantially in the reverse order. 
         [0084]    The shift transmission  10  makes possible—as described above—a type of gear stage change which can be controlled simply, with the gear stage change also being able to take place under load. No substantial power losses occur during the gear stage change due to the pump/motor configuration of the hydrostatic machines  18 ,  20 . The hydrostatic machines  18 ,  20  rather support the gear stage changes in an advantageous manner, whereby it can be designed particularly efficiently. In addition, it becomes clear from the above descriptions that friction clutches can be completely dispensed with. Only the constructionally simple valves V 1  and V 2  and the hydrostatic machines  18 ,  20  have to be controlled in a suitable manner. 
         [0085]    The use of hydrostatic machines  18 ,  20  for the coupling of the input shaft  12  and of the mechanical transmission branches  26 ,  28  additionally enables a plurality of advantageous further developments. 
         [0086]    As already noted above, the outflow line  58  has the valve V 3 . This is generally closed during the above-described procedures. In addition, a restrictor valve D which can be regulated by the transmission control unit  70  and a cooling device  74  are arranged in the outflow line  58 . These components can be utilized, for example, on a start-up of the vehicle. In this respect, the driving torque should be transferred via the first gear stage so that the first transmission branch  26  is selected and the corresponding hydrostatic machine  18  is short-circuited. The second transmission branch  28  is not selected. 
         [0087]    In this situation, the input shaft  12 —and thus the housing  16  of the hydrostatic machine  18 —rotates very fast (rotational speed of the drive unit), while the selected transmission branch  26  does not show any rotation since the vehicle is stationary. A high speed of rotation difference between the housing  16  and the rotor  22  is thus present, which brings about a large conveying capacity of the hydraulically short-circuited machine  18  and results in an increased heat development there. In order gradually to increase the degree of coupling between the input shaft  12  and the selected transmission branch  26 , the valve V 3  is opened, with the regulable restrictor valve D being in an opened position. Expediently, the pressure lines  54 ,  54 ′ are additionally blocked (aforesaid first position of the valve V 1 ). 
         [0088]    The counter pressure against which the hydrostatic machine  18  has to work is increased by a gradual closing of the restrictor valve D. This counter-pressure acting against the pump capacity of the machine  18  has the result that the coupling of the rotor  22  with the housing  16  is creased. An increasing portion of the driving torque is therefore transferred to the first transmission branch  26  by the closing of the restrictor valve D and the vehicle starts up. 
         [0089]    In other words, the counter-pressure acting against the pump power can be controlled via an intervention into the conveyed volume of the hydraulic fluid, which results in a coupling of the rotor  22  with the housing  16  since the driving torque transferred by the input shaft  12  to the mechanical transmission branches  26  is directly proportional to the fluid pressure which is effectively generated due to the conveying capacity of the hydraulic machine  18 , on the one hand, and the intervention by means of the pressure control  53 , on the other hand. 
         [0090]    Start-up states can thus be realized in a simple manner by the provision of the valve V 3  and of the regulable restrictor valve D without an additional start-up element being required. In addition, the heat arising in the machine  18  can be led away in an efficient manner by the cooling device  74 . 
         [0091]    The restricted hydraulic fluid can be supplied back to the hydrostatic machines  18 ,  20  via the supply line  56  in communication with the outflow line  58 . The hydraulic control unit  60  can moreover balance any fluid losses—for example at the rotary leadthroughs  64 —by hydraulic fluid conveyed from the sump  72  by means of the pump  68 . 
         [0092]    Instead of the switch valve V 3  and the restrictor valve D, a single regulable valve can also be provided (proportional valve, restrictor valve), as is shown in  FIG. 4 . 
         [0093]      FIG. 4  shows a further embodiment of the pressure control device  53 . 
         [0094]    Instead of the valve V 1  with three possible switch states, two 2/4 way valves V 1 ′ and V 1 ″ are provided which each permit two switch states, namely a switch state for the connection of the pressure lines  54 ′ and  54   a ′ or  54  and  54   a  respectively—short circuit of one of the hydrostatic machines  18 ,  20 —and a switch state for the block of the other hydrostatic machine  20  and  18  respectively. The valves V 1 ′, V 1 ″ and V 2  are designed such that, on a failure of the hydraulic control unit  60  and on a subsequent drop in the control pressure in the control lines  66 , the hydrostatic machines  18 ,  20  are automatically coupled so that, for example, an unintentional simultaneous block of both hydrostatic machines  18 ,  20 , which is damaging for the components of the shift transmission, can be precluded. In addition, such valves V 1 ′, V 2 ″, V 2  having two positions can be controlled in a simple manner. 
         [0095]    In contrast to the embodiment of the pressure control  53  shown in  FIG. 3 , the embodiment of  FIG. 4  does not have a valve V 3  for the separation of the outflow lone  58  from the hydraulic system to the coupling of the hydrostatic machines  18 ,  20 . This function is satisfied here by the restrictor valve D which is hydraulically controlled by the control line  66 . The embodiment of  FIG. 4  is structured more simply in an advantageous manner by the omission of the valve V 3  and of an electrical control line for the control of the restrictor valve D by the transmission control unit  70 —see  FIG. 3 , dashed line. In addition, the restrictor valve D is arranged in the rotating region Ro, whereby the advantage results that the rotary leadthrough  64  in the outflow line  58  is arranged behind the restrictor valve D in the flow direction of the hydraulic fluid. The rotary leadthrough  64  is therefore no longer part of the part of the pressure control  53  to which high pressure is applied. Leak losses are thereby minimized and the rotary leadthrough  64  can be designed in a less complex and/or expensive manner. An automatic opening of the restrictor valve D on the drop of the control pressure can be provided to bring the vehicle into a state in which the drive unit is substantially completely decoupled from the transmission branches  26 ,  28 . 
         [0096]    In the above descriptions, only the “parallel” position of the respective valve V 2  is looked at which is shown in  FIGS. 3 and 4  and which results in a connection of the pressure lines  54 ′ and  54   a ′ or  54  and  54   a  respectively. In specific cases, a “cross-over” coupling of the hydrostatic machines  18 ,  20  can also be advantageous (second switch state of V 2 ). 
         [0097]    If, for example, a first gear stage and the reverse gear are simultaneously selected, a torque is admittedly provided via both transmission branches  26 ,  28 . However, the two mechanical transmission branches  26 ,  28  do not rotate; the vehicle is stationary. A rolling away of the vehicle when stationary or at a hill can thereby be prevented, for example (“geared neutral” or “hill hold” function). In this state, the hydrostatic machines  18 ,  20  are in the above-described balanced state in which the pressure balance has already taken place. 
         [0098]    Provision can furthermore be made that the hydrostatic machines  18 ,  20  have a variable geometry—as variable hydrostatic machines  18 ′,  20 ′—with the piston spaces  50  of the cylinders  51   a - e  of the variable hydrostatic machines  18 , 20  being adjustable, for example, by means of wobble plates so that the throughput of the hydraulic fluid per revolution of the rotor  22  or  24  can be variably controlled—both in a pump operation and in a motor operation. Other hydrostatic machine types then the machine type discussed above with radial pistons can also do this. 
         [0099]    Such hydrostatic machines  18 ′,  20 ′, with variable geometries make it possible that a “geared creep” can be realized with a “cross-over” coupling. 
         [0100]    For this purpose, for example, the first gear stage and the reverse gear stage are selected and the hydrostatic machine  18 ′,  20 ′ acting as a pump has a larger conveying capacity than the machine  20 ′ or  18 ′ operated as a motor. With the balanced state being adopted in the “cross-over” configuration, torque is transmitted via both transmission branches  26 ,  28 , with these then rotating in opposite senses. In sum, a small propulsion of the vehicle is produced and a transmission ratio can be set which is lower in amount than the transmission ratio of the smallest gear stage (G 1  or R) of the mechanical transmission branches  26 ,  28 . 
         [0101]    If instead of the reverse gear stage a forward gear stage—for example the second gear stage—is adopted under otherwise the same conditions, i.e. differently large conveying capacities of the hydrostatic machines  18 ′,  20 ′, a drive of the vehicle likewise results due to the transmission ratios of different amounts of the selected gear stages  26 ,  28 , said drive, however, being directed in the opposite direction—compared with the above-described case of the “geared creep”. In other words, a “hydraulic reverse gear” can thus be realized. The torques transferred via the two transmission branches  26 ,  28  in this respect have a different sign. 
         [0102]    It can therefore very generally be stated that balanced states can be generated with the help of variable hydrostatic machines  18 ′,  20 ′ and a suitable combination of gear stages on a hydraulic coupling of the hydrostatic machines  18 ′,  20 ′, said balanced states ultimately having the effect of additional transmission ratios. Such shift transmissions can therefore be used very flexibly and versatilely. The use of variable hydrostatic machines  18 ′,  20 ′ for the carrying out of a gear stage change will be explained in the following with reference to  FIG. 5 . 
         [0103]    Such balanced states can, however, also be generated with a fixed geometry of the hydrostatic machines, with the adopted state then corresponding to the fixedly set respective volume throughput per revolution of the hydrostatic machines. 
         [0104]      FIG. 5  shows an embodiment of the pressure control  53  of a variant of the shift transmission  10  with variable hydrostatic machines  18 ′,  20 ′. A gear stage change in this respect takes place in a substantially analog manner, as described above with reference to  FIG. 3 . The variable hydrostatic machine  20 ′ is, however, configured before the hydraulic coupling of the hydrostatic machines  18 ′,  20 ′ such that its fluid volume throughput per revolution, that is its volume displacement per revolution, is smaller than the corresponding fluid volume throughput per revolution of the hydrostatic machine  18 ′ if this were not blocked. The fluid volume throughput per revolution of the hydrostatic machine  20 ′, which is operated as a pump in the starting state, is in particular very small so that the quantity of the hydraulic fluid circulated in the idling state is small. 
         [0105]    After the second gear stage has been selected and after the two hydrostatic machines  18 ′,  20 ′—switch states of the valves V 1 ′, V 1 ″ and V 2  as shown in FIG.  3 —have been hydraulically coupled to one another, the fluid volume throughput per revolution of the hydrostatic machine  20 ′ is generally raised, whereas the rotational speed of the input shaft is reduced. During the conveying capacity increase of the hydrostatic machine  20 ′, the fluid volume throughput per revolution of the hydrostatic machine  18 ′ operated as a motor in this state remains constant. In this situation, an increasing torque transfer takes place via the second transmission branch  28  associated with the hydrostatic machine  20 ′, while the torque transferred via the first transmission branch  26  falls by the same amount. With amounts of the torques transferred via the two mechanical transmission branches  26 ,  28  of equal size, the balanced state described with reference to  FIG. 3  is substantially present. 
         [0106]    On a further lowering of the rotational speed, the fluid volume throughput per revolution of the hydrostatic machine  18 ′ is reduced in the continued coupled state of the hydrostatic machines  18 ′,  20 ′, while the fluid volume throughput per revolution of the hydrostatic machine  20 ′ remains constant or is increased even further. More and more torque is thereby transferred via the second mechanical transmission branch  28 . A substantially complete torque transfer from the first transmission branch  26  to the second transmission branch  28  is reached when the drive rotational speed has reached the rotational speed level of the second transmission branch  28 . To conclude the shift procedure, the hydrostatic machine  20 ′ is then blocked by an actuation of the valve V  1 ″ and the hydrostatic machine  18 ′ is simultaneously short-circuited by the valve V 1 ′. 
         [0107]    The respective fluid volume throughput per revolution, i.e. the respective geometry of the two hydrostatic machines  18 ′,  20 ′ can also be varied simultaneously or with time overlap on this gear stage change. 
         [0108]    The above-described variant of the shift transmission  10  having variable hydrostatic machines  18 ′,  20 ′ allows even smoother gear stage changes. In addition, the above-described concepts can be realized with respect to a geared creep and to a hydraulic reverse gear as well as with respect to a plurality of intermediate gear changes. 
         [0109]    The embodiment of the pressure control  53  shown in  FIG. 5  has no outflow line  58 . Consequently, there is also no valve V 3  arranged in the extent of the outflow line  58 , no regulable restrictor D and no cooling device  74 . These components can, however, generally also be integrated in the embodiment shown in  FIG. 5 . 
         [0110]    All discussed embodiments of the pressure control  53  can be in communication with further components of a hydraulic system. For example, the pressure lines  54 ,  54 ′,  54   a,    54   a ′ can be connectable via a switch-in valve (not shown) to an all-wheel drive clutch (AWD clutch) to actuate it. An effective monitoring of the pressure state of the hydrostatic machines  18 ,  18 ′,  20 ,  20 ′ is also possible by such a connection. 
         [0111]    It must still be noted with respect to the above-explained respective pressure control  53  that the switch valves (V 1 , V 2 , V 3 ) can have suitable control edges to effect soft transitions between the different switch states. 
         [0112]    In addition, a “fail-safe” function is preferably realized. As can be seen from the arrangement of respective compression springs in accordance with  FIGS. 3 to 5 , the valves (V 1 , V 1 ′, V″, V 2  and V 3 ) of the pressure control  53  are automatically brought into an open position in the case of an operational failure (switching the hydraulic control unit pressure free) to switch the transmission load free. 
         [0113]      FIG. 6  shows that the shift transmission  10  can also be combined with a hybrid drive in a simple manner. The part of the shift transmission  10  from the housing  16  to the right corresponds to the embodiment which was discussed above with reference to  FIG. 1 . To the left of this, the torsion damper  14  is in turn provided which is, however, combined with a clutch  78 . The shift transmission  10  can thereby be separated from the drive unit (not shown) so that a driving torque can be generated on the housing  16  by an electrical drive unit  80 . The electrical drive unit  80  can also be used as a generator on braking for the generation of electrical energy. 
         [0114]      FIG. 7  shows a further embodiment of the shift transmission  10  which again corresponds in large parts to the embodiment shown in  FIG. 1 . The rotor of the electrical drive unit or of the generator  80  is here rotationally fixedly coupled to the hollow shaft  30  of the first transmission branch  26 . In this case, the clutch  78  can be dispensed with. 
         [0115]      FIG. 8  shows a further embodiment of the shift transmission  10 , wherein the hydrostatic machines  18 ,  18 ′,  20 ,  20 ′ are arranged between mechanical transmission branches  26 ,  28 . This embodiment can also be combined with a hybrid drive in a simple manner. 
         [0116]      FIG. 9  shows a further application possibility of the hydrostatic machines  18 ,  18 ′,  20 ,  20 ′ in accordance with the idea underlying the invention. The hydrostatic machines  18 ,  18 ′,  20 ,  20 ′ here do not have any common housing rotationally fixedly connected to the input shaft  12 . The respective housing  16  of the pumps  18 ,  18 ′,  20 ,  20 ′ is instead fastened in a stationary manner, that is it does not rotate. The driving torque of the input shaft  12  is transferred via planetary transmissions  82  to the mechanical transmission branches  26 ,  28 . A sun gear  84  of the respective planetary transmission  82  is here rotationally fixedly connected to the rotor  22  or  24  of the associated pump  18 ,  18 ′,  20 ,  20 ′. The mechanical transmission branches  26 ,  28  are rotationally fixedly coupled to a respective planetary carrier  86  at which planetary gears  88  are rotationally journalled. The driving torque of the input shaft  12  is transferred to a respective annular gear  90 . The planetary gears  88  mesh with the respective sun gear  84  and with the respective annulus gear  90 . The planetary transmissions  82  can naturally also be configured differently than described here by way of example. 
         [0117]    In this embodiment, the rotors  22 ,  24  act so-to-say as “brakes” with which the respective sun gears  84  can be braked or held firmly. The planetary transmissions  82  thus act as differential gears for the transfer of a driving torque of the input shaft  12 . If one of the pumps  18 ,  18 ′,  20 ,  20 ′ is hydraulically blocked and if the other is hydraulically short-circuited, the driving torque of the input shaft  12  is transferred completely via the mechanical transmission branch  26  and  28  respectively associated with the blocked pump  18 ,  18 ′,  20 ,  20 ′. This embodiment can likewise be controlled by the pressure control  53  which was described above with reference to  FIGS. 3 to 5 . However, advantages result in a construction respect since the housings  16  do not rotate, which, for example, simplifies the guidance of the control lines  68 . 
         [0118]    The foregoing description of the exemplary embodiments has been provided for purposes of illustration and description. It is not intended to be exhaustive or to limit the inventions. Individual elements or features of a particular embodiment are generally not limited to that particular embodiment but, where applicable, are interchangeable and can be used in a selected embodiment even if not specifically shown or described. The same may also be varied in many ways. Such variations are not intended to be regarded as a departure from the invention, and all modifications are intended to be included within the scope of the invention. 
       REFERENCE NUMERAL LIST 
       [0119]    shift transmission
     12  input shaft     14  torsion damper     16  housing     18 ,  18 ′,  20 ,  20 ′ hydrostatic machine     22 ,  24  rotor     26 ,  28  mechanical transmission branch     30  hollow shaft     32  synchronizing device     34  transmission shaft   G 1 -G 7 , R transmission gears   r,  38  gear     36  back gear shaft     40  output shaft     44  center of the rotor     46  axis of rotation     48  piston     50  piston space     51   a - e  cylinder     52 ,  52 ′ valve     53  pressure control     54 ,  54 ′,  54   a,    54   a ′ pressure line     56  supply line     58  outflow line     59  check valve     60  hydraulic control unit     62  check valve     64  rotary leadthrough     66  control line     68  pump     70  transmission control unit     71  hydraulic fluid filter     72  sump     74  cooling device     78  clutch     80  electrical drive unit     82  planetary transmission     84  sun gear     86  planet carrier     88  planetary gear     90  annulus gear   V 1 , V 1 ′, V 1 ″,   V 2 , V 3 , valve   D restrictor valve   M motor   Ro rotating region   S stationary region