Transmission Combination, Traction Drive and Method for Controlling the Transmission Combination

Abstract

A transmission combination includes a hydrostatic transmission and a mechanical transmission having a clutch and a control device for calibrating a grinding point of the clutch. A traction drive includes the transmission combination. A method includes calibrating the clutch.

Claims

1. A transmission combination, comprising: a hydrostatic transmission that has a hydraulic pump with an adjustable expulsion volume and with at least one hydraulic motor that is configured to be supplied with pressure medium by the hydraulic pump; a mechanical transmission that is combined with the hydrostatic transmission and has a clutch; and a control device configured to calibrate a grinding point which occurs when the clutch closes, from which time torque is configured to be transmitted via the clutch, wherein the grinding point is configured to be calibrated by the control device as a function of a response of the expulsion volume or pressure medium volume flow of the hydraulic pump or a rotational speed of the hydraulic motor.

2. The transmission combination according to claim 1, wherein: the at least one hydraulic motor is two hydraulic motors that are configured to be supplied with pressure medium via the hydraulic pump, the drive shafts of which are configured to be connected in a rotationally fixed fashion via a clutch of the transmission for the purpose of power compounding, and the grinding point is configured to be calibrated by the control device as a function of a response of the expulsion volume or pressure medium volume flow of the hydraulic pump or a rotational speed of a hydraulic motor of the two hydraulic motors.

3. The transmission combination according to claim 1, wherein the hydraulic pump is configured such that the expulsion volume depends on a setting of a closed-loop control device of the hydrostatic transmission and on a working pressure of the pressure medium.

4. The transmission combination according to claim 1, wherein a setting of a closed-loop control device is configured to be kept constant by the control device for the purpose of calibration.

5. The transmission combination according to claim 1, wherein a rotational speed of the hydraulic pump is configured to be kept constant by the control device for the purpose of calibration.

6. The transmission combination according to claim 1, wherein the hydraulic motor is configured with an adjustable expulsion volume or with a constant expulsion volume, and wherein the expulsion volume of the hydraulic motor is configured to be kept constant by the control device or is constant for the purpose of calibration.

7. A traction drive, comprising: a hydrostatic transmission that has a hydraulic pump with an adjustable expulsion volume and with at least two hydraulic motors that are configured to be supplied with pressure medium by the hydraulic pump; and a drive machine configured to drive the hydraulic pump, wherein the hydraulic motors have respective drive shafts that are configured to be coupled to at least one wheel, one chain, or one axle of the traction drive in order to transmit torque.

8. A method for calibrating a grinding point of a transmission, the transmission including a hydrostatic transmission that has a hydraulic pump with an adjustable expulsion volume and with at least one hydraulic motor that is configured to be supplied with pressure medium by the hydraulic pump, a mechanical transmission that is combined with the hydrostatic transmission and has a clutch, and a control device configured to calibrate a grinding point which occurs when the clutch closes, from which time torque is configured to be transmitted via the clutch, the method comprising: calibrating the grinding point as a function of the response of the expulsion volume or pressure medium volume flow of the hydraulic pump or the rotational speed of a hydraulic motor.

9. The method according to claim 8, wherein calibrating the grinding point as a function of the response of the expulsion volume or pressure medium volume flow of the hydraulic pump or the rotational speed of the hydraulic motor comprises: continuously detecting and/or determining the expulsion volume or pressure medium volume flow of the hydraulic pump or the rotational speed of the hydraulic motor, activating the clutch in the closing direction with a control signal starting value, changing the control signal value in order to increase a closing force of the clutch, aborting in the case of a control signal value at which a significant response is detected, and storing this control signal value as a control signal value at the grinding point in the control device.

10. The method according to claim 9, wherein before changing the control signal value, the drive shaft of a second hydraulic motor is defined.

11. The method according to claim 9, wherein before changing the control signal value, the expulsion volume of the hydraulic motor is reduced.

12. The method according to claim 9, wherein changing the control signal value takes place continuously or incrementally.

13. The method according to claim 9, further comprising repeating the sequence according to claim 9 at least once, wherein before activating the clutch in the closing direction with the control signal starting value in the repeated sequence, the method further includes setting a new control signal starting value as a function of the most recently saved control signal value at the grinding point.

14. The method according to claim 13, wherein in setting the new control signal starting value as a function of the most recently saved control signal value at the grinding point, the new control signal starting value is calculated from the most recently saved control signal value at the grinding point reduced by one or more of: the increment of the control signal value of the previous step sequence, or a fraction or a multiple thereof, a hysteresis of the control signal value, and a tolerance value.

15. The method according to claim 13, further comprising: determining a mean value from the saved control signal values at the grinding point.

16. The method according to claim 15, further comprising: determining control signal values at the grinding point which lie outside a specified bandwidth or a standard deviation, followed by one of: excluding control signal values at the grinding point from the mean value formation that lie outside the bandwidth or standard deviation, or rejecting the control signal values at the grinding point and repeating the method.

17. The transmission combination according to claim 1, wherein the transmission combination is for a traction drive.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

[0053] An exemplary embodiment of a traction drive according to the disclosure with a transmission combination according to the disclosure as well as diagrams of an exemplary embodiment of a driving according to the disclosure are illustrated in the drawings. The disclosure will now be explained in more detail on the basis of the figures of these drawings.

[0054] In the drawings:

[0055] FIG. 1 shows an exemplary embodiment of a traction drive having the transmission combination in a schematic illustration,

[0056] FIG. 2 shows a characteristic diagram of a hydraulic pump of the traction drive according to FIG. 1, and

[0057] FIG. 3 shows time profiles of activation variables of the transmission combination according to FIGS. 1 and 2, during a method for calibrating the grinding point of a clutch of the transmission combination.

DETAILED DESCRIPTION

[0058] According to FIG. 1, a traction drive 1 has a hydrostatic transmission 2 with a hydraulic machine 4 which operates as a hydraulic pump in the traction mode of the traction drive 1 and two hydraulic machines 6 and 8 which operate as hydraulic motors in the specified traction mode. The two hydraulic motors 6, 8 are connected, on the one hand, to the hydraulic pump 4 via working lines 10 and 12, and, on the other hand, 14 and 16 fluidically in parallel in a closed hydraulic circuit. The hydraulic machines 4, 6, 8 can be adjusted in their expulsion volume, while they are each embodied as an axial piston machine with a swashplate design or/axle design.

[0059] The hydraulic pump 4 is connected via its drive shaft 18 to a drive machine 20 which is configured as a diesel engine. A first hydraulic motor 6 of the two hydraulic motors 6, 8 has a first drive shaft 22, and the second hydraulic motor 8 has a second drive shaft 24. A compounding transmission 26 with two input shafts 28 and 30 is connected downstream of the hydrostatic transmission 2 as a mechanical transmission. The first input shaft 28 is connected here in a rotationally fixed manner to the first drive shaft 22, and the second input shaft 30 is connected in a rotationally fixed fashion to the second drive shaft 24. An output shaft 32 of the compounding transmission 26 is connected in a rotationally fixed fashion to a differential 34 of a drive axle 36.

[0060] The compounding transmission 36 comprises a clutch 38 which is embodied as a multi-disk clutch. The latter has a first clutch section 30 which is connected in a rotationally fixed fashion to the first input shaft 28. It additionally has a second clutch section 42 which is connected in a rotationally fixed fashion via a gearwheel arrangement 44 (illustrated only schematically) to the second input shaft 30 of the compounding transmission 26. By activating the clutch 38, which involves engagement of the second clutch section 42, the two input shafts 28 and 30 therefore the two drive shafts 22 and 24 can be connected to one another in a rotationally fixed fashion.

[0061] An activation element 46, which is configured as a hydraulic cylinder, is provided for activating the clutch 38. The piston 48 of said activation element 46 is coupled in a tension-resistant and shear-resistant fashion via a piston rod to the second clutch section 42. The hydraulic cylinder 46 has a piston space in which a compression spring 50 is arranged. The piston space is connected here permanently to a fuel tank T via a fuel tank line. On the piston rod side, the hydraulic cylinder 46 has an annular space 52 which is connected via a control pressure line 54 to a connection S of a closed-loop control device 56 which is embodied as a pressure-regulating valve. Said closed-loop control device 56 has a pressure connection P which is connected via a pressure line 58 to a feed pump 61 which is driven, together with the hydraulic pump 4, by the drive shaft 18. The feed pump 61 sucks in pressure medium from the fuel tank T. Furthermore, the pressure-regulating valve 56 has a fuel tank connection T which is connected to the fuel tank T. The pressure-regulating valve 56 is continuously adjustable and has two end positions a, b. In a first end position a, in which the valve body is pre-loaded by means of a spring 60, the pressure connection P is connected to the control pressure connection S, and the connection T is shut off with respect to the connection S. The pressure-regulating valve 56, to be more precise its valve body can be activated into a second end position b via an electromagnet 62. During the energization of the latter, and if the second end position P is completely occupied, the control pressure connection S is connected to the fuel tank connection T, and the pressure connection P is shut off. In the first end position a, charging or filling of the angular space 52 with pressure medium takes place exclusively, whereas in the second end position b exclusively emptying or discharging of pressure medium from the annular space 52 takes place. Regulating positions of the valve body are therefore possible in the two end positions a and b, in which regulating positions the connections P, S and T have a respective pressure medium connection to one another. In order to feed back the control pressure in the annular space 52, which is to be regulated and is present at the control pressure connection S, the annular space 52 is fluidically connected via a control line or a control duct to a control face of the valve body of the pressure-regulating valve 56, which is equivalent to the spring 60.

[0062] The mechanical transmission 26 also has a first rotational speed-detection unit 64 via which the rotational speed of the first input shaft 28 and therefore the first clutch section 40 and of the first drive shaft 22 can be detected. The rotational speed of the second clutch section 42 can be detected via a second rotational speed-detection unit 66 of the mechanical transmission 26, and therefore indirectly the rotational speed of the second input shaft 30 and of the second drive shaft 24 can be detected, given knowledge of the transmission ratio of the gearwheel arrangement 44.

[0063] A closed-loop control device 70, which interacts with an adjustment device 72 to adjust the expulsion volume of the hydraulic pump 4, is assigned to the hydraulic pump 4. The first hydraulic motor 6 and the second hydraulic motor 8 have a closed-loop control device 74 or 78 and an adjustment device 76 or 80.

[0064] The drive machine 20, the closed-loop control devices 70, 74 and 78, the solenoid 62 and the rotational speed-detection units 64, 66 are each connected to the control device 68 via a signal line.

[0065] FIG. 2 shows a characteristic diagram of the hydraulic machine 4 which is configured as a hydraulic pump. The relationship between the pumping pressure or working pressure p.sub.HP, a control pressure p.sub.SHP which is applied by the closed-loop control device 70 and acts on the adjustment device 72 of the hydraulic pump 4 and the specific expulsion volume v.sub.Hp, resulting as a function of the specified pressures p.sub.HP, p.sub.SHP of the hydraulic pump 4 are illustrated therein. The working pressure p.sub.HP varies here between −400 bar and +400 bar, and the expulsion volume varies between 0% and approximately 90%.

[0066] Since the hydraulic pump 4 is configured with a fully pivotable swashplate, specific expulsion volumes of +90% to −90% are covered. The positive values correspond here to a positive pivoting angle of less than 0° and the negative to a negative pivoting angle of above 0°.

[0067] It is characteristic of the already mentioned “load-sensitive” behavior of the hydraulic pump 4 that the control pressure p.sub.SHP acts in the direction of a deflection of the swashplate from its 0° position, whereas the working pressure or pump pressure p.sub.HP which is present in one of the working lines 9 or 17 is effective in the direction of the resetting of the swashplate toward the pivoting angle 0°. If the group of curves of the working pressure or pump pressure p.sub.HP is considered, their non-linear and discontinuous profile is apparent. The discontinuity is due to the fact that in addition to the adjustment force of the adjustment device 72 which results from the control pressure p.sub.SHP and the supporting force resulting from the working pressure p.sub.HP the working piston on the swashplate also acts on a spring packet on the swashplate which centers the swashplate in its zero position.

[0068] The diagram according to FIG. 2 is valid for a rotational speed, constituting the exemplary embodiment shown, of 2000 rpm of the hydraulic pump 4. This characteristic diagram and further rotational-speed-dependent characteristic diagrams of the same type are saved in the control device 68.

[0069] If, for example, a rotational speed of the drive machine 20 of 2000 rpm is predefined by means of the control device 68, the diagram according to FIG. 2 applies. For the explanation thereof, it is assumed that owing to the load situation at the differential 34 and therefore at the transmission output shaft 32 the working pressure p.sub.HP of 200 bar is present in the working line 9 or 17. In addition it is to be assumed that the pivoting angle of the hydraulic pump 4 is in the positive range between 0 and 100%. Irrespective of whether just one of the hydraulic motors 6, 8 or both contribute to the drive power of the transmission output shaft 32, it is to be assumed that the swept volume of the hydraulic motor or motors 6, 8 is constant. According to the curve of the working pressure p.sub.HP for 200 bar, which extends on the right in the quadrant I in FIG. 2, it is then possible to read off directly which control pressure p.sub.SHP the pressure-regulating valve 70 of the hydraulic pump 4 has to comply in order, for example, to achieve a specific expulsion volume of the hydraulic pump 4 of 50%. In the specified example, this is approximately 14 bar. Correspondingly, the closed-loop control device 70 is then energized via the control device 68, for example with a control signal value I.sub.HP.

[0070] It will now be assumed, with otherwise constant operational variables, that the load at the transmission output shaft 32 rises in such a way that the load pressure p.sub.HP rises from 200 to 300 bar. If the control pressure of 14 bar were to be maintained, this would bring about, according to FIG. 2 on the basis of the 300 bar line in the first quadrant, a reduction in the expulsion volume of 50% to approximately 20%, which would be equivalent to a corresponding decrease in the output rotational speed of the hydraulic motor or motors 6, 8 and therefore to a drop in velocity of the traction drive 1.

[0071] In order, on the other hand, to maintain the expulsion volume of 50%, with the specified change in load to 300 bar according to FIG. 2 an increase in the control pressure from 14 to 16 bar is necessary (cf. top dashed curve in FIG. 2).

[0072] As mentioned, such a behavior of the hydraulic machine 4 which is configured as a hydraulic pump is also referred to as “load-sensing”. This means, in principle, that the expulsion volume or delivery volume VHP of the hydraulic pump 4 changes as a function of the working pressure or pump pressure p.sub.Hp with otherwise constant setting of the closed-loop control device/of the pressure-regulating valve 70 (p.sub.SHP). The hydraulic pump 4 which is configured in such a way therefore has the property that the rotational speed or the rotational speeds of the hydraulic motor or motors 6, 8 cannot be predefined rigidly, since the expulsion volume of the hydraulic pump 4 results from the above-mentioned equilibrium of the control pressure p.sub.SHP and working pressure p.sub.HP. This characteristic is utilized beneficially for the method for calibrating the grinding point of the clutch 38.

[0073] FIG. 3 shows the time profile of relevant operational variables of the traction drive 1 during a calibration method according to the disclosure for determining the grinding point of the clutch 38 according to FIG. 1. The time t is plotted from left to right in FIG. 3. In an upper quarter of the diagram, on the one hand the profile of the breaking force F.sub.B of a parking brake of the traction drive 1 is shown between 0-100% (not shown in FIG. 1). Furthermore, the upper quarter shows the selected value of the direction-of-travel lever with F for forward travel, R for reverse travel and N for neutral. When reverse travel F is selected, the expulsion volume V.sub.HP is deflected here from 0 generally in the positive direction, whereas when the reverse travel R is selected deflection generally takes place in the negative direction. In contrast, when neutral N is selected, any accelerator pedal request of an operator is ignored, with the result that the expulsion volume V.sub.HP cannot be influenced by the activation of the accelerator pedal. This is a necessary precondition for the fact that the calibration process remains uninfluenced by any random activation by the accelerator pedal by the operator.

[0074] In the fourth quarter in FIG. 3 from the top, control signal values I.sub.HM1 and I.sub.HM2 of the closed-loop control devices 74 and 78 of the two hydraulic motors 6 and 8 are illustrated. Furthermore, the control signal value I.sub.HP of the closed-loop control device 70 of the hydraulic pump 4 is illustrated. All three specified control signal values are illustrated here as currents I between 0% and 100% on the basis of the electromagnetic activation of the closed-loop control devices 70, 74 and 78. In the same diagram the rotational speed n.sub.mot of the drive machine 20 can be read off on the right, said rotational speed n.sub.mot being, for example, 800 rpm during the calibration.

[0075] In the third quarter of the diagram according to FIG. 3, the profile of the control signal value I.sub.V for the energization of the solenoid 62 of the pressure-regulating valve 56 is illustrated. This control signal value I.sub.V represents here the input signal during the calibration method.

[0076] In the fourth, bottom section of the diagram according to FIG. 3, the profile of the rotational speed m.sub.HM2 of the second hydraulic motor 8 according to FIG. 1 is illustrated as a (system) response to the input signal I.sub.V.

[0077] The traction drive 1 according to FIG. 1 is outside a roller test stand in the normal driving mode. Owing to a regular maintenance cycle or if the operator of the traction drive 1 has the impression that recalibration of the clutch 38 is necessary, the driver/operator can initialize the calibration by means of the control device 68. For this purpose, according to FIG. 3 the parking brake of the traction drive 1, by means of which the axle 36 according to FIG. 1 can be arrested, must firstly be set in preparation to a sufficiently large braking force F.sub.B. The setting is set advantageously to the neutral value N in order to calibrate the direction of travel lever.

[0078] The calibration function is then selected by the operator. In contrast to the exemplary embodiment shown, the braking force F.sub.B and the neutral value N of the direction of travel lever can also be set in an automated fashion, for example by means of the selection of the calibration function.

[0079] In the diagram according to FIG. 3, the selection of the calibration function is at the time t.sub.0. This signal involves the control device 68 setting the expulsion volume V.sub.HM1 of the first hydraulic motor 6 to its maximum value of 100% by means of a maximum control signal value I.sub.HM1 which is transmitted to the closed-loop control device 74. At the same time, the closed-loop control device 78 of the second hydraulic motor 8 receives a control signal value I.sub.HM2 of approximately 20% from the control device 68, with the result that the expulsion volume V.sub.HM2 thereof rises to approximately 20%. With the control signal value n.sub.Mot the control device 68 adjusts the rotational speed n.sub.Mot of the drive machine 20 constantly to approximately 800 rpm. Shortly after the time t.sub.0 the traction drive 1 is therefore in the stationary state with a braking force F.sub.B of the parking brake of approximately 90%, the direction of travel switch in the neutral position N, an expulsion volume V.sub.HM1 of the first hydraulic motor 6 of 100%, a rotational speed n.sub.Mot of the drive machine 20 and of the hydraulic pump 4 of 800 rpm. A control signal value I.sub.HP is still 0% at the time just after t.sub.0. This means that the closed-loop control device 70 of the first hydraulic pump 4 does not control the adjustment device 72, and the swashplate is in its zero position at a pivoting angle of 0°, that is to say V.sub.HP=0. Accordingly, a pressure medium volume flow Q of the hydraulic pump 4 is also zero. This also entails the fact that the hydraulic motors 6, 8 are stationary despite their already set expulsion volumes of V.sub.Hm1=100% and V.sub.HM2 of 20%. Accordingly, the rotational speed-detection units 64, 66 also detect a rotational speed of 0.

[0080] Starting from a time t.sub.1, the expulsion volume V.sub.HP of the hydraulic pump 4 is then increased by increasing the control signal value I.sub.HP up to the time t.sub.3. The expulsion volume V.sub.HP then has a value of approximately 40% of its maximum. It is also apparent here from FIG. 3 that the rotational speed n.sub.HM2 of the second hydraulic machine 8 rises in parallel with the increasing of the expulsion volume V.sub.HP.

[0081] At the time t.sub.3 the entire traction drive 1 is in a stationary state which is necessary for the calibration. At the time t.sub.3 the control signal value I.sub.V for the pressure-regulating valve 56 is equal to 0. That is to say the solenoid 62 according to FIG. 1 is non-energized, as a result of which the pressure-regulating valve 56 is prestressed into its first end position a by means of the spring 60. Accordingly, pressure medium is applied to the annular space 52 via the feed pump 61, for which reason the second clutch section 42 is disengaged.

[0082] In the text which follows, according to the disclosure the grinding point of the clutch 38 is determined by means of the control device 68 as a function of a response of the rotational speed n.sub.HM2 of the second hydraulic motor 8 to the control signal value I.sub.V with which the closing force of the clutch 38 is increased. According to FIG. 3, for this purpose at the time t.sub.4 the control signal value I.sub.V is set to a starting value I.sub.V0,0 and the solenoid 62 is energized. This energization is equivalent to the start of a regulating mode of the pressure-regulating valve 56, since now the signal value I.sub.V0,0 acts on the one side of its valve body and the pressure equivalent of the spring 60 and the pressure at the control pressure connection S act on the other side. A regulating position of the pressure-regulating valve 56 between the two end positions a and b is therefore obtained, as a result of which pressure medium is relieved from the annular space 52 to the fuel tank T via the connection S and the connection T. Accordingly, the piston 48 is retracted, moved by the spring 50, in the direction of the grinding point. The starting valve I.sub.V0,0 is selected here to be below an estimated value at the grinding point, with the result that the latter still cannot be reached at this control signal value. For this reason, subsequently an incremental increase in the control signal value I.sub.V with an increment size of s according to FIG. 3 occurs. Approximately at time t.sub.5, the two clutch sections 40 and 42 according to FIG. 1 then enter into frictional engagement, with the result that a torque M.sub.HM2, albeit only a small one, is transmitted.

[0083] The torque M.sub.HM2 at the second drive shaft 24 is calculated from the product of the working pressure p.sub.HP of the hydraulic pump 4 and the expulsion volume V.sub.HM2 of the second hydraulic motor 8. A relatively high working pressure p.sub.HP therefore results from the frictional engagement of the two clutch sections 40, 42 owing to the expulsion volume V.sub.HM2, set to a constant value, of the second hydraulic motor 8 (proportional to I.sub.HM2). However, according to FIG. 3 the closed-loop control device 70 of the hydraulic pump 4 continues to be energized with the constant control signal value I.sub.HP, and the resulting control pressure P.sub.SHP of the adjustment device 72 therefore also remains constant. However, as is shown on the basis of the explanation according to FIG. 2, a control pressure p.sub.SHP which remains constant as the working pressure p.sub.HP rises brings about a reduction in the expulsion volume V.sub.HP of the hydraulic pump 4. Accordingly, when the rotational speed n.sub.HP remains the same, the pressure medium volume flow Q thereof drops, which subsequently leads to a reduction in the resulting rotational speed n.sub.HM2 of the second hydraulic motor 2, according to FIG. 2.

[0084] If this reduction in the rotational speed of the second hydraulic motor 8 reaches a significant value at which the reduction in the rotational speed is greater than the noise of the detection signal of the rotational speed m.sub.HM2 by at least a factor of 2, the control signal value I.sub.V which is then present is saved as a control signal value at the grinding point I.sub.V,cal1 in the control device 68.

[0085] With this first calibration step, the calibration could be terminated since it is now known for a known input signal I.sub.V,cal1 that the two clutch sections 40, 42 have entered into frictional engagement near to the grinding point. However, it proves advantageous to repeat the step sequence carried out in this way for the purpose of calibration. This is shown by FIG. 3, according to which the described step sequence is repeated four times with respectively newly calculated starting values I.sub.V0,1 to 4. The respectively following starting value of the control signal I.sub.V0,i is for this purpose calibrated by means of the control device 68 from the previous control signal value at the grinding point I.sub.v,cal,i, reduced by the step s, a valve hysteresis of the pressure-regulating valve 56 and a tolerance value. An exemplary value for the step s is then 20 mA, for the valve hysteresis 30 mA and for the tolerance value 10 mA.

[0086] In the exemplary embodiment of the method, after five step sequences with five determined control signal values at the grinding point I.sub.v,cal,1-5, the mean value thereof is calculated as I.sub.v,cal,m. The calibration ends with the resetting of the values F.sub.B, I.sub.HP, I.sub.HM2, I.sub.V to their output values 0 by means of the control device 68.

[0087] In the exemplary embodiment (illustrated in FIG. 1) of a transmission combination according to the disclosure, the hydrostatic transmission has two hydraulic motors and the mechanical transmission has, apart from various gearwheels and other elements, a single clutch. Of the two hydraulic motors, the one hydraulic motor is continuously operationally connected to the output shaft 32, while the other hydraulic motor is connected to the output shaft only when the clutch is closed.

[0088] A transmission combination according to the disclosure can also comprise a hydrostatic transmission with a different number of hydraulic motors or a mechanical transmission with a different number of clutches. It is therefore possible, for example, for the hydrostatic transmission to comprise two hydraulic motors and for the mechanical transmission to comprise three clutches, wherein when the clutches 1 and 2 are closed both hydraulic motors are operationally connected to an output shaft, and when the clutch 2 is closed only the one of the two hydraulic motors is connected to the output shaft, and when the clutch 3 is closed the other of the two hydraulic motors is operationally connected to the output shaft with a different down step ratio than when the clutch 1 is closed.

[0089] It is also conceivable that just a single hydraulic motor is present, by which a sun gear of a planetary gear mechanism can be driven. The planetary gear mechanism also comprises a planetary carrier with planetary gears and a ring gear. The ring gear can be connected in a rotationally fixed fashion to the sun gear via a clutch or in a rotationally fixed fashion to a frame via a further clutch. Depending on which clutch is closed, the down step ratio from the drive shaft of the hydraulic motor to the shaft of the planetary carrier is different.

[0090] In this context, it is to be noted that an arrangement which can be activated and which can connect a rotatable element in a fixed fashion to a frame or housing and therefore can be stationary with respect to the frame is generally referred to as a brake. The term clutch which is used in the description is also intended to comprise such arrangements which are generally referred to as a brake.

[0091] A transmission combination is disclosed having a hydrostatic transmission which has a hydraulic pump with an adjustable expulsion volume and with at least one hydraulic motor which can be supplied with pressure medium by the hydraulic pump, and with a mechanical transmission which is combined with the hydrostatic transmission and has a clutch. A control device of the transmission combination is configured here in such a way that it can be used to determine a grinding point of the clutch as a function of the response of an, in particular, kinetic, operational variable of the transmission, dependent on the expulsion volume of the hydraulic pump, to a torque at the grinding point. The kinetic operational variable can be, in particular, the pressure medium volume flow of the hydraulic pump or a rotational speed which is dependent thereon. The response of the expulsion volume of the hydraulic pump itself can also be used as an operational variable for this purpose.

[0092] Furthermore, a traction drive and a vehicle, in particular a mobile working machine, with such a transmission combination are disclosed.

[0093] Furthermore, a method for calibrating the transmission combination with a step in which the grinding point of the clutch is determined as a function of the response of the, in particular kinetic, operational variable of the transmission, dependent on the expulsion volume of the hydraulic pump, to the torque which occurs at the grinding point is disclosed.

LIST OF REFERENCE SYMBOLS

[0094] 1 Hydrostatic traction drive

[0095] 2 Hydrostatic transmission

[0096] 4 Hydraulic pump

[0097] 6 First hydraulic motor

[0098] 8 Second hydraulic motor

[0099] 10, 12, 14, 16 Working line

[0100] 18 Driveshaft

[0101] 20 Drive machine

[0102] 22 First driveshaft

[0103] 24 Second driveshaft

[0104] 26 Compounding transmission

[0105] 28 First input shaft

[0106] 30 Second input shaft

[0107] 32 Output shaft

[0108] 34 Differential

[0109] 36 Drive axle

[0110] 38 Clutch

[0111] 40 First clutch section

[0112] 42 Second clutch section

[0113] 44 Gearwheel arrangement

[0114] 46 Hydraulic cylinder

[0115] 48 Piston

[0116] 50 Spring

[0117] 52 Annular space

[0118] 54 Control pressure line

[0119] 56 Pressure-regulating valve

[0120] 58 Pressure line

[0121] 60 Spring

[0122] 61 Feed-in pump

[0123] 62 Solenoid

[0124] 64 First rotational speed-detection unit

[0125] 66 Second rotational speed-detection unit

[0126] 68 Control device

[0127] 70 Closed-loop control device hydraulic pump

[0128] 72 Adjustment device hydraulic pump

[0129] 74 Closed-loop control device first hydraulic motor

[0130] 76 Adjustment device first hydraulic motor

[0131] 78 Closed-loop control device second hydraulic motor

[0132] 80 Adjustment device second hydraulic motor

[0133] F.sub.B Force brake pedal

[0134] F Forward travel

[0135] N Neutral

[0136] R Reverse travel

[0137] I.sub.HP Control signal value hydraulic pump

[0138] I.sub.HM1 Control signal value first hydraulic motor

[0139] I.sub.HM2 Control signal value second hydraulic motor

[0140] n.sub.mot Rotational speed drive machine

[0141] I.sub.V Control signal value clutch

[0142] I.sub.V0 Control signal starting value

[0143] I.sub.V0,i Control signal starting value calculated

[0144] I.sub.V,cal Control signal value at grinding point

[0145] n.sub.V,cal,m Mean value control signal value at grinding point

[0146] n.sub.HM1 Rotational speed of first hydraulic motor

[0147] n.sub.HM2 Rotational speed of second hydraulic motor

[0148] n.sub.HM2,cal Rotational speed of second hydraulic motor at grinding point

[0149] P Pressure connection

[0150] S Control pressure connection

[0151] T Fuel tank connection