DE29513686U1 - Hydrostatic-mechanical transmission with power split - Google Patents

Description

25.08.95
G128N3
Hydrostatic-mechanical transmission with power split

The invention relates to a continuously variable transmission with hydrostatic power split as described in the preambles of the claims. The object of the invention is to reduce the number of components to a minimum, to reduce installation space and costs, to improve noise behavior, efficiency and comfort and to achieve a simplification of the design.

The problem is solved by the features listed in the main claims and further claims. Further advantageous embodiments emerge from the subclaims and the description.

The invention is explained using exemplary embodiments based on drawings. Show

Figure 1 shows the gearbox structure and the arrangement of the individual components;

Figure la Gearbox diagram for 2-range gearbox

Figure 2 shows a further embodiment of the invention;

Figure 2a Gearbox diagram for 1-range gearbox

Figure 2b Gearbox diagram for 2-range gearbox

Fig. 2c and d Gearbox diagram for 1-range gearbox as another embodiment
Fig. 2e and f Speed plan

Figure 3 the transmission diagram with provision for all-wheel drive;

Figure 4 shows the transmission diagram for longitudinal design and front-wheel drive.

Figure 5 Circuit diagram

Figure 6 Shift correction diagram

Figure 7 Switching valve with non-return valve

The transmission design according to Figure 1 shows a hydrostatic-mechanical power split transmission, in particular for front-wheel drive vehicles in transverse design with an integrated axle differential 15. The structure of the transmission is such that on the central shaft, which forms the drive shaft 16, the hydrostatic transmission 36, which in particular is used as

Compact transmission is designed and has an adjustment unit A and preferably a constant or adjustment unit B. A summation transmission 37 is arranged in addition to or after the hydrostatic transmission 36, which sums up the hydraulic and mechanical power. At least two clutches Kl and K2 are assigned to the summation planetary transmission 37, which alternately transmit the power summed up in the summation planetary transmission 37 via an output gear 10, another drive gear 11 and an intermediate shaft 12 to the axle differential transmission 15. The hydrostatic transmission 36, the summation planetary transmission 37 and the clutch pack 43 and the output gear 10 are arranged one after the other in the order listed. The transmission preferably has two clutches Kl and K2, which are arranged one above the other in a short manner, the clutches themselves preferably being designed as positive clutches with no drag losses. A common coupling member 44, which is in drive connection with the output shaft 10, is in constant drive connection with the output shaft 10. In order to achieve a short design, the front or input side housing opening 20 is closed off by a front cover 19 or 38, which is preferably designed as a stamped sheet metal part and has a noise-insulating effect in that the noise radiation emanating from the hydrostatic transmission 36 is shielded to the outside, in particular towards the drive side. A direct connection or contact of the front cover 19 or 38 with the hydrostatic transmission is avoided according to the invention. This front cover 19 or 38 is preferably made of noise-insulating material or of an anti-drum sheet or a so-called sandwich sheet. The type of front cover 19 or 38 shown in Figures 1 and 2 requires very little installation space, which has an impact on the length of the transmission. The shaft sealing ring 40 is advantageously inserted in the front cover 19; 38 mentioned without contact with the hydrostatic transmission 36. According to the design in Figure 1, the front cover 19 described is held in an additional housing cover 2, which simultaneously serves as a bearing element 51 for the bearing of the hydrostatic transmission 36 relative to the main housing 1. According to the design in Figure 1, the hydrostatic transmission 36 is mounted at two points, bearing point 51 and bearing point 42, relative to the housing 1, with noise-reducing elements 50 and 41 being mounted between them. To absorb the reaction moment of the hydrostatic transmission 36, support elements 8 and 9, preferably elastic or noise-reducing, are mounted between the driving surfaces. According to the embodiment in Figure 2, the hydrostatic transmission 36 is also mounted at two points, with one bearing point 42 serving opposite the housing 1 and the second bearing point without

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direct mounting on the housing via the drive shaft or central shaft 16 is carried out by a central bearing on the input side, which preferably sits directly on the drive shaft as a roller bearing 39. The front cover 38, which serves as a cover and for sound insulation or soundproofing of the internal noise of the gearbox, is connected directly to the gearbox housing 1 as shown in Figure 2. To further reduce vibration and noise, an elastic or sound-insulating element 2 is to be placed between the housing and the front cover 38, which can also serve as a seal at the same time. A separate device or one or both of the sound-insulating bearing elements 50 or 41 can be used to axially fix the hydrostatic gearbox 36. In the case of a design with simple hydrostatic bearing, a device for axial fixation at the bearing point 42 can preferably be provided as shown in Figure 2. It is also sensible to use the hydrostatic transmission 36 with an axial fixation 51 in relation to the front cover 38, whereby the radial bearing of the hydrostatic transmission in the front area in relation to the housing preferably remains free. It is preferably provided here by a bearing 84 on the output or crankshaft 60 of the drive motor and/or the front central bearing 39 located on the transmission drive shaft 16, which is installed in the hydrostatic housing, for example. The axial fixation is preferably carried out by an intermediate ring, e.g. designed as a friction ring 51, which is particularly effective as a friction damping element and absorbs axial forces resulting from helical gearing, for example.

The oil supply for the hydrostatic transmission and/or the control and regulating device and/or the clutch control and lubrication is provided, preferably in extreme space conditions, from the external central hydraulic system of the vehicle, whereby the overall length of the transmission can be reduced to a minimum in accordance with certain vehicle requirements. In the event that no external hydraulic supply is possible or the overall length of the transmission allows it, it is possible to use the feed pump 49 at the end of the transmission, preferably as an externally attachable pump unit, which can be driven directly or indirectly via the drive shaft 16. In order to achieve a short design, the output gear 10 is supported by a main bearing 27 and a second bearing 45 on a bearing carrier 46 connected to the housing 1. The second bearing 45 is preferably a needle bearing, which cannot be arranged in the area of the clutch pack 43 in a way that determines the overall length. The main bearing 27 is preferably arranged directly inside the gear 10 so that there is no influence on the overall length. This main bearing 27 is preferably designed as a fixed bearing

designed to absorb the axial forces resulting in particular from the helical gearing. It can preferably be designed as a cylindrical roller bearing in such a way that the roller track is incorporated in the output gear 10 and in the bearing carrier 46. However, a special axial bearing (not shown) can also be provided to absorb the axial forces, so that the main bearing 27 can also be designed as a floating bearing. A four-point bearing (not shown) or similar can also serve as the main bearing 27, as an independent four-point bearing or similar, whereby the second bearing 45 can be omitted. The design of the bearing carrier 46 as an oil guide body, which contains the inflow line for the control oil of the range clutches K1 and K2, is also advantageous in terms of costs and installation space. The oil lines are preferably shown in the form of cast-in pipes 28 or correspondingly designed oil lines in order to reduce costs and provide further advantages. The bearing carrier 46 is connected to the housing 1 via a screw connection 32, either from the inside of the housing or, in a form not shown, from the outside of the housing. Depending on the gearbox design, the bearing carrier 46 is attached to the gearbox input (Fig. 1a; 2c; 2d) on the housing 1 or on the front cover 19 or is attached to the opposite end of the housing, as shown in Fig. 1, inside the housing or from the outside.

The drive connection from the output gear 10 to the differential 15 is made via an intermediate shaft 12, which establishes the drive connection via gears 11, 13 and 14. For the assembly of the intermediate shaft 12 and the gear 11 meshing with the output gear 10, a housing opening 21 is provided, which is closed by the second housing cover 3, which preferably carries the bearing 24. The bearing adjustment of the roller bearings or tapered roller bearings 24 and 25 of the intermediate shaft 12 is carried out in a simple and time-saving manner using a shim 26.

The design and arrangement of the individual gear components shown allows the housing to be made in one piece, so that costly joints can be omitted. The described gear components arranged on the central axis or drive shaft 16 can be installed from the front housing opening 20. The intermediate shaft 12 and the bearings 24 and 25 with the associated gear 11 can be installed very easily through the housing opening 21. The differential gear 15 can be installed through a third housing opening 22, which can be closed by a corresponding housing cover 23, preferably designed as a sheet metal cover. The housing itself is simple in terms of shape and can be manufactured inexpensively.

The system of hydrostatic power splitting shown is based on the fact that, as already described, two shift ranges can be switched via clutch 1 and clutch 2, whereby the first shift range also includes the reverse range, as can be seen from the speed diagram in Fig. 2f. Both shift ranges work with power splitting. The drive shaft 16 is connected to the first hydrostatic unit A and a member of the summation planetary gear (the ring gear 29). The second hydrostatic unit B is in drive connection with a second ring gear 30 of the summation planetary gear. First planetary gears 33 and second planetary gears 34 are mounted on the web 32 of the summation planetary gear, and are in mutual meshing. The first planetary gears 33 mesh with the first ring gear 29, and the second planetary gears 34 mesh with the second ring gear 30. A sun gear 31 meshes with the second planetary gears 34. In the first switching range, the web 32 can be connected to the output shaft or the output gear 10 via a clutch K1. In the second switching range, the sun gear 31 can be coupled to the output shaft or the output gear 10 via the clutch K2.

The functional sequence is, as shown in the speed plan according to Fig. 2f or 2e, such that in the starting state the hydrostatic transmission is set to a certain negative speed for the second hydrostatic unit B, which corresponds to a certain adjustment variable - starting manipulated variable of e.g. 60%. When starting, with clutch K 1 closed, the power accumulated in the summation planetary gear is transferred to the output via the web 32, whereby an output speed of the web shaft 32 from zero is generated by reducing the speed from the aforementioned hydrostatic setting VN in accordance with the starting manipulated variable. When the hydrostatic transmission is regulated to the opposite direction of rotation, preferably up to setting W of a certain maximum adjustment, synchronous operation is achieved for the clutch elements K2 with simultaneous block rotation · of all elements of the summation planetary gear. After switching the second range clutch K2 and opening the first range clutch Kl, the hydrostatic unit is adjusted in the opposite direction, whereby the gear ratio end point is reached at the end of the hydrostatic adjustment in the opposite direction. For the reverse range, the hydrostatic adjustment is further adjusted from the aforementioned vehicle standstill position VN - starting control variable - in the negative adjustment direction, whereby the maximum reversing speed or the gear ratio end position of the reverse range is reached at the end of the negative hydrostatic adjustment.

With this gearbox design, it is possible to achieve a short design, particularly when using claw clutches, with or without deflector teeth, as known from EP 0 276 255 and EP 0 343 197, for example, whereby the two clutches K1 and K2 are advantageously arranged one above the other.

In a further embodiment, not shown, the clutches K1 and K2 are connected upstream of the summation planetary gear, with the output gear 10 being directly connected to a member of the summation planetary gear. In this configuration, the second hydrostatic unit B is preferably alternately connected to one of two input shafts of the summation planetary gear (not shown).

In Figure 2a, a hydrostatic-mechanical power split transmission is shown as a single-range transmission. According to the invention, this transmission has a hydrostatic transmission 36, which is preferably designed as a compact transmission and has a first hydrostatic unit A with adjustable volume and a second hydrostatic unit B with constant or adjustable volume, to which a three-shaft summation planetary transmission 37a is assigned, the first shaft 61 of which is connected to the drive shaft 16 and the first hydrostatic unit A and the second shaft 63 of which is connected to the second hydrostatic unit B and the third shaft 62 of which is in drive connection via a drive train 10, 11, 12, 13, 14 with the output shaft 64, 65 or the differential 15, the summation planetary transmission 37a being arranged coaxially to the hydrostatic transmission 36. In this transmission design, the summation planetary gear 37a is also spatially arranged downstream of the hydrostatic transmission 36 in the following order: drive motor, hydrostatic transmission 36, summation planetary gear 37a. The differential gear 15 and an intermediate shaft 12 belonging to the output train are arranged parallel to the input shaft 16 and the hydrostatic transmission 36.

In a further embodiment according to Figures 2c and 2d, the invention provides for the summation planetary gear 37a; 37 to be placed between the drive motor and the hydrostatic transmission 36. This has the advantage that the intermediate shaft 12 can be omitted and instead a less expensive intermediate gear 13a is used, which establishes the drive connection between the third shaft 62 of the summation planetary gear via further gears 10 and 14 with the differential. In this embodiment according to Figures 2c; 2d, the drive shaft 16 is guided on the input side through the output gear 10 or the output member and the second hydrostatic unit B and connected to the first shaft 61; 29 of the summation

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planetary gear 37a; 37 and connected to the hydrostatic unit A, which is spatially downstream of the hydrostatic unit B.

Similar to the design according to Fig. 2c, 2d, the two-range transmission system, as shown in Figure la, can also be implemented. Here, the drive shaft 16 is guided through the output member or through the output shaft 10, the clutch K2 and through the sun gear 31 and connected to the first shaft 29 of the summation planetary gear 37 and the first hydrostatic unit A, which in this design is spatially arranged downstream of the second hydrostatic unit B.

The three-shaft summation planetary gear 37a preferably has a sun gear 61, a ring gear 63 and a carrier shaft 62, wherein the sun gear 61 is connected as a first shaft to the drive shaft 16 and the first hydrostatic unit A, the ring gear 63 is connected as a second shaft to the second hydrostatic unit B and the carrier 62 is connected as a third shaft to the output member 10.

In a further embodiment not shown, the three-shaft summation planetary gear can also be designed with two ring gears and a carrier shaft, similar to the summation planetary gear 37 according to Fig. 2b. In this case, the first ring gear 29 is connected to the drive 16 and the first hydrostatic unit A, the second ring gear 30 to the second hydrostatic unit B and the carrier shaft 32 to the output member 10, with intermeshing planetary gears 33 and 34 being arranged on the carrier shaft 32, which in turn mesh with the ring gears 29 and 30. Which embodiment is to be preferred is decided by the required stationary gear ratio of the planetary gear and/or the spatial and structural conditions. Both planetary gear designs are functionally identical.

The functional sequence of the single-range transmission design, according to Figure 2a; 2c or 2d, is the same as the first shift range of the two-range transmission according to Figure 1; la or 2b, as described in more detail elsewhere and in the patent claims and shown in the speed diagrams Fig. 2e and 2f. In this design, the second hydrostatic unit B is advantageously equipped with secondary control, preferably for an overdrive range.

To improve efficiency, the invention provides a device that can be used for all transmission designs or for all power-splitting transmissions for bridging or releasing the load on the hydrostatic transmission 36 at one or more transmission ratio points. The respective transmission ratio point is held here as long as the operating situation allows it. This state is referred to as "fixed point switching" and is described in EP 0 599 263,

DE 43 39 864 or DE 44 17 335 are described in more detail. In the case of a transmission design according to the single-range system according to Fig. 2a; 2c; 2d, a purely mechanical operation is achieved in the final gear ratio by connecting the output 10 directly to the drive or drive shaft 16 via a clutch 67; 68. When the relevant clutch 67; 68 is closed, a block circuit of the summation planetary gear 37a is effected, so that the power transmission can take place purely via the drive shaft 16 directly or indirectly to the output 10 during block rotation of the summation planetary gear. In this switching state, the hydrostatic transmission is set torque-free or differential pressure-free, such that the hydrostatic adjustment is automatically adjusted accordingly or that a bypass valve (not shown) is connected between the two working pressure lines or the hydrostatic circuit, which creates a short circuit between the lines mentioned after appropriate control. In hydrostatically power-free operating states, i.e. when the adjustment variable is "zero" or the displacement volume is "zero", the invention also provides a clutch or brake 69 that connects the second hydrostatic unit B to the housing, so that the support torque is absorbed on the housing and the hydrostatic transmission is not burdened with unnecessary power loss.

In order to achieve smooth switching with hydrostatic bridging, the invention further provides that the switching on process preferably always takes place in the synchronous state of the relevant clutch elements, whereby the synchronous state or the synchronous and/or switching signal is determined from a speed comparison of at least two transmission elements in a known manner, e.g. via speed sensors. The switching off process is also designed so that the relevant clutch 67; 68; 69 is set to a torque-free state before the opening signal is initiated. This is done in such a way that the hydrostatic transmission automatically initially produces the load-free state of the clutch by means of a corresponding load-dependent correction of the hydrostatic adjustment device to a correspondingly adapted new displacement volume Vnew, which is adjusted depending on the load state. The load-dependent adjustment correction is carried out in particular from load-dependent operating values of the engine and/or transmission. This can include one or more operating signals or operating variables, such as throttle valve position, accelerator pedal position, accelerator pedal change speed, engine control signal in connection with the current engine speed, brake signal, external operating influences, temperature, air pressure, etc., which are suitable for determining or calculating the respective load state or load moment.

The current engine torque or the current load state can be determined or calculated in the control device or in the electronics using the above-mentioned operating variables or operating signals, from which the required adjustment correction or the required new displacement volume Vnew is determined in order to automatically activate the opening signal for the relevant clutch or the exit from the hydrostatic bridging switching state (fixed point switching) without any switching shock. In many applications, only two variables such as the accelerator or accelerator pedal position or throttle valve position or engine control signal and engine speed η mot are sufficient for acceptable switching quality.

The switching sequence shown can also be used for the usual range switching from one switching range to another, e.g. from range 1 to range 2 or vice versa with the same signal processing in the control/regulation. During overrun operation, the system recognizes the overrun behavior and the overrun torque and torque size from the increased engine speed and, if applicable, the accelerator pedal position or throttle valve position or engine control size, whereby a reversal of the adjustment correction direction is taken into account during overrun operation.

In conjunction with the described bypass valve (not shown), it is possible according to the invention to implement a constant hydrostatic adjustment correction depending on the respective operating state within the switched hydrostatic bridging phase with no load, so that when this state is left, the correct adjustment size or the correct displacement volume Vnew is spontaneously available, so that a spontaneous switching sequence without switching shocks into the new stepless range is ensured. With this mode of operation, the bypass valve is first closed before leaving this switching position and then the opening signal for the old clutch Kl or K2 or the bridging clutch 67; 68; 69 is activated.

The invention provides an automatic switching mechanism for the fixed point circuit 67; 68; 69, which controls the nearest fixed ratio point in an efficiency-oriented manner, taking into account the most favorable consumption ratios, taking into account the respective engine efficiency and transmission efficiency. For this purpose, consumption-oriented values of the drive motor and efficiency values or efficiency-determining data of the transmission are stored, whereby in the vicinity of a fixed ratio point with the aid of

The most fuel-efficient operating state is determined based on the aforementioned engine and transmission data and, if necessary, automatically by adjusting the engine speed and the transmission ratio accordingly, the relevant ratio fixed point is reached by closing one of the clutches 67; 68; 69; K1 and K2. Leaving the ratio fixed point is also determined and activated by a more or less precise calculation process based on the same criteria in the control device.

The hydrostatic bridging devices 67; 68; 69 can be designed very cost-effectively as positive-locking clutches with clutch teeth that can be produced without cutting, as described in more detail in EP 0276 255, or as friction clutches, preferably as cone clutches, as described in DE 41 26 650.

The speed diagram according to Fig. 2e shows the speed curve of the second hydrostatic unit B and the second shaft 63 of the summation planetary gear 37a for the single-range transmission according to Fig. 2a, 2c and 2d. If the second hydrostatic unit B is designed as an adjustment unit, it is possible to increase the total gear ratio range for both forward and reverse operation by reducing its adjustment volume, whereby the forward driving range has the size BV and the reverse range has the size BR'. Figure 2f shows the speed curve of the second hydrostatic unit B or the second shaft 30 of the summation planetary gear 37 of the two-range transmission described earlier according to Fig. 1; la; 2b; 3.

According to this invention, a further, additional device is very advantageous, in which the hydrostatic transmission 36 can be used as a brake retarder in conjunction with the described fixed-point circuit or hydrostatic bridging device. This device is described in more detail in the aforementioned publications EP 0 599 263 and DE 44 17 335.

The possibility of braking energy recovery and/or energy storage is also provided and is explained in more detail in the aforementioned, well-known publications.

To reduce noise emissions, the hydrostatic transmission 36 is also mounted elastically or in a noise-insulating manner in relation to the housing 1 in the transmission design according to Fig. 2c and 2d, for which purpose elastic bearing elements 50, 51 are used for radial support and one or more torque-supporting, noise- and vibration-insulating elements or torque support bodies are used for torque transmission. A common body 41 can be used for radial, axial and torque support, which can be inserted into corresponding

Recesses 8 of the hydrostatic transmission and the housing are used. Precise radial fixation of the hydrostatic transmission is ensured by machining the corresponding radial bearing surfaces of the hydrostatic transmission and the housing and the pockets or recesses 8 between the housing and the hydrostatic transmission can be unmachined, possibly cast or embossed recesses.

The transmission according to Fig. 2a, 2b, 2c, 2d is also suitable for all-wheel drive, whereby the installation takes place in a longitudinal design and one output shaft 64 serves, for example, for the front-wheel drive and the other output shaft 65 serves, for example, for the rear-wheel drive, whereby the differential 15 can serve as a longitudinal differential, if necessary with differential lock 66.

The transmission designs shown according to the invention are characterized by the following advantages: short design, small installation space, small number of components, good transmission efficiency, since the majority of operation - approx. 85% of operation - is with hydrostatic bridging, low fuel consumption, in particular due to the ability to utilize a large overdrive range, reduction in noise emissions through special extra encapsulation of the hydrostatic transmission 36 mounted in the main housing 1 and use of special, aforementioned noise-insulating intermediate elements 50, 51, 41, 8 and 9.

The entire transmission structure can also be implemented for a power split transmission with more than two shift ranges, whereby more than two clutches can be switched for more than two shift ranges. The respective limitation lies in the installation space available or depends on the installation length available for the transmission of the respective vehicle.

For all-wheel drive, it is very advantageous to design the transmission with an output shaft 55, which creates the drive connection to the rear axle. A bevel gear 59 is used for this purpose, which is brought into drive connection with an output member 14. A bevel gear 53 is advantageously connected directly to the differential drive gear 14, which drives the output bevel gear 54 and the corresponding output shaft 55. The bevel gear 59 can be very advantageously inserted into the transmission housing 1.

The gearbox design according to Fig. 4 provides an axle differential 58 that is located transversely to the drive shaft and is suitable for a front-wheel drive vehicle with a longitudinally installed engine. For this design, too, the gearbox housing 1 is made in one piece, as is the case with the gearbox designs described above.

The invention further provides a device that can also be used for other power split transmissions to optimize the shifting quality, in particular for the range shifts from one shift range to another. As described in more detail in the known publication EP O 599 263 A2 and is part of this invention, an adjustment correction of the hydrostatic transmission is carried out during the shifting phase after the new clutch has been closed and the old clutch has been closed. This shift correction device can be used for all power split transmissions, in particular hydrostatic-mechanical power split transmissions with two or more shift ranges.

The adjustment correction for gear shift optimization is preferably controlled/regulated via operating values of the drive motor M. The size of the adjustment correction X; Y; Z is determined by the respective or current drive torque Tan of the transmission or the engine torque Tmot. The closing process of the new clutch after the end of the old shift range takes place in the synchronous state of the clutch elements of the new clutch in a known manner, with two or more speed sensors preferably determining the synchronous state by comparing the speeds of at least two transmission elements and triggering the switching pulse for the new clutch. The current drive torque of the transmission or engine torque Tan determines the load size of the hydrostatic transmission and the corresponding speed slip of the hydraulic motor/pump and thus also the size of the adjustment correction required in each case within both closed clutches. Each drive torque is therefore also assigned a specific adjustment size or displacement volume Vnew or adjustment correction size Z; tk, which corresponds to the torque-free state of the old clutch, after which the opening signal for this clutch is triggered. Torque changes within the switching phase, where both clutches are closed, are automatically taken into account in this correction device, since the current drive torque Tan or engine torque Tmot always determines the opening signal of the old clutch or triggers the opening of the old clutch via corresponding resulting signals. Regardless of whether it is a traction upshift, traction downshift, overrun upshift or overrun downshift, the system recognizes the most favorable opening point or the new adjustment volume Vnew to trigger the opening signal of the old clutch. The recognition in the control device of whether overrun or traction operation is carried out by engine parameters or load variables/signals, if necessary also external influencing variables, whereby in traction operation each engine torque Tmot and

Engine speed &eegr; mot is assigned a specific control variable such as throttle valve position; driving or accelerator pedal position, accelerator pedal change rate, temperature, possibly also air pressure, etc. For example, if the accelerator pedal is suddenly released during the gearshift phase, the immediate opening signal at Vth can be given because the drive torque or engine torque immediately drops to zero or even assumes negative torque, even though VaIt and correction variable X were at their maximum at the start of the gearshift.

During overrun operation, the system also recognizes the respective operating state by revving the engine accordingly when the overrun torque is present and, depending on the speed, recognizing a corresponding braking torque or negative engine torque or overrun torque value according to signal b and determining the torque value Tmot; Tan and torque direction from the corresponding signal values b, a or braking signal f and using this determining and signaling the shift correction value and direction and Vneu for opening the old clutch. During overrun operation, the correction values and directions are reversed, whereby the displacement volume or the control value VaIt is smaller than Vneu when shifting up and VaIt is larger than Vneu when shifting down. This corresponds to the general characteristic shift correction behavior in hydrostatic-mechanical, but also purely mechanical power split transmissions.

The speed slip values of the hydrostatic transmission can be very different before and after the range shift with the same input and output torque at the range shift point, depending on the transmission design and range division. The correction value ratios X to Y or Z are correspondingly different. Accordingly, the correction values or correction value ratios X to Y or Z are programmed into the control and regulation system, i.e. each drive torque for each range shift point is assigned and pre-programmed with its own control value or displacement volume value Vneu, which can be corrected depending on the load conditions and other factors that change during the shift phase, so that a range shift without any switching shocks is guaranteed in all operating states.

Depending on the type of engine control, whether power or speed control, e.g. RQ, RQV or other type of control, the signal sizes suitable for determining the engine torque are programmed into the control and regulating device in the drive control or control and regulating device. For example, with an RQ control, each throttle valve

or accelerator pedal position and a given engine speed, so that the throttle valve or accelerator pedal position signal and the engine speed signal provide information about the respective engine torque, according to which the new displacement volume or the adjustment variable Vnew is determined and the opening signal for the old clutch can be initiated from this. With RQV control, each accelerator pedal or accelerator pedal position corresponds to a given engine speed value regardless of the engine torque. With this or a similar type of control, it is therefore necessary to use a corresponding signal to determine the engine torque, which corresponds to the filling level of the fuel injection or a similar signal value that is suitable for determining the torque.

With this correction variant, in contrast to known devices for gearshift correction, it is not necessary to determine or calculate the adjustment correction value Z from a hydrostatic pressure signal or a correction value X or VaIt given before the start of the gearshift, but the opening signal can always be determined from the currently effective drive torque Tmot or Tan. An appropriate adjustment travel sensor or sensor that signals the adjustment angle or adjustment travel of the hydrostatic transmission usually ensures precise implementation and signaling of the new control value or Vneu.

Alternatively, the known adjustment pressure or electrical adjustment current can be used to measure the adjustment size or displacement volume, provided that these are suitable for determining the correct opening signal for the old clutch.

A pressure sensor for detecting the respective hydrostatic pressure is not required with this correction variant.

All shift correction devices described in this patent application can be used for both hydrostatic-mechanical and purely mechanical power split transmissions. In a mechanical power split transmission, the respective torque ratios on the continuously variable converter apply to the correction ratios X and Y, whereby the torque ratios correspond to the pressure ratios of the working pressures of the hydrostatic transmission.

The invention further provides a shift correction device for the range shifts for both upshifting and downshifting, traction or overrun shifting, according to which the shift correction value Z is determined over an adjustment correction time tk, wherein preferably

the size of the required correction values X, Y or Z or tk is determined from the above-mentioned respective operating values such as engine torque or from the torque-determining operating values. The adjustment correction time tk determines the size of the correction value or path Z at the switching time or in the switching phase depending on the adjustment speed. The control current or the delivery volume Qk is therefore decisive for this. The effective control current Qk is determined or influenced by various values that are effective at the switching time, such as feed pump delivery volume, speed, control pressure (constant or variable), internal throttling effects, oil temperature, etc. The size of the adjustment speed or the adjustment current Qk can be determined experimentally in dependence on the various operating conditions, - engine speed, engine control signals such as throttle position, accelerator pedal position, oil temperature, etc. - From these aforementioned values, the control device or the electronics recognizes which adjustment correction time is required in which operating condition to determine the respective correction size or the new displacement volume Vnew. To accurately determine the correction time tk, the temperature, in particular oil temperature, oil viscosity and, if applicable, other factors influencing the volume flow of the adjustment devices, leakage oil changes and the switching time are also taken into account by appropriate signal processing in the control device.

According to the invention, the control and regulating device or electronics also automatically change and adapt predefined values, which are dependent in particular on the operating time and/or type of use. The adaptation mentioned can be implemented in various ways, e.g. in such a way that malfunctions or disturbances in the control device or electronics are detected, from which, for example, a shock-generating signal or disturbance signal causes a change in one or more predefined variables or fixed values, so that despite changing operating values, e.g. the amount of leakage oil, the correction variables tk; Z are adapted in such a way that good switching quality is maintained or improved. The triggering signal for this correction or internal correction of preferred specified or prevailing variables can be a torque change of an engine or transmission element, particularly during the switching phase, or a torque or speed changing signal or/and rate of change or/and change of a mass force or generally a signal indicating a shock or jolt. A suitable signal variable for changing internal specifications

·

or fixed values can be the engine speed signal, particularly in the constant driving range, whereby a deceleration shock triggers an increase in the engine speed, which automatically brings about a corresponding change in one or more of the internal default values or fixed values and/or signal size, so that in this case, for example, a corresponding increase in the adjustment correction value tk or Z is achieved. In the event of an acceleration shock, a reverse change in the aforementioned fixed values is brought about. Instead of the signals used to change internal fixed values, there can also be a signal that results from the change in an above-mentioned inertia force, whereby, for example, the change in mass within the switching sequence results in a decision to shorten or lengthen the switching correction time tk or the adjustment correction size Z. The amount of change in the internal fixed values depends essentially on values that change over the course of the operating time, which can, for example, be dependent on the wear of individual elements and/or dependent on a significantly changing operating characteristic of a vehicle. This means that in order to optimize or maintain good switching quality, the electronics or control system makes the decision to change one or more internal fixed values preferably from the information of several switching operations in order to determine the most suitable change value.

The invention further provides that, with regard to the reduction in switching time, the synchronous speed range, which includes the degree of synchronous inaccuracy, can be of different sizes depending on one or more operating parameters. This means that the signal for closing the new clutch in a range shift or the clutch for fixed point shifting can be triggered at a greater or lesser distance tkS from the absolute synchronous point. The electronics take into account the required closing time from the time of the signal triggering to the start of the active closing process, for example in the case of a very rapid gear ratio change. Accordingly, the clutch closing signal is initiated accordingly early before the synchronous state is reached or before the permissible synchronous range is reached. This is particularly important during high acceleration processes, in which a correspondingly high kick-down effect is also effective, or during braking processes, which require a correspondingly high gear ratio reduction in the transmission. The aforementioned synchronous range can vary in size depending on the type of clutch - friction clutch, e.g. in the form of a multi-disk or cone clutch or positive-locking clutch with or without deflection teeth.

· · ■

The exact values should preferably be determined experimentally. The electronics obtain the information for the most suitable switching point from, for example, the speed/force of change of the gear ratio and/or the actuation force/speed of the accelerator pedal and/or brake pedal or other suitable operating parameters or influencing factors that result from the experimental determination and findings. The feedback signal to indicate the closed new clutch provides the impulse for initiating the correction adjustment Z or tk.

Likewise, as explained in more detail above, the opening time for the old clutch can also be varied differently or brought forward by a corresponding amount tkv.

The switching correction device with time-dependent switching correction has the advantage that a costly hydrostatic pressure sensor and in some cases also a hydrostatic adjustment sensor (potentiometer; travel sensor) can be dispensed with. This process is suitable for both the range switching and the fixed point switching KB; KH; KD in transmissions as known from DE 43 39 864 and EP 0 599 263 and also for reversing switching, e.g. for the reversing operation of a work machine - wheel loader, tractor front loader etc. -. Particularly when using form-locking clutches with a preferred design as deflection gears or friction clutches, the opening signal can be initiated before the torque is completely transferred from the old to the new clutch, i.e. before the end of the adjustment correction Z or correction time tk, which greatly reduces the switching time, since after the opening signal has been given, the relevant or old clutch can be pressed open by the deflection function. The opening signal is thus brought forward by the time tkv, whereby tkv can be influenced by one or more of the above-mentioned operating signals and/or change signals. The adjustment process of the hydrostatic transmission can thus be continued largely continuously even during the opening process of the aforementioned clutch, whereby a functional overlap of the opening signal or the opening process of the aforementioned clutch and the hydrostatic adjustment is effective, which results in a reduction in switching time and high switching quality.

The system detects overrun, pull or up/downshifts from the respective engine speed and the throttle valve position or the size of the engine control, from which the correction direction of the hydrostatic adjustment device is determined.

In support of the shift correction device, according to the invention, an additional influence or reduction of the engine torque within the shift phase can also be provided by automatic throttle release, e.g. when using an electronic accelerator pedal, in order to achieve optimal shifting quality for the range, reversing and fixed point shifts in all operating situations.

When closing a new clutch, the filling process of this clutch inevitably causes a reduction in the control pressure for the closed clutches, which can lead to a reduction in torque or even to the opening of the closed or old clutch. This applies to all hydraulically operated clutches or comparable devices or consumers. To prevent this, a device 64' as shown in Figure 7 is provided, which prevents or reduces oil backflow and thus a reduction in the control pressure for the old or closed clutch. Further advantages of this device 64', which is preferably designed as a check valve within the switching valve 64 or as a separate device, are that the feed oil quantity and thus the feed pump 36 can be made smaller and/or a provided hydraulic accumulator 36' can also be made smaller, or this can be dispensed with entirely. Furthermore, when using a form-locking coupling, in particular in a design such as in DE 41 26 650 A1, which is preferably a component of this invention and is shown therein in Figures 3, 3a and 3e, it can be designed with a lower driver profile, whereby the switching path and thus the pressure oil volume can be reduced to a smaller level. The aforementioned coupling with form-locking coupling teeth is a coupling device in which a rotationally fixed but axially displaceable coupling ring is arranged on a coupling carrier, which is acted upon by an axially displaceable hydraulically actuated piston, whereby when the coupling is closed, the aforementioned coupling ring engages in the corresponding counter profile of the second coupling half. The aforementioned coupling profile can be designed in a repelling or non-repelling form.

The aforementioned device 64' is preferably designed as a check valve within a switching valve 64, shown in Figure 7d, e.g. in such a way that a control piston 64a has a movable closure element 64b, which closes the inflow line 64e to the clutch via a spring element 64c after the clutch is closed. After the clutch is closed, the same pressure conditions are present in the control line K and the control pressure P, whereby a low spring force of the spring 64c is sufficient to close the closure element

or to bring the piston ring 64b into the closed position. The switching valve is preferably controlled by a solenoid valve 64d, which acts as a pilot valve. In the open state, the switching valve is held in the neutral position against the force of a spring 64c when the solenoid valve 64d is not controlled, with the clutch line 64e being connected to the return line 64g without pressure. The closing body 64b is held in a fixed position relative to the switching piston 64a by the control pressure P against the pressure of the spring 64c, until the clutch is closed after a switching operation has taken place.

The oil non-return valve in the clutch pressure line has the additional advantage that changes in the system pressure caused by any switchable consumers or others prevent the clutch from opening.

To improve transmission efficiency, the transmission is provided with a pressure valve 65' (Fig. 5) which effects a pressure modulation of the system pressure or the feed pressure depending on various operating parameters. The pressure valve 65 ^V is set to a minimum pressure which is sufficient to supply the hydrostatic transmission and/or the clutch control Kl or K2 at low load and/or at low engine speed. At higher loads, the system pressure is increased accordingly by a load-dependent signal, in particular hydrostatic pressure signal e and/or at higher engine speeds via a speed pressure signal b. This means that with predominantly partial load operation, there is little power loss through the feed pump 81.

When the continuously variable power split transmission is designed with a secondary-controlled hydrostatic transmission 36, as shown in Fig. 5a, the adjustment control is designed according to the invention in such a way that only one adjustment cylinder 79 or common adjustment gear is effective for both the primary control and the secondary control, i.e. for the control of the hydrostatic units A and B. This can also be used in a manner not shown for radial piston hydrostats. The adjustment cylinder 79 is preferably arranged parallel to the drive axis of one or both of the hydrostatic units A and B. An adjustment piston 80 is operatively connected to the swivel plate 75 of the first adjustment unit A. The entire positive and negative adjustment range of the first hydrostatic unit A is traversed via a piston travel SV. With the swashplate position VO the delivery volume of the first hydrostatic unit A is set to zero delivery volume, which with a power split transmission at a certain forward driving speed at

hydraulic power zero as shown in Fig. 2e and 2f. With the swivel plate position W, the hydrostatic unit A is set to maximum delivery volume, which corresponds to equal speed and block circulation of all elements of the summation planetary gear 37 and 37a in both the two-range transmission (as per Fig. 1; 2b and others) and the single-range transmission (as per Fig. 2c; 2d). In the single-range transmission as per speed plan Fig. 2e, from this point onwards

Point W is the secondary control, whereby further adjustment of the adjustment piston 80 results in a resetting of the delivery volume of the hydrostatic unit B, in that the secondary adjustment disk 76 is reset from the maximum adjustment angle to a correspondingly small adjustment angle via a possible secondary adjustment path SS. In position VS, the final gear ratio of the transmission is reached as per design Fig. 2a; 2c. The secondary control can, as shown in Fig. 5a, be implemented via an adjustment piston 80, which presses on a corresponding pressure piece 82 at the start of the secondary control. In another design not shown, an internal, separate adjustment piston in the same adjustment cylinder 79 acts on the swivel disk 76 of the second hydrostatic unit B. In a further embodiment not shown, the secondary control can already take place within the primary control range, whereby, for example, the secondary adjustment takes place within the adjustment path or adjustment angle aV, whereby the pressure piece 82 must be made correspondingly longer. In this case, for example, with a maximum adjustment angle of 20° for both hydrostatic units A and B, the secondary control could begin after an adjustment path aV 10*, whereby at the end point of the transmission with an adjustment angle aV = 20* corresponding to the end position W, the angle aS of the secondary swivel plate 76 would correspond to approximately 10°. This means that with this embodiment, at the end of the primary adjustment in position W, the end point VS is reached with a secondary adjustment angle aS of approximately 10'.

In this last-described embodiment, the adjusting piston 80 is also positively connected to the swivel plate 75 via an articulated rod, whereby the piston rod 80' is preferably designed as a hollow body and the connecting piece between the piston 80 and the swivel plate 75 represents an internal push and pull rod with an articulated connection. The spring 77 can be omitted in this case.

A compression spring 83 is used to fix the position of the adjusting piston 80. Other spring devices not shown, preferably in the area of the adjusting piston 79 with a dual function, such as those known from conventional adjusting devices of hydrostatic transmissions, can also be used here, e.g. to fix a neutral position, which e.g.

Starting point, i.e. driving speed zero of the vehicle. A spring lock at VO can also be useful, especially for resistance-free starting of the engine in cold conditions or in winter operation.

Particularly in the case of the single-range transmission as per Figure 2a; 2c; 2d, since it does not contain any range clutches, the transmission has a bypass valve 84 as shown in Figure 5, which is connected between the two working pressure lines of the hydrostatic units A; B, which is particularly useful when starting the engine, towing, starting the engine by pushing or also for optimizing a fixed point position described or generally for creating a load-free state. The bypass valve can be actuated automatically or manually, depending on the type of requirements. For example, in preparation for the start-up process after actuating the selector lever, automatic actuation of the bypass valve is advantageous. When the vehicle is stationary, particularly when the selector lever is in the "Neutral" and/or "Park" position, the bypass function is always switched on. When the direction of travel is preselected as "reverse" or "forward", the bypass valve 84 is automatically closed according to a predetermined characteristic. The closing process is preferably designed in such a way that, in the case of a single-range transmission that has no range clutch or no output-side separating clutch, after the engine has started, the adjustment device of the hydrostatic transmission automatically assumes the starting position, which corresponds to the hydrostatic control variable VN. Depending on the transmission design, this control variable can correspond to approximately 60% of the maximum negative adjustment angle, as shown, for example, in the speed diagram in Figure 2e. Preferably, the adjustment device is always locked in place at this point by a mechanical device, in particular a spring device, when the vehicle is at a standstill, so that no or almost no differential oil quantities between the working pressure lines of the two hydrostatic units A; B need to be compensated via the bypass valve 84. In most applications, the bypass valve 84 can therefore always remain closed in the neutral position, so that it only needs to be activated to start the engine in winter operation or when the oil is very cold and/or to start the engine by pushing it. The activation can preferably be triggered manually or automatically via the control and regulating device or electronics, whereby the closing process of the bypass valve is preferably activated electronically in the case of electronic control. This ensures that the bypass valve 84 is always open when the engine is at a standstill. The bypass valve 84 can be designed in different ways. Depending on the application and requirements, a spontaneous or gradual pressure build-up or pressure reduction is possible.

Closing process is provided. For engine starting, it is recommended to implement a gradual continuous pressure build-up or continuous closing process, which preferably works in a pressure-dependent manner and can be used advantageously after pre-selecting the direction of travel for setting the swivel disc to the starting position VN, for example by supporting the adjustment process in a pressure-dependent manner until the setting point VN is reached.

Claims (28)

■ Claims 1. Hydrostatic-mechanical transmission with power split consisting of a hydrostatic transmission with a first hydrostatic unit (A) of adjustable volume and a second hydrostatic unit (B) of constant or adjustable volume and a power split device with a summing planetary gear (37; 37a), in which the drive power is divided into a hydraulic and a mechanical branch, which is summed up again before the transmission output, characterized in that the power split transmission is assigned a differential gear (15; 58), in particular serving as an axle differential, in particular for driving a front-wheel drive vehicle, which is arranged offset from the drive shaft (16) and is driven via intermediate members (10, 11, 12, 13, 14 or 10, 13a, 14), that the hydrostatic transmission (36) preferably has a common structural unit for the hydrostatic unit (A and B) in the form of a compact transmission and the structural units hydrostatic transmission (36), summation planetary transmission (37; 37a) and optionally clutches (44) are arranged coaxially to one another, that the transmission is designed as a single-range transmission without range shift clutches or as a two-range transmission with preferably two range shift clutches (Kl; K2), and the transmission in both embodiments operates over the entire forward and reverse gear ratio range with power splitting and the reverse range (BR or BR') is included in the first shift range or adjustment range of the hydrostatic transmission, such that in the starting state the hydrostatic transmission is set to a certain or defined negative adjustment value (VN) and that the hydrostatic transmission (36) is preferably mounted relative to the housing (1) via noise-insulating or noise-damping, in particular elastic and/or friction-damping elements (9; 50; 51; 8) and the noise and vibration-insulating elements (9; 50; 51; 8) are preferably effective against radial forces, axial forces and circumferential forces of the hydrostatic transmission relative to the housing (1). 2. Transmission according to claim 1, characterized in that the assemblies - hydrostatic transmission (36), summation planetary transmission (37; 37a) and if necessary differential transmission (15; 58) - are arranged in a common housing (1), wherein the housing (1) is designed in one or more parts and preferably a central housing opening (20) is provided, which serves as an assembly opening for the hydrostatic transmission (36), the summation planetary gear (37; 37a), if necessary the clutches (Kl, K2) and the driven gear (10) and that if necessary for the assembly of an intermediate shaft (12) belonging to the drive train, a second housing opening (21) is included, which is opposite the central housing opening (20) and arranged axially offset from the central axis and can be closed by means of a housing cover (3), which preferably contains the bearing (24) for the intermediate shaft (12) (Fig. 1; 2a; 2b) or that the structural units - hydrostatic transmission (36), summation planetary transmission (37; 37a) and if necessary, differential transmission (15; 58) - are arranged in a common housing (1), wherein the housing (1) is formed in one or more parts and preferably has a central housing opening (20) which serves as an assembly opening for the hydrostatic transmission (36), the summation planetary transmission (37; 37a), if necessary, the clutches (K1, K2) and the driven gear (10). 3. Transmission according to claim 1, characterized in that the structural units - hydrostatic transmission (36), summation planetary transmission (37; 37a), if necessary clutch pack (43) and driven gear (10) (Fig. 1; 2a) or - driven gear (10), if necessary clutch (43), summation planetary transmission (37; 37a) and hydrostatic transmission (36) (Fig. la; 2c) - are arranged one behind the other on an axis in the order listed, whereby the driven gear (10) is connected to a common coupling member (44) for at least two clutches (Kl, K2) or to a shaft (62) of the summation planetary transmission (37a), whereby the driven gear (10) is connected by a main bearing (27') which is designed as an independent bearing (e.g. four-point bearing, double tapered roller bearing or double-row angular contact ball bearing) or is mounted by a main bearing (27) and a further bearing (45) on a bearing carrier (46) that is firmly connected to the housing (1), the main bearing (27) preferably being designed as a fixed bearing and the second bearing (45) preferably being designed as a floating bearing, that the bearing carrier (46) is preferably designed as an oil guide body and serves for the oil supply and for switching the clutches (Kl, K2) and/or for lubricating oil and/or as a control oil line for the hydrostatic stabilization via a corresponding oil line in the hydrostatic drive shaft (48) for a modulable pressing of at least one of the hydrostatic units (A; B) onto its control disk and/or that in a further embodiment not shown, the structural units
Hydrostatic transmission (36), optionally clutch pack (43), summation planetary gear - are arranged one behind the other in the listed order, wherein the output gear (10) is provided with a Member of the summation planetary gear, wherein the output gear (10) is mounted by an independent main bearing (27') or a main bearing (27) and further bearing (45) on a bearing carrier (46) firmly connected to the housing (1), and wherein the main bearing (27) is preferably designed as a fixed bearing and the second bearing (45) is preferably designed as a floating bearing. 4. Gearbox according to claim 1bis3, characterized in that the gearbox housing (1) is closed at the front by a front cover (19; 38) which is made as a stamped sheet metal part, preferably from anti-drumming sheet or sandwich sheet or from noise-reducing material. 5. Transmission according to one or more of the preceding claims, characterized
characterized in that the transmission has only two clutches (Kl) and (K2), which serve to switch two hydrostatic-mechanical shift ranges, the first shift range also including the reverse range, or that the transmission has at least two shift clutches (Kl) and (K2), whereby the first shift range operates purely hydrostatically or in power split mode, or that the transmission has two or more shift ranges with several shift clutches, whereby all ranges operate in power split mode. 6. Transmission according to one or more of the preceding claims, characterized
characterized in that the feed oil supply comes from the central hydraulics of the entire vehicle, or that the feed pump is accommodated in the hydrostatic transmission (36) in a known manner, or that the feed pump (49) can be mounted as an externally attachable pump unit and driven via the central shaft (48). 7. Transmission according to one or more of the preceding claims, characterized
characterized in that the hydrostatic transmission (36) has a bearing device (42), according to which the bearing bodies (9; 50) consist of noise-insulating or sound-damping material, preferably an elastomer, and that a torque-driving device (9) is provided in the form that the circumferential forces are supported on noise-damping elements and that the hydrostatic transmission (36) is radially mounted on two or only one bearing point (42) on the main housing (1) and a second bearing point or the counter bearing takes place via a central bearing (39) sitting on the drive shaft or central shaft (16) and/or a bearing (84) sitting in the drive machine (crankshaft 60). 8. Transmission according to one or more of the preceding claims, characterized characterized in that the hydrostatic units (A, B) are designed as axial piston machines or as radial piston machines (not shown), the first hydrostatic unit (A) having a distributable volume and the second hydrostatic unit (B) having a constant or adjustable volume. 9. Transmission according to one or more of the preceding claims, characterized
characterized in that a bevel gear (59) is provided for all-wheel drive, the bevel gear (53) of which is rotationally connected to the differential drive gear (14) and that the bevel pinion (54) connected to the output shaft (55) for the rear-wheel drive can be inserted into the gear housing (1) with the associated bevel pinion bearing. 10. Hydrostatic-mechanical transmission with power splitting consisting of a hydrostatic transmission with a first hydrostatic unit (A) of adjustable volume and a second hydrostatic unit (B), of constant or adjustable volume and a power splitting device with a summing planetary gear (37; 37a), in which the drive power is divided into a hydraulic and a mechanical branch, which is summed up again before the transmission output, wherein preferably the power splitting transmission is assigned a differential gear (15; 58), in particular serving as an axle differential, preferably for driving a front-wheel drive vehicle and wherein preferably the differential gear (15; 58) is arranged axially offset from the drive shaft (16), which is connected via intermediate links (10, 11, 12, 13, 14 or 10, 13a, 14) is driven characterized in that the summation planetary gear is preferably designed with three shafts and the hydrostatic units (A and B) of the hydrostatic gear (36) are arranged concentrically or coaxially to one another, the drive shaft (16) being constantly connected to the first hydrostatic unit (A) and a first shaft (61; 29), the second hydrostatic unit (B) being constantly connected to the second shaft (63; 30) of the summation planetary gear and the third shaft (62) being constantly connected to the differential (15) via a drive train (10, 11, 12, 13, 14 or 10, 13a, 14) and the summation planetary gear (37a; ^ 37) being placed coaxially to the hydrostatic gear (36) (Fig. 2a; 2c; 2d). 11. Transmission according to one or more of the preceding claims, characterized
characterized in that the summation planetary gear (37a) is designed as a three-shaft with a sun gear (61), a carrier shaft (62) and a ring gear (63), the sun gear (61) with the first hydrostatic unit (A) and the drive shaft (16) connecting the ring gear (63) with the second hydrostatic unit (B) and the web (62) with the output (10) or the output train (10, 11, 12, 13, 14 or 10, 13a, 14) and the differential (15) is in permanent drive connection or that the transmission is designed with a four-shaft summation planetary gear (37) and has two ring gears (29, 30), a web (32) and a sun gear (31), whereby a first ring gear (29) is permanently connected to the first hydrostatic unit (A) and the drive shaft (16), the second ring gear (30) is permanently connected to the second hydrostatic unit (B) and the web (32) and the sun gear (31) can be connected via clutches (Kl or K2) to the output (10 to 15) via the corresponding drive train (10, 11, 12, 13, 14, 15 or 10, 13a, 14, 15). 12. Transmission according to one or more of the preceding claims, characterized
characterized in that the summation planetary gear is designed with three shafts and has a first ring gear (29), a second ring gear (30) and a carrier shaft (32) on which intermeshing planet gears (33 and 34) are arranged, the first ring gear (29) being permanently connected to the drive shaft and the first hydrostatic unit (A), the second ring gear (30) being permanently connected to the second hydrostatic unit (B) and the carrier being permanently connected to the drive train (10, 11, 12, 13, 14 or 10, 13a, 14) and the differential (15) (not shown). 13. Transmission according to one or more of the preceding claims, characterized
characterized in that the transmission in the version (Fig. 2a) with a three-shaft planetary gear (37a) represents a single-range transmission with power split over the entire gear ratio range forwards and backwards, whereby at driving speed "zero" the hydrostatic unit (A) is set to a certain negative displacement volume between the maximum negative adjustment and displacement volume "zero" and the first shaft (61) and the second shaft (63) of the summation planetary gear (37a) have opposite directions of rotation and with this setting when the displacement volume is increased a reverse speed or negative output speed and when it is reduced up to adjustment size "zero" and beyond that up to the maximum negative adjustment the forward driving range extends and that in the version (Fig. 2b) with a four-shaft summation planetary gear (37) a second hydrostatic-mechanical switching range with power split is connected, whereby in The first gear range includes the reverse range, which has the same effect as the single-range transmission (according to Fig. 2a; 2c) through a negative partial adjustment from a certain setting point of the hydrostatic transmission and the forward range consists of a part of the negative and the entire positive adjustment range of the hydrostatic transmission and wherein the second switching range with power split occurs with synchronous operation of the clutch elements (Kl, K2) to be switched and that preferably a switching correction device is provided which functions in such a way that the old clutch only opens after a certain load-dependent adjustment correction in an almost torque-free state. 14. Transmission according to one or more of claims 10 to 13, characterized
characterized in that a hydrostatic bridging device is provided, whereby by closing one or more clutches (67; 68; 69; K1 and K2) the hydrostatic transmission (36) can be set without load, wherein preferably the hydrostatic bridging device is designed such that a clutch (67; 68) or several clutches (K1, K2) connects members (62, 63) of the summation planetary transmission to one another or bring them into block circulation or into a rolling power-free state, so that the hydrostatic transmission (36) can be set without load and/or that a brake or a clutch (69) is provided which connects the second hydrostatic unit (B) to the housing (1) in the hydrostatic power-free state, so that no torque loads the hydrostatic transmission in this operating state and keeps the hydrostatic circuit free of differential pressure and/or that a range block circuit is provided, such that the transmission ratio is held at the range limits by closing two Range clutches (Kl and K2), wherein at the same time the hydrostatic transmission (36) is placed in a no-load state, in such a way that a bypass valve is provided between the two working pressure lines of the two hydrostatic units (A, B) and within this switching state creates a short circuit between the two working pressure lines, so that at least an approximately torque-free or differential pressure-free state of the hydrostatic transmission is ensured or that the hydrostatic adjustment is regulated so that no differential pressure occurs in the hydrostatic circuit. 15. Transmission according to one or more of the preceding claims, characterized
characterized in that the hydrostatic lock-up clutch (67; 68) and/or the clutch or brake (69) is engaged or closed when the clutch elements to be switched are running synchronously and that the opening signal for this aforementioned clutch also occurs in a virtually load-free state in which the hydrostatic transmission is adjusted before the clutch is opened so that it corresponds to the load state when the corresponding clutch is open, whereby the volume flow setting or the hydrostatic Adjustment device is preferably implemented by load-dependent signals, in particular signals determining the engine torque and/or the drive torque of the drive shaft (16) such as throttle valve position; accelerator pedal position, accelerator pedal adjustment speed, engine speed, engine control signal; brake signal. 16. Transmission according to one or more of claims 10 to 15, characterized
characterized in that for the range switching and/or the devices for hydrostatic bridging (clutch 68; 67; 69 or K1, K2), the hydrostatic control device for determining the displacement volume is designed such that within the aforementioned closed clutches or the associated operating states, the hydrostatic adjustment follows a constant load-dependent specification or adaptation in a load-free or at least almost load-free state of the hydrostatic transmission, so that when the aforementioned hydrostatic-relieving switching or operating state is canceled, that is to say when the clutch in question is opened and the bypass valve is closed or closing at the same time, a torque-free or at least almost torque-free state of this clutch is given, so that no switching shock occurs. 17. Transmission according to one or more of claims 10 to 16, characterized
characterized in that the hydrostatic transmission (36) is designed as a compact transmission and contains both hydrostatic units (A and B) and that this compact transmission represents a common component for the single-range transmission (as per Fig. 1 and 2b) and for the transmission version with two shift ranges (as per Fig. 1 and 2b), whereby version (Fig. 2a) is provided with one shift range for the low power range and version according to (Fig. 2b) with two shift ranges for the medium and higher power range. 18. Transmission according to one or more of claims 1 to 17, characterized
characterized in that the summation planetary gear is designed with three shafts, the first shaft being extendable as a sun gear (61) or as a ring gear (29) (not shown). 19. Transmission according to one or more of claims 10 to 18, characterized
characterized in that the hydrostatic transmission is designed in the form of an axial piston machine and/or a radial piston machine, wherein both units (A and B) can be realized parallel next to one another or radially one above the other in a concentric or coaxial arrangement or that one of the two units (A, B) can be extended as an axial piston machine and the other as a radial piston machine. 20. Transmission according to one or more of claims 10 to 19, characterized characterized in that the summation planetary gear (37, 37a) is spatially arranged downstream of the hydrostatic gear (36) (according to Fig. 2a; 2b) or is spatially interposed on the input side of the gear between the drive or drive motor and the hydrostatic gear (36) (according to Fig. 2c, 2d). 21. Continuously variable transmission, in particular with hydrostatic power splitting, preferably for motor vehicles with preferably several shift ranges and/or with a device for hydrostatic bridging in the form of a fixed point circuit (KH, KB, KD), with a first hydrostatic unit (A) with adjustable volume and a second hydrostatic unit (B), preferably with constant volume, with a summing planetary gear for summing up the hydraulic and mechanical power, in which the switching takes place without load interruption and/or without switching shock of the above-mentioned shift ranges or the fixed point circuits, wherein the hydrostatic units (A) and (B) exchange their function as pump and motor when changing ranges and the range change takes place at synchronous speed of the clutch elements of the new clutch or the elements for the fixed point circuit to be switched and the old clutch or clutch of the fixed point circuit (KH; KB; KD) only after the adjustment correction (Z) or a setting correction time (tk) at a new displacement volume (Vnew), characterized in that the opening signal for the old clutch in the case of a range change or clutch of the fixed point circuit (KH; KB; KD) is derived from one or more torque-determining operating variables of the drive motor (M) (throttle valve position or motor control signal, accelerator pedal travel signal (a), motor control signal, speed signal (b), temperature signal, etc.) or that the opening signal is formed directly or indirectly from a torque sensor that is operatively connected to the drive of the transmission and/or the drive motor (M) and/or that the switching correction size (Z) is determined via a correction time (tk) that depends on the respective load state of the engine and/or transmission, whereby the switching correction size (Z) is determined via a correction adjustment time (tk) or according to a pre-programmed time factor, such that the correction time (tk) required for the switching correction (Z) is determined from the respective load state of the engine and/or transmission, whereby in a transmission design with several different switching range sizes for each range switching point (e.g. 1/2; 2/3; 3/4; or. 4/3; 3/2; 2/1) and torque value (Tmot; Tan) a correspondingly adjusted own adjustment correction value or correction time (tk) or a correspondingly adjusted new Displacement volume (Vnew) is assigned and pre-programmed to determine the opening signal of the old or relevant clutch, so that the adjustment correction variable adapted to the respective load condition takes effect for switching shock-free operation. 22. Transmission according to claim 21, characterized in that load or torque changes occurring during the switching phase (with the old and new clutch engaged) are taken into account in such a way that the opening signal is determined by the currently effective load state of the engine and/or transmission or the torque (Ta) and/or that the opening signal or the new displacement volume (Vnew) specified for each load state and each switching point (range switching; fixed point switching) undergoes a correspondingly adapted correction depending on the current operating state and/or currently effective operating values such as (accelerator pedal change, change speed, acceleration state, throttle valve position, temperature, etc.) and/or the type of clutch design (clutch factor fk, friction clutch, positive clutch). 23. Transmission according to one of claims 21 and 22, characterized
that the switching correction device detects and processes a disturbance function and/or shock signal in such a way that one or more change signals become effective which change one or more of the internal fixed values or predetermined quantities, so that an automatic targeted adaptation of the switching correction quantities (Z; tk) takes place, wherein preferably the change signal for changing fixed values and/or predetermined quantities or basic quantities is a rotational speed signal and/or a speed change signal and/or a mass change signal and/or a signal which can be the change of a continuity in some form and that the change of internal fixed values results in particular from one or more disturbance signals which occur within one or more circuits. 24. Transmission according to one or more of the preceding claims, characterized
characterized in that for the range switching or fixed point switching the signal for closing the new clutch or the clutch in question is triggered within a defined synchronous range, whereby the synchronous range or synchronous error range can be changed automatically depending on various operating parameters, whereby different response times (tkS) for the start of the switching signal can be preprogrammed or implemented. ·· t · 25. Transmission according to one or more of the preceding claims, characterized
characterized in that a device (64') is provided for the clutch control which, when switching a clutch or another consumer, prevents or reduces a pressure drop in the closed clutch, preferably by means of a non-return function outside or inside a clutch switching valve. 26. Transmission according to one or more of claims 1 to 32, characterized
characterized in that the housing (1) has a housing opening (22) lying transversely to the shaft axes for the mounting of the differential gear (15; 58). 27. Transmission according to one of the preambles of the preceding claims, characterized
characterized in that the hydrostatic units (A and B) are designed as adjusting units, the adjustment of both units being carried out via a common adjusting cylinder or adjusting servo or adjusting gear and preferably both units (A and B) being controlled by a common adjusting piston (80). 28. Transmission according to one or more of the preceding claims, characterized
characterized in that a bypass valve (84) is provided which is connected between the two working pressure lines of the hydrostatic units (A and B), which produces a bypass function, particularly in the case of transmission designs without range-shift clutches, particularly when the vehicle is at a standstill and/or when being towed and/or to produce a neutral or load-free operating state.