DE29816863U1 - Continuously variable transmission, especially with power split - Google Patents

Description

Michael Meyerie &Iacgr; ··. : : 'j '·„·'·„·

Engineering office for J , * « &iacgr; · « *

Research and Development · ·· _&iacgr;

Automotive transmission

Kiefernweg 9 ■ Lochbrücke

D-88074 Meckenbeuren

a 3? 2 he

G122H35

Continuously variable transmission, especially with power split

The invention relates to a continuously variable transmission, preferably with power split, with a hydrostatic or mechanical continuously variable converter according to the preamble of claim 1 and further independent claims.

The object of the invention is to create a power split transmission that can be installed in a main housing or in a vehicle frame, e.g. of a tractor, in a simple and time-saving manner without dismantling the vehicle main housing or vehicle frame. In addition, different power sizes should be possible with a hydrostatic or mechanical variator (converter 4c, 4d) that is as similar in size as possible. The object is achieved by the features listed in the main claims. Further details can be found in the subclaims and the description. They show in a schematic representation:
Fig.l and 2, Fig.7 to 14 and 19 to 20 different embodiments of the

continuously variable hydrostatic power split transmission with two
Forward driving ranges in inline arrangement of the components hydrostatic transmission,
Summation planetary gear and, if necessary, reverse drive device.
Fig. 15 to 18 different embodiments of the reversing device.

Fig.21 to 25 Gearbox designs with more than two forward shift ranges.

Fig.26 to 29 the arrangement of the components for a mechanical stepless

Power split transmission.

Fig. 30 to 33 Gearbox with two forward ranges.

Fig. 34 Speed and function plan for automatic group switching.

Fig.35 to 44 Gear concepts according to the invention as mechanical

Power split transmission.

The transmission according to the invention can be designed in various ways and is characterized

in particular, the power split transmission (HVG / MVG) forms a complete unit which, in the form of a modular design, can be used in any housing shape of an overall transmission or engine or frame of a vehicle. Depending on the vehicle requirements, the transmission can be designed as a hydrostatic-mechanical power split transmission (HVG) or a mechanical power split transmission (MVG) with a different number of driving ranges for forward and reverse travel. It can be used as a standard transmission for various types of vehicle, such as tractors, work machines, commercial vehicles, distribution vehicles or buses, etc. A special feature is that a transmission family is created for a larger power range, e.g. for tractors, whereby, for example, the same hydrostatic transmission or the same-sized converter 4c; 4d can be used for a power range of 70 to 300 hp. The delimitation of the individual power ranges is achieved by a corresponding number of shift ranges and/or assigned group gears and/or adaptation gears at the gear input, whereby, for example, only one forward range is used for the lowest power range and four or more forward ranges are used for the highest power range. The number of ranges and division of the range sizes determine the gearbox corner power required for the respective tractive force. The gearbox (HVG) or (MVG) is preferably constructed in an inline design. This means that the hydrostatic gearbox 4c and the coupling gearbox 5c, which contains the summation planetary gear and, if applicable, the range clutch, are arranged coaxially to one another. Depending on the vehicle concept, the drive shaft Ic can advantageously be arranged on the same axis as the drive shaft of the drive motor. In order to be able to fully utilize the power capacity of the hydrostatic transmission 4c, the invention provides a high-drive transmission (HT) connected upstream of the hydrostatic transmission 4c, preferably designed as a planetary transmission, which allows the speed of the input shaft Ic or the hydrostatic transmission to be adjusted to its permissible values (Fig. 31). For example, this transmission (HT) can enable the maximum drive speed of 2,300 rpm on the input shaft lca to the permissible input speed of 3,500 rpm of the hydrostatic transmission on the input shaft Ic. The high-drive transmission (HT) can be designed as a known transmission transmission with spur gear stages (not shown) or, as shown in Fig. 31, more advantageously as a three-shaft planetary transmission, with the sun gear S being fixed to the housing, the web St with the input shaft lca and the ring gear H forming the output shaft. According to the increase in speed between the two shafts, gearbox input shaft lca and hydrostatic input shaft Ic, the gearbox corner power and thus

The maximum tractive force is increased accordingly based on the same final speed of the vehicle.

Another possibility for adjusting the necessary traction is provided by the use of a group transmission (GR) (Fig. 32), whereby, for example, a field group and a road group are provided in the known manner with the switching positions A and S as shown in Fig. 32. The field group is designed for a maximum speed of 30 km, for example, and the road group for a maximum speed of 50 km.

The power split transmission (HVG or MVG) can therefore, e.g. depending on the required tractive force or depending on the power of the vehicle, e.g. tractor, be combined with the high-driver HT (Fig. 31) and/or a slow/fast or field/road group GR (Fig. 32). The power split transmission HVG used in each case is designed as a single-range, two-range, three-range or four-range transmission Fig.21; 22, whereby the hydrostatic transmission 4c advantageously forms a largely uniform structural unit for all power sizes from the same basic components A and B or the complete hydrostatic transmission 4c.

The gearbox design according to Fig. 30 includes the power-split basic gearbox HVG with two forward driving ranges without its own reverse range. According to the invention, this gearbox HVG is assigned a step gearbox VG/GR, which has a reversing group WG for speed reversal with the corresponding switching R. This step gearbox VG is also equipped with a step manual gearbox, which in the known manner allows a step switching for slow operation L or A and fast operation S or H, e.g. field or road operation of a tractor, as explained in more detail in the output speed diagram Fig. 34. The field level here is designed for a maximum of 30 km/h, for example, and the road group for 50 km/h. The continuously variable branching gearbox HVG or MVG can be created from the basic units of the overall gearbox program.

The transmission program provides that, depending on the various vehicle requirements, particularly with regard to the traction requirements, the respective transmission design - single-range, two-range, three-range, four-range transmission - can be combined with the high-driver transmission HT (Fig. 31) arranged on the input side and/or with a downstream group transmission GR (Fig. 32; 33). When using the group transmission GR, the respective group - field group A or road group S - can be preselected when the vehicle is at a standstill.

The invention provides a control and regulating device which makes it possible to

to switch from one group to the other while driving. The control system has a special group switching program for this purpose, which provides that, as shown in Fig. 34, at the speed or ratio end of the working group or at the ratio/speed point PAl, the group switching is set to neutral and within a time phase dependent on an adjustment speed, the gear ratio is reduced until synchronous operation of the clutch elements of the road group is achieved, after which the road group is automatically switched on. The synchronization point of the relevant clutch elements is sought by a preferably electronic speed comparison of suitable gear elements using speed sensors or other known devices. The group switching S-A is therefore automated accordingly, whereby the clutch itself can be designed as a friction clutch or a positive-locking clutch. A hydraulically actuated positive-locking clutch is advantageous here, preferably with deflection teeth as described in more detail in European Patent Specification 0276 255. A known multi-disk clutch or cone clutch, as shown in DE19 14 724 in Fig. 42, 43, 44, can be used as a friction clutch. When using friction clutches, it is sensible to provide a clutch overlap within the switching phase or within the necessary gear ratio change.

The invention provides an alternative, automatically effective control program for the automatic switching process from working group A to road group S and vice versa from S to A. This program can be activated automatically depending on one or more operating parameters and/or depending on predefined time parameters and/or depending on factors that determine economic efficiency, such as transmission efficiency and/or engine efficiency. For example, a switch from one group to another can take place when the control device recognizes that an operating state can be operated in the other switching group with lower fuel consumption and/or with better noise behavior. For example, at a driving speed of 25 km/h (see Fig. 34) and working group A is engaged, the control device will recognize that this operating state can be operated in road group S with lower fuel consumption. The control program provides for an automatic switch from group A to group S to be triggered in the manner described above. The trigger signal can also be triggered manually using a corresponding operating device (button; lever). In the case of automatic triggering, it is advantageous to include a time factor in the program, i.e. that the switching process only takes place after a predetermined dwell time at the corresponding transmission point.

or can be triggered within a limited gear ratio range and/or a constant speed and/or constant load values or operating values. In order to largely prevent load interruption or to reduce the interruption period to a minimum, it is advisable to use the group switching via friction clutches or positive clutches with deflection teeth, in accordance with the above-mentioned EP patent.

When shifting up from A to S, the first range can be fully extended, e.g. up to 30 km/h, as shown in Fig. 34, after which an automatic switch to S takes place to increase the speed further. When shifting down, the signal to switch from group S to group A will, in the worst case, only occur at a speed point PS2 to PA2 if there is a minimum gear ratio difference Ai, which prevents the required output speed from exceeding the final speed point PAl. However, a switch to the other group can also take place at lower speeds or in the lower gear ratio range, e.g. at 15 km/h, if the drive control device recognizes that this speed point can be driven more efficiently in the other gear group. For this purpose, the corresponding transmission and engine characteristics are programmed into the electronic control device 5, from which the switching signal is generated depending on the respective transmission ratio and the respective load condition, e.g. hydrostatic pressure, ratio, from which the respective hydrostatic power component can also be determined, and possibly other operating values. The continuously variable transmission with the group switching described above can be used for both work machines and road vehicles of various types.

Switching to the other range preferably takes place after a defined dwell time within a defined gear ratio range in order to avoid switching back and forth from one gear range to the other too frequently. The appropriate values must be determined experimentally. For example, in a transport operation at 25 km/h, switching to road group S could only be triggered after a dwell time of approx. 30 seconds. After a switching process has been triggered, the next switching process should ideally take place after a longer dwell time. A load-dependent switch from range S to A, on the other hand, should take place as spontaneously as possible in order to bring the specific load values, e.g. hydrostatic pressure, to correspondingly lower values.

The invention further provides that an optical and/or acoustic display is provided, which indicates for the respective switched group whether this operating state in

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this group or whether it is better to drive in the other group. The driver could be informed via a corresponding light signal and/or monitor and/or acoustic display, e.g. verbal request or indication, whether a group change makes sense. Depending on the type selected, the clutches - whether force-locking or form-locking clutches - the switching devices, as known from powershift transmissions or from automatically switchable multi-stage transmissions, can be used for the automatic group change.

According to the invention, the driver can decide whether the group switching should be automatic or manual by making the appropriate pre-selection using the appropriate pre-selection devices.

For the application of mechanical power split transmissions MVG, the invention provides a transmission series that provides a transmission series with two or more shift ranges to meet various vehicle requirements, in particular with regard to different power levels.

In Fig.35, 39, 40 is shown a transmission system with a belt transmission 4d and an associated power split transmission, which consists of a

Summation planetary gear 301; 101; 201; 401 with associated range clutches Kl and K2. A fully stepless driving speed from zero to top speed is achieved via two power-split shift ranges. This means that in the starting state the gear ratio is "infinite" and a starting or separating clutch between the engine and the transmission can be omitted. The MVG transmission can be equipped with various mechanical stepless variators or converters, as well as friction transmissions of any type with a primary and secondary part. Preferably a belt transmission 4d is provided, consisting of a primary unit 4dA and a secondary unit 4dB. The associated summation planetary transmission 301; 101; 201; 401 is designed with four shafts and has two input shafts El and E2 as well as two output shafts Al and A2. The first input shaft El is connected to the drive shaft Ic and the primary unit 4dA and the second input shaft E2 to the secondary unit 4dB. The variable speed thus flows and the variable power is fed via the second input shaft E2. In the summation planetary transmission both power branches of the shafts El and E2 are summed up and forwarded together alternately via the two output shafts Al and A2 to the output. The first output shaft Al of the summation planetary gear and the second output shaft A2 are thus alternately connected to the output shaft 106.

According to the invention, the summation planetary gear can also be extended in different ways for mechanical power split transmissions MVG (see 301 Fig.35; 101 Fig.38; 201 Fig.42; 401 Fig.39). All embodiments have in common that in the starting state at zero driving speed the gear ratio is "infinite", whereby a starting clutch can be saved and that the continuously variable converter is set to a large, preferably maximum, ratio, whereby the first driving range is driven through with the clutch Kl closed by increasing the speed of the second input shaft E2 from a low speed to the same speed as the first input shaft El. At this ratio point, all elements of the summation planetary gear 101; 201; 301; 401 as well as the elements of the second range clutch K2 have reached synchronous operation, after which the second shift range can follow seamlessly without load interruption by closing the second range clutch K2 and opening the first range clutch Kl.

The gearbox designs according to Fig.35, 38, 39, 40 each have two power-split forward driving ranges. As shown in Fig.38, these gearboxes are designed with a device for a reverse driving range. For this purpose, a planetary gear stage PR is provided, in which, for example, the ring gear of the planetary gear PR is connected to the first output shaft Al, the sun gear is connected to the output shaft 106 and the web can be connected to the gearbox housing via a coupling KR. Instead of this planetary gear PR, various reverse driving devices, such as those shown in Fig. 15, 16, 17 or 18, can be used. Furthermore, in the gearbox designs Fig.38 and 39, the drive shaft Ic is guided through the gearbox in order to be connected to the other shaft end 114; 2e to enable the possibility of a power take-off shaft connection, e.g. on a tractor or PTO connection. Furthermore, alternatively, the shaft 114; 2e can also be used as a drive shaft, whereby according to certain vehicle requirements, e.g. when used in a car with front-wheel drive, the transmission output gear 138 can be arranged on the transmission input side in a vehicle-friendly manner.

The summation planetary gear 301 Fig.35 has two planetary gear stages Pl and P2, whereby the carrier shaft 126 of the first planetary gear stage forms the first input shaft El and is in drive connection with the primary unit 4dA of the converter 4d and the drive shaft Ic as well as the ring gear 127 of the second planetary gear stage P2. The ring gear 125 of the first planetary gear stage is in drive connection with the second input shaft E2 and the primary unit 4dB of the continuously variable converter 4d. The sun gears 137 and 124 of both planetary gear stages are connected to the second output shaft A2. Intermeshing planet gears 122 and 123 are on the carrier shaft 128 of the second planetary gear stage P2.

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arranged, one 122 engaging the ring gear 127 and the other 123 engaging the sun gear 124.

The carrier shaft 128 of the second planetary gear stage P2 forms the first output shaft Al, which can be connected to the clutch Kl.

The summation planetary gear 401 according to Fig. 39 has a web 133 which is connected to the first input shaft E1. Intermeshing first planetary gears 132, second planetary gears 130 and third planetary gears 131 are arranged on this web 133, the third planetary gears 131 meshing with a ring gear 134 connected to the second input shaft E2. A second ring gear 135 connected to the first output shaft A1 meshes with first planetary gears 132 and a sun gear 136 connected to the second output shaft A2 also meshes with first planetary gears 132.

The summation planetary gear 101 according to Fig. 40 is designed such that the second input shaft E2 is connected to a web 103 on which intermeshing first planetary gears 107 and second planetary gears 108 are arranged, with a ring gear 102 connected to the first input shaft E1 engaging with first planetary gears 107. A ring gear 104 connected to the first output shaft A1 meshes with second planetary gears 108 and a sun gear 105 connected to the second output shaft A2 also meshes with second planetary gears 108.

A further embodiment of the summation planetary gear 201 according to Fig. 42 provides that the second input shaft E2 forms the planet carrier 144, on which intermeshing first planetary gears 139 and second planetary gears 140 are arranged, with a ring gear 141 connected to the first input shaft El engaging first planetary gears 139, a second ring gear 142 connected to the first output shaft Al engaging second planetary gears 140 and a sun gear 143 connected to the second output shaft A2 also engaging second planetary gears 140. In this summation planetary gear design 201, the first input shaft El is arranged above the second input shaft E2, with the drive connection between the summation planetary gear and the drive shaft Ic being made via a first spur gear stage dA and a second spur gear stage dAl.

The gearbox designs according to Fig.41 to 44 each have an additional gearbox 112; 112a; 113 to create more than two switching ranges. The additional gearboxes 112 and 112a are designed as countershaft gearboxes, with the output shaft 106 being arranged offset from the drive shaft Ic. The output shaft 106a can also be arranged coaxially to the drive shaft Ic using an adaptation stage 152. In the gearbox designs according to Fig.41, 42, 43, the power is transmitted in the first and second switching ranges via the spur gear stage 146 with alternating

Clutches K1 and K2 and in the third shift range via a further spur gear stage 145 to the output shaft 106. For a possible fourth shift range, a further spur gear stage 148 is provided, see Fig. 43, whereby the power is transmitted via the second output shaft A2 and to the output shaft when clutch K2 is closed.

In the case of the gearbox design according to Fig. 44, the associated gearbox 113 is designed in a planetary gear design according to the invention. In this gearbox design, the power summed up in the summation planetary gear is transferred to the output in the first and second shift ranges with the clutches Kl and K2 alternately engaged via a planetary gear stage P4 with the clutch/brake KV closed. In the third shift range, the power is transferred directly from the variator 4d via the second input shaft E2 to the output shaft 106 by closing the clutch K3. In the fourth shift range, with the clutch K2 closed, the power is directed to the output shaft after the clutch K4 is closed without an intermediate gear stage. For reverse operation, an additional planetary stage P3 is provided, which, with the clutch or brake KR engaged, directs the power to the output via two shift stages with the clutches Kl and K2 alternately engaged.

The additional gear 113 consists of two planetary gear stages P3 and P4, whereby the sun gears 115 and 116 are connected to an intermediate shaft 109, which transmits the power when the clutch K1 or K2 is closed. The ring gear 117 of the planetary gear stage P4 is connected to the housing via a clutch or brake KV in the first and second shift ranges. The ring gear 118 of the planetary gear stage P3 is firmly connected to the output shaft 106 and to the web 117a. The web 118a of the planetary gear stage P3 can be connected to the housing 1 via a clutch or brake KR when reversing. The intermediate shaft 109 is used to transmit power in the first, second and, if necessary, fourth shift range.
Functional description

The functional sequence is identical for all transmission versions Fig.35 to 44 in relation to the respective switching range. In the starting state with the preselected direction of travel "forward", the first output shaft Al has the speed "zero" due to the correspondingly adjusted ratio of the summation planetary gear and the ratio setting of the variator 4d. The secondary unit 4dB of the variator 4d is set to a low speed, preferably to its maximum own ratio. In this state, clutch 1 is preferably switched automatically according to the preselected direction of travel. By resetting the ratio of the

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Variator, the second input shaft E2 begins to increase the speed, whereby the first output shaft Al and thus the output shaft 106; 109; 109a are set in rotation accordingly. After reaching the preferred maximum adjustment value of the variator, the second input shaft E2 and thus all elements of the summation planetary gear 301; 101; 201; 401 as well as the elements of the clutch K2 have reached synchronous operation. The shift into the second shift range now takes place by closing the clutch K2 and subsequently opening the clutch Kl. The variator 4d can now be regulated back down to its maximum gear ratio, which corresponds to the final gear ratio point or the end of the second shift range. The power summed up in the summation planetary gear is transmitted in the second shift range via the second output shaft A2. In the case of a design with several shift ranges as shown in Fig.41 to Fig.44, a third shift range can be connected at the end of the second shift range, so that a direct drive connection is established between the secondary unit 4dB of the variator and the output shaft, e.g. by closing a clutch K3 (see Fig.44; 41; 42). In the first and second shift ranges, the power in these transmission designs is transmitted via an additional gear ratio, e.g. in the design shown in Fig.44 via the planetary gear stage P4 with the clutch KV closed. After the third shift range has been engaged via the clutch K3, the clutch KV is opened and the variator 4d is adjusted in the opposite direction again until, at the end of the third shift range, synchronization of the clutch elements K4 is achieved in accordance with the speed of the second output shaft A2 and the intermediate shaft 109 still connected to it via the clutch K2. Now the shift into the fourth gear range takes place by closing this clutch K4 and then opening the clutch K3. Now the power transmission is again split by repeatedly adjusting the variator back to the other end of the adjustment until the final gear ratio point of the gearbox is reached. For reversing, the clutch or brake KR is closed after the preselected direction of travel R, which results in a speed reversal via the planetary gear stage P3. The functional sequence for reversing takes place over two reverse ranges with the clutch Kl or K2 closed with the same functional sequence as for forward driving. For the gearbox versions with countershaft gear 112, 112a, the same functional sequence applies as for the version according to Fig. 44 with planetary gear design 113; this means that in the third shift range, the power is transmitted via a spur gear stage 145 with the clutch K3 closed and, if necessary, in the fourth shift range, via the spur gear stage 148 with the clutch K2 and K4 closed. For reverse operation, the same can be done with the clutch K1 or K2 closed within two shift ranges.

Power is transmitted to the output via a spur gear stage 147 provided with an intermediate gear 147a.

The mechanical converter or variator 4d can be arranged in different ways depending on the given spatial conditions. For example, it is intended to arrange the two units 4dA and 4dB on an axis line with the drive shaft Ic, as shown in Fig. 37 or as shown in Fig. 36, next to or above the drive shaft Ic, which in each case enables a very compact design.

The transmission according to Fig.l; 2; 4 has two forward and one reverse range. The summation planetary transmission PSl; PSl' has four shafts and has two input and two output shafts. Intermeshing planetary gears Pll' and P12' are arranged on a planet carrier Al', with a ring gear HE2 in drive connection with the second hydrostatic unit B engaging in first planetary gears Pll' and the ring gear HEl in drive connection with the drive shaft Ic; Ic' engaging in second planetary gears P12', and with the carrier shaft αΓ as the first output shaft Al' with a first clutch K1 and the second output shaft A2' forming a sun gear SA, which engages in first planetary gears Pll' and can be connected to a second clutch K2. The two output shafts, web Al and sun gear SA, can be connected to the output shaft 2c via the two clutches Kl and K2. A second, preferably three-shaft planetary gear unit PRl' is used for the reverse range, whereby its ring gear H3' can be connected to the web Al', its sun gear SR' to the output shaft 2c and its web shaft PT2' can be connected to the gearbox housing via a clutch or brake KR. This gearbox design has two hydrostatic-mechanical forward ranges and one hydrostatic-mechanical reverse range. In another, not shown, embodiment, the first output shaft - web Al' - is connected to the sun gear SR' and the sun gear SA to the ring gear H3'. This design is advantageous for vehicles with lower reversing speeds. The reversing device can be designed in different ways, as shown in Figs. 15 to 18.

In transmission designs according to Fig. 1; 2; 7 to 21, according to the invention, the hydrostatic transmission 4c' with the units A and B is coaxial with the summation planetary transmission PSl'.

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the clutches Kl and K2 and the planetary gear SR'. This has the advantage that the spur gear stages connecting the hydrostatic transmission can be omitted.

In gearbox designs according to Fig. 2; 7; 9 and others, the drive shaft Ic" is guided through the gearbox, which can be used on the output side of the gearbox as a PTO connection or as a power take-off drive or also as a drive shaft. In this design, the gearbox output takes place via an output-side spur gear stage 2d, which creates the drive connection between an axle differential DIF and the output shaft 2c'. The axle differential DIF is arranged coaxially offset to the drive shaft Ic". This gearbox is particularly suitable for work machines, such as tractors. Depending on the vehicle requirements, the spur gear stage 2d can be omitted, as shown in Fig. 3, with the output shaft 2c' representing the bevel pinion shaft for the bevel drive of the differential gear DIF. The bevel gear drive KE is designed as a hybrid drive with an axle offset AX, so that the output PTO shaft or power take-off shaft can be arranged on one of the two drive shafts TW with sufficient distance.

The summation planetary gears PSl and PSl' according to Fig.l and 2 are functionally identical and only differ in that, depending on the arrangement of the hydrostatic transmission 4c and 4c', the position of the two ring gears HEl and HE2 in the axial direction is swapped.

In the gearbox design according to Fig.4, the hydrostatic gearbox 4c is arranged offset parallel to the drive shaft Ic", whereby the drive of the first hydrostatic unit A is coaxial at the gearbox output via a spur gear stage 10b" and the drive connection of the second hydrostatic unit B is via a spur gear stage 228' arranged on the input side. The shift drum 5c, which contains the summation planetary gear and the shift clutch, is also arranged here on the drive shaft Ic", whereby the connection for the power take-off shaft or PTO can be implemented very advantageously for a tractor in the extension of this shaft 2e.

The transmission design according to Fig.6 is similar to the design according to Fig.4 and has the difference that the hydrostatic transmission 4c', which is arranged parallel to the drive shaft, is designed in such a way that the second hydrostatic unit B is arranged on the drive shaft 6c', so that both hydrostatic units A and B can be driven on the drive side, with a

first spur gear stage 10b' the first hydrostatic unit A and via an adjacent second spur gear stage 228' the drive connection between the second hydrostatic unit B and the shift drum 5c arranged on the drive shaft Ic can be established.

Instead of the summation planetary gear PSl', another functionally identical summation planetary gear PSl" according to Fig. 5 can be used according to the invention, which consists of two planetary gear sets, the first planetary gear set being connected to the first input shaft El' via a carrier shaft StI". The second input shaft E2' is connected to the ring gear H2" of the second planetary gear stage. The first output shaft Al' is firmly connected to the ring gear Hl" of the first planetary gear stage and the carrier shaft St2" of the second planetary gear stage and the second output shaft A2' is connected to the sun gear Sl' of the first and to the sun gear S2' of the second planetary gear stage.

The transmission design according to Fig.6 has a hydrostatic transmission 4c' arranged parallel to the drive shaft, which is designed in such a way that the drive connections to the first hydrostatic unit A and the second hydrostatic unit B are possible on the transmission input side. The drive shaft Ic is connected to the first hydrostatic unit A via a first spur gear stage 10b'. An adjacent spur gear stage 228' connects the second hydrostatic unit B to the summation planetary gear arranged in the shift drum 5c. The shift drum 5c with the summation planetary gear and the shift clutches is also arranged on the drive shaft Ic and can establish the connection to the power take-off shaft on a tractor or a PTO in an axial position to the drive shaft Ic in an extension (shaft 2e). The output to the differential DIF takes place via a spur gear stage 2d. The axle differential DIF and the second output shaft 2c" can be designed offset from the drive shaft, as shown in Fig.6, or in an axially identical design, as shown in Fig.3.

The gear design according to Fig.7 is similar to the gear design according to Fig.2, but with the difference that instead of the summation planetary gear PSl', the functionally equivalent summation planetary gear PSl", as shown in Fig.5, is used. Furthermore, the front wheel drive with the output shaft 2VA and the corresponding clutch KVA, e.g. for use in a tractor, is shown.

The gearbox design according to Fig.9 shows the installation of the gearbox in a modular design, e.g. for use in a tractor. The gearbox shown here corresponds to the gearbox design according to Fig. 7. The two components - hydrostatic gearbox 4c and the shift drum or gearbox drum 5c' with summation planetary gear and the clutches and if necessary the reversing planetary stage PRl' and a brake or clutch KR -

are arranged coaxially to one another and in this design are preferably inserted as a complete unit from the outside into the main housing 1, with the input shaft lca being connected to the drive shaft Ic via a connecting element Fl and the output shaft A' being connected to a PTO or power take-off shaft via a further connection F2 and an intermediate shaft 2e coupled to the drive shaft and leading through the transmission with a connecting element F3. The connections Fl, F2 and F3 are preferably mechanical connections which can be designed as a screw flange connection and/or a drive connection with a tooth profile, e.g. as a sleeve connection, among other things of a known type. For the hydrostatic transmission 4c, an elastic torque support 4e is expediently provided to support the reaction torque, as is advantageously shown and described in the as yet unpublished DE 197 27 360.2 in Fig.lba, lbb and lbc. In this case, for example, the elastic support is advantageously implemented by rubber elements 4f and 4g, which have a spherical or conical shape, inserted into recesses or drive openings in the housing and the hydrostatic transmission. The drive openings 4h' mentioned can be incorporated directly as shown or housed in a separate, non-illustrated, rotationally fixed element, e.g. a sheet metal body, as shown in the aforementioned DE in Fig. lba part 9e.

This gearbox design according to Fig.8 is very cost-effective, since the hydrostatic gearbox 4c on the drive shaft Ic and also the shift drum or transmission drum 5c' with the summation planetary gear and the shift clutches are arranged coaxially to each other. In this way, the otherwise very costly drive connections through spur gears, which are required, for example, when the hydrostatic gearbox is arranged offset on the axle, are eliminated. When used in a tractor, there is sufficient length so that this design can be used without any problems. Apart from the drive connection to the axle differential DIF via a spur gear stage 2d, no spur gear stages or gears are required up to the drive of the rear axle. The oil lines for the control oil supply for the clutches Kl and K2 and possibly also for KR as well as the lubricating oil supply, etc. can be routed in a very cost-effective and simple manner through a carrier element Ib (see also Fig. 11), which can also be coupled to the housing 1, and transferred to the corresponding components. Plug connections, e.g. with plug connection 3a Fig. 11 (not shown in detail) of a known type between a control unit le; Ie' and the carrier element Ib of the transmission module HVG establish the necessary oil connection. The plug connections mentioned are pipe connections with elastic sealing elements, e.g. O-rings at the respective connection points, also ensure noise insulation and also allow compensating movements, e.g. between the hydrostatic transmission 4c and

the housing due to the elastic torque deposit 4e. The HVG gearbox can be designed in various ways and can be a single or two-range gearbox, as shown in Fig. 1, 2, 7, 8 to 14 and 26, or a three-, four- and multi-range gearbox, whereby, for example, the internal gearbox structure or the gearbox concept as shown in Fig. 21 can be used.

The gearbox design according to Fig. 7, 9 to 14 and 1 to 3 is preferably used for tractors of the lower power class, whereby the first switching range can preferably be between 0 and 16 and the second switching range between 16 and 50 km/h. With this range division, the best efficiency, at which the hydrostatic power is zero, is at the main speed point of 8 km/h.

The function of this gear design and all gear designs Fig. 1 to 8 is identical in terms of the respective shift ranges to the gear design described in EP 0 831 252 A2 or DE 197 41 510Al according to Fig. 3, whereby it must be taken into account that in the gear designs with two forward and only one reverse range, the respective reverse range is functionally the same as the first forward range, whereby the reversal of the direction of rotation is implemented in each case by an additional planetary drive.

Function of the gears Fig. 1 to 20

When the engine is started and the direction of travel "forward" is preselected, the second hydrostatic unit B is set to a negative adjustment value, preferably the maximum negative adjustment value. The summation planetary gear is designed so that the first output shaft Al has the "zero" rotational tooth, so that the first range clutch Kl can be coupled to the stationary output shaft A. The first range is now driven through by hydrostatic adjustment, preferably up to its end setting point, in the positive direction of rotation until synchronous operation of the second output shaft A2 with the clutch K2 is achieved. After the clutch K2 is closed and the clutch Kl is opened, the second switching range is now driven through by hydrostatic re-regulation in the other adjustment direction up to the end of its adjustment. A reversing device, preferably via a planetary gear, is used for reverse operation, in various designs as shown in Fig. 15 to 17 and already described elsewhere. In the reverse range, the first output shaft Al is put into drive connection with the output shaft 2c; 2c' by closing a clutch or brake KR, whereby the speed reversal takes place by connecting a web shaft Pt2 to the housing.

The gearbox designs according to Fig. 21, 22 and 23 are gearbox designs with more than two shift ranges. These gearbox designs have a summation planetary gearbox P2 with two input shafts El and E2 and three output shafts Al, A2 and A3. These gearbox designs are described in more detail in the known publications DE39 29 209 or EP 0 386 214 or DE 40 27 724. As with the gearbox designs mentioned above from Fig. 1, the same functional sequence applies for the first and second shift ranges, although a reduction gear P3 is interposed in the first and second shift ranges, which serves to transmit power when engaged with the corresponding clutch or brake KV and another clutch K5. In the third shift range, in this gearbox design, the third output shaft A3 of the summation planetary gearbox is connected to the output by engaging a clutch K3. At the end of the second shift range, the third output shaft A3 and all elements of the clutch K3 have reached synchronism, so that this clutch K3 can be closed and the clutch KV can be opened. The clutch K2 remains closed. By adjusting the hydrostatic drive back to the other adjustment limit, the third shift range can now be passed through, whereby at the end of this range all elements of the clutch K4 have reached synchronism, so that they can be closed and then the clutch K3 can be opened. By repeatedly adjusting the hydrostatic drive back to the other adjustment limit, the fourth driving range can now be passed through.

A further feature of the invention is particularly characterized by the fact that the continuously variable transmission HVG; MVG, which preferably forms a common structural unit, can be introduced into the transmission housing on one side, through a housing opening la' on the right, left, top or bottom, and is fastened to the transmission housing 1 by a fastening device F5 and F6, as shown in Fig.8 and Fig.9, the fastening device F5 and F6 being designed in such a way that distortions of the housing 1, as can occur, for example, in tractors with a self-supporting housing, in particular in the case of one-sided impact loads or due to uneven road surfaces, cannot have a damaging deformation effect and/or vibrations on the transmission HVG; MVG. For this purpose, the continuously variable transmission is fastened at one point, e.g. on the input or output side, via two or more fastenings F5 point II, F5 point III and on the other side only at one fastening point F6 point I. The first-mentioned fastening point F5 has at least two fastening points F5II, F5III, at which the reaction torque of the gearbox is absorbed; no torque or reaction torque of the gearbox is to be transmitted to the opposite fastening point F6, since this fastening point F6 is primarily used to fix the position of the gearbox HVG; MVG.

In this way, the above-mentioned load-dependent or impact-dependent housing deformations of the main housing 1 are kept away from components of the continuously variable transmission HVG; MVG and the noise and vibration behavior is also reduced to the outside. The fastening type corresponds to a three-point bearing based on the respective corner points of the fastening F6 point I and F5 point II and point III (see Fig. 9, it shows the top view of the flange surface of Fig. 8). In order to ensure secure support of the reaction torque, any number of fastening points can be used on the torque-absorbing side F5, e.g. in the form of a screw connection made up of several screws. The fastening side F5 can, as mentioned, be on the gearbox input or output side or in between, but care must be taken to ensure that there is a sufficiently large distance between the fastening points F6 and F5. The output side F5, which is subject to a higher torque, should preferably be used for this purpose. Centering pins can be used to fix the position precisely. The gearbox attachment F5, F6 can preferably be provided on the gearbox housing 1, as mentioned, or also on the gearbox cover la. However, it is preferable to attach the entire gearbox to the gearbox housing. This has the advantage that the mounting frame of the gearbox HVG; MVG can be provided with openings for mounting the shaft connections Fl and F2 of the input and output shafts of the gearbox. A further advantage with regard to noise behavior is provided by this solution in that structure-borne vibrations of the gearbox can be kept away from the housing cover. In order to be able to compensate for manufacturing inaccuracies or with the aim of being able to allow greater manufacturing inaccuracies, it is expedient to provide the shaft connections Fl and F2 at the gearbox input and output with a correspondingly large amount of play or to provide cardan connections in a known manner, e.g. in the form of toothed sleeves. The invention described has the further advantage that the gearbox HVG; MVG is possible and thus workshop costs and the time spent in the workshop can be reduced to a minimum and, in addition, the HVG; MVG gearbox can be completely manufactured and tested separately and can be used as a replacement gearbox.

A further embodiment of the invention provides, as shown in Fig. 12, that the continuously variable transmission or continuously variable power split transmission HVG; MVG as a compact transmission in complete or partially complete design is mounted on the input and output side via a centering bearing F7 and F8, whereby the reaction torque of the transmission HVG; MVG on the transmission housing 1 or on one of the fixed housing elements, e.g. the

Gearbox cover la is supported. The central gearbox bearing F7 or F8 can also be designed with intermediately mounted noise and vibration reducing bearing elements F9' to optimize the noise behavior. The torque support can be designed in different ways, e.g. similar to that shown in Fig. 11, in which the intermediately mounted elastic elements 4f or 4g are intermediately mounted between a support element F7' or F8' connected to the housing and the housing element G of the continuously variable transmission to absorb the reaction torque, whereby any number of elastic elements 4f or 4g, e.g. in spherical 4f or conical shape 4g, can be provided. The torque support can also be designed for the axial fixation of the gearbox. With regard to ease of assembly, a support element F7' or F8' is provided on the input and output sides, preferably in flange form with centering, which enables the MVG; HVG gearbox to be connected to the main housing 1 in a simple and easy-to-assemble manner, for example via a screw connection F4'. The gearbox design described above can be used for inline construction, as shown in Figs. 7 to 14 and 19, as well as for designs with an axially offset hydrostatic gearbox, as shown in Figs. 4; 6; 23, although the two gearbox bearing points F7 and F8 must be coaxial with the input shaft Ic (not shown in the drawings). Hydrostatic gearboxes with summation and shift gearbox 5c also form a complete structural unit in the latter design. In this embodiment, as shown in Fig. 12, the continuously variable gearbox MVG; HVG can be inserted and installed laterally through a housing opening la' as in the previously described embodiments in order to take advantage of the advantages of ease of installation and service, as already described. However, it can also be installed in a gearbox housing in the usual way, e.g. without a side housing cover.

The invention is further characterized in that, as shown in Figures 10; 11; 13; 14 and 26 and partially explained above, the complete or almost complete continuously variable power split transmission (hydrostatic split transmission HVG) forms a transmission unit which can be designed in different ways, i.e. in different embodiments of the power split, in particular, as already mentioned, with a different number of shift ranges, e.g. as a single-range transmission (not shown) or as a 2-range transmission Fig.7 or 4-range transmission, e.g. Fig.21; 26, and can be adapted to different vehicle requirements. The power split transmission HVG is designed as a compact transmission with or without a control unit Ie, which can be very easily introduced into the housing opening la' of a main housing 1, e.g. in a tractor, and only on one side, in particular the output side, via a connecting device.

•&igr;'&iacgr;&iacgr;&iacgr;' **i "i

in particular a screw connection F4, can be connected to the main housing 1. The housing opening la' can be relatively small here, since the special gear shape of the HVG, in particular the inline design, which in the embodiment shown can be implemented in an almost cylindrical shape and as shown in Fig. 11, requires little installation or assembly space la". The control unit Ie or Ie' is expediently attached to the gear housing 1, e.g. with plug connections, as already described, for the control pressure, lubricating oil connection, etc. or, as shown in Fig. 26, to the entire HVG gear.

The drive-side connection to the vehicle engine can be made in the usual known manner, e.g. by means of a sleeve or flange connection (not shown) or also a direct connection to the flywheel of the drive engine by means of an intermediate vibration damper, as is known per se.

By eliminating spur gear stages, which are required as a drive connection for example in the case of a parallel offset hydrostatic transmission, there are particular advantages in terms of installation space, costs, efficiency, ease of assembly and service and also noise behavior. Another advantage is that a very quick transmission exchange (replacement transmission) is possible and workshop costs and workshop time can be reduced to a minimum. This fulfills the essential tasks of this invention.

In the case of gearboxes as shown in Figures 7 to 11, the drive motor is preferably arranged on the same axis as the gearbox drive shaft Ic, with the axle differential DIF being offset from the drive shaft. This has the advantage that a simple power take-off connection is possible without intermediate spur gear stages, as the drive shaft Ic is guided centrally through the gearbox and the power take-off connection can be implemented via a shaft connection F3, e.g. in the form of a plug connection with a drive profile. In the design according to Figures 13, 14 and 26, the gearbox output is at the same height as the axle differential DIF. In this design, it is not necessary to guide the drive shaft Ic through the gearbox to the output. Depending on the vehicle requirements, a corresponding selection can be made from the various gearbox designs. In the gearbox design according to Fig. 14, the drive shaft Ic' is shown offset to the drive shaft of the gearbox HVG, whereby the drive connection to the drive shaft Ic and the gearbox HVG is established via a spur gear stage 26c'. In this gearbox design, the power take-off shaft connection is made by a shaft 2e arranged offset parallel above or next to the gearbox HVG. In this embodiment according to Fig. 14, it is expedient to place the housing opening on one of the gearbox sides (not shown).

The HVG transmission can also be designed as a separate continuously variable transmission with a completely enclosed housing, as shown in Fig. 26, which can be attached to a vehicle axle via a corresponding connection (screw connection F4), for example. As shown in Fig. 11, according to the invention, the hydrostatic transmission 4c is designed as a separate or separately encapsulated unit with the hydrostatic units A and B, as is the summation planetary transmission in various embodiments, as already or subsequently described, and the clutches or range clutches and/or brakes for switching the corresponding switching ranges are advantageously combined according to the invention to form a unit in the form of a transmission roller 5c. The hydrostatic transmission 4c and the transmission roller 5c are connected to one another via a housing or carrier element G. According to the invention, the hydrostatic transmission 4c is, as already described in a similar form, mounted relative to the housing and support element G in a way that reduces noise and vibrations by means of a corresponding, preferably elastic damping device 4e, 4h. According to the invention, elastic elements 4f and 4g, in particular made of rubber, are mounted in a very cost-effective manner in corresponding recesses 4e' that are incorporated into the hydrostatic housing 4c and the support element G. The elastic damping elements can be designed in a spherical shape 4f or a conical shape 4g or in other shapes not shown. The corresponding recesses 4e' on the hydrostatic transmission 4c and the carrier element G are preferably tapered bores, so that when a reaction torque occurs on the hydrostatic transmission, in addition to the torque support, an axial force corresponding to the torque is generated, which is supported on a corresponding damping element 4h, which is preferably designed as a friction ring. This has the advantage of very effective friction damping. The aforementioned tapered bores 4e' are very cost-effective, and can each be produced in a single operation. The oil and control pressure connections from the control unit Ii to the hydrostatic branching transmission HVG are expediently made via plug connections, which can also be implemented cost-effectively using known connections 3a with O-ring seals.

The transmission design according to the invention has the further advantage that the axle distance a' is very small and can therefore be adapted to the most diverse vehicle requirements.

As already mentioned, various designs of the power split transmission are possible. Fig. 19 shows a power split transmission in which the clutches Kl, K2 are arranged between the hydrostatic transmission 4c and the summation planetary transmission PS3. The summation planetary transmission PS3, as shown in Fig. 19, has four shafts and has three input shafts and one output shaft. The first input shaft El is

with the first hydrostatic unit A and the first shaft a2' of the

Summation planetary gear is permanently connected. The second input shaft El' and the second input shaft E2' of the summation planetary gear can be alternately connected to the second hydrostatic unit B via clutch Kl or K2. The output shaft Al' of the summation planetary gear is permanently connected to the output shaft A' of the gearbox or, as shown, can be connected via a clutch KV. The summation planetary gear PS3 has a carrier shaft connected to the output shaft Al', on which intermeshing planetary gears Pl are arranged. A ring gear Hl', which can be connected to the first range clutch Kl, and a sun gear Sl', which can be connected to the second clutch K2, engage with first planetary gears. Second planetary gears mesh with a ring gear H2' connected to the drive shaft El. The summation planetary gear PS3 can be designed in different ways, e.g. with two planetary sets as shown in EP 0 242 372 Fig. 1. The gear design according to Fig. 19 has two forward and two reverse ranges, whereby a planetary gear PR4 is preferably provided for reverse operation, which causes a reversal of the direction of travel or rotation when the clutch or brake KR is engaged. The output shaft Al' can be equipped with a sun gear and the output shaft A' with a ring gear, whereby the web of the planetary gears can be connected to the gear housing via a clutch or brake KR for reverse travel. For forward travel, the output shaft Al' is connected to the gear output shaft A' via a clutch KV, as already explained. Depending on the vehicle requirements, the maximum reversing speed can be adjusted as required by selecting the planetary gear from the planetary gear series PR4, PR5, PR6 and others, as shown in the drawings in Fig. 15, 16 and 17 and described earlier.

Another gear design according to Fig. 20 is functionally identical to the design according to Fig. 19. The only difference is a different structure of the summation planetary gear 187. The summation planetary gear 187 consists of two planetary gear stages Pl and P2, whereby the input shaft E2' is connected to the ring gear 182 of the first planetary gear stage and both sun gears 185 and 186 of the first and second planetary gear stages are connected to the input shaft E2'. The web 188 of the second planetary gear stage P2 is coupled to the first input shaft El and the web shaft 183 of the first planetary gear stage Pl and the ring gear 184 are coupled to the output shaft Al'.

With the aim of being able to adapt the maximum reversing speed or starting traction to different vehicle requirements, the invention provides for various

Embodiments of the reversing device as shown in Fig. 15, 16, 17, 18. The aforementioned reversing devices are preferably applicable to the transmission designs according to Fig. 7 to 14, the essential difference being that an output member of the transmission can be connected to various planetary gear designs - PR4, PR5, PR6 -, whereby in PR4 the aforementioned transmission output member is connected to a sun gear S4 and the transmission output shaft A'; A" is connected to a ring gear H4 and the carrier shaft St4 can be connected to the housing via a clutch or brake KR. In the design according to Fig. 16, the reversing device has a planetary gear PR5, with intermeshing planetary gears Pll and P12 being arranged on a carrier shaft, with a ring gear H3' connected to a gear output member engaging first planetary gears Pll and the gear output shaft A'; A" being connected to a further ring gear H3" which engages second planetary gears P12. The reversing device according to Fig. 17 has a planetary gear PR6 with intermeshing planetary gears Pll and P12 which are mounted on a carrier shaft St6. An output member of the gear engages a sun gear S6 which meshes with first planetary gears Pll. The output shaft A'; A" is connected to a ring gear H3' which meshes with second planetary gears P12 meshing sun gear S6'. The carrier shafts St4; St5; St6 can be connected to the housing via a brake or clutch KR for the reversing range. For lower reversing speeds, the design as shown in Fig. 15 is suitable, and for high reversing speeds, the design as already shown in the complete gearboxes Fig. 10 to 14. If approximately the same starting tractive forces as when driving forward and the same speed ratios are required, the reversing devices with the planetary gearboxes PR5 or PR6 as shown in Fig. 16 or 17 are suitable.

For the requirement for approximately equal forward-reverse speeds or the same number of reversing ranges, a complete reversing gear according to Fig. 18 is connected to the drive shaft A'; A". As shown in Fig. 17, a planetary gear PR6 is used with a carrier shaft that can be coupled to the housing 1; G, on which intermeshing planet gears Pll and P12 are arranged, with a sun gear S6 attached to the output shaft A' engaging in the first planet gears Pll and a sun gear S6' connected to the output shaft a2 engaging in the second planet gears P12. All embodiments, as shown in Fig. 15 to 17, can be used for this, provided that the corresponding first planetary member S4; H3'; S6 is connected to the output shaft A'; A".

The function of the transmission according to Fig. 19 corresponds in terms of the switching method or the switching sequence and the hydrostatic speed behavior of the transmission design according to Fig.7, whereby in

23

In the starting state at a driving speed of "zero" and the clutch Kl closed, the hydrostatic transmission is preferably fully regulated from one gear ratio end point to the other, with all elements of the summation planetary transmission PS3 and the elements of the second range clutch K2 running synchronously at the end of the first shift range. In this state, the clutch K2 can be closed and then the clutch Kl opened and then the hydrostatic transmission can be adjusted in the opposite direction up to its adjustment end point, at which the final gear ratio of the transmission or the maximum speed of the transmission can be reached. The transmission thus has two hydrostatic-mechanical shift ranges similar to the aforementioned design according to Fig. 7. All clutch shifts take place with synchronous operation of the clutch elements, with the shaft El' of the summation planetary transmission in the starting state at a driving speed of "zero" having the same speed in the same direction of rotation as the second hydrostatic unit B in order to ensure that the clutch Kl closes smoothly.

In Fig. 27, a transmission design according to the invention with a mechanical converter 4d, in particular a belt transmission and a downstream transmission 5d, which is preferably designed as a power split transmission, is shown. Similar to the aforementioned transmission designs, this transmission according to the invention is also designed as a compact transmission or mechanical split transmission MVG, which can be installed through a transmission opening la' in a transmission housing, e.g. of a tractor, or can be attached to a transmission, e.g. of a drive axle IA, via an identical or similar connecting device F4. In order to achieve the most compact design possible, the mechanical converter transmission consisting of a primary unit (conical disk 4dA) and a secondary unit (conical disk 4dB) is constructed in such a way that the first primary unit 4dA is connected to the drive shaft Ic via a gear stage dA and the secondary unit 4dB is connected to a gear element of the summation planetary gear or summation gear 5d via a further gear stage dB. The

Summation gear 5d can be designed as a planetary gear with one or more shift ranges, similar to the hydrostatic-mechanical gear designs described above. Here too, as shown, the output shaft of the gear can run coaxially to the input shaft Ic through the gear and be connected coaxially to a gear, e.g. an axle gear IA. In a further embodiment not shown, the differential gear DIF or the input shaft 2"' of the drive axle IA or the corresponding subsequent gear is arranged offset from the input shaft Ic and is in drive connection with the gear output shaft A' via a corresponding spur gear stage 2d.

Here, as with the gearbox design according to Fig. 10, it is possible to guide the gearbox input shaft Ic centrally through the gearbox in order to realize, for example, a coaxial through-drive for a shaft 2e' or power take-off shaft or PTO.

For all transmission designs with an inline arrangement of the hydrostatic transmission and the shift or transmission roller, as shown in Figs. 7 to 19, it is advisable to design the second hydrostatic unit B, which is usually designed as a constant unit, for a smaller delivery volume than the adjustment unit A in order to be able to compensate for the speed slip caused by leakage oil for the range shifts. This can be easily compensated for by a correspondingly smaller angle of the swash plate or, in the case of an adjustable hydrostatic unit B, by appropriate secondary adjustment. For transmission designs with a parallel offset hydrostatic transmission, the speed slip can be compensated for by appropriately adjusting the ratio in the connecting spur gear stages.

The gearbox design according to Fig. 22 is identical to the gearbox design 21, but with the difference that the third output shaft A3 overlaps the clutches Kl and K2 and can be coupled to the carrier shaft PST3 via the clutch K3 and therefore the output shaft A' can be arranged directly above the continuous drive shaft Ic.

The gearbox design according to Fig. 24 has a five-shaft summation planetary gearbox 157 with two input shafts El and E2 and three output shafts Al; A2 and A3. A countershaft gearbox 156 is arranged downstream of the summation planetary gearbox. This gearbox design is largely identical and functionally identical to the known gearbox according to EP 0386214 as per Fig. 9 and DE 40 27 724A Fig. 9 and 91, but with the difference that instead of the downstream planetary gearbox, a countershaft gearbox 156 is used. In the first and second shift range, with the clutch Kl or K2 alternately engaged, the summed power is transmitted to the output shafts Al or A2 via the spur gear stage 158 with the clutch 162 closed. In the third shift range, the power is transmitted via the third output shaft A3 to another spur gear stage 161 with the clutch 165 closed. In a fourth shift range, the power is also transmitted via the same spur gear stage 161, with the power flowing via the second output shaft A2 with the clutch K2 and clutch 164 closed. A fifth shift range can also be switched here, with another spur gear stage 160 being transmitted via the third output shaft A3 of the summation planetary gear with the clutch 166 closed.

The gear concept according to Fig.25 with the same five-shaft summation planetary gear 157 provides that the summation planetary gear has three planetary gear stages PVl; PV2 and

PR are assigned. The sun gears 169 and 167 of the two planetary gear stages PR and PVl are connected to the first output shaft Al and the sun gear 168 of the planetary gear stage PV2 is connected to the second output shaft A2 of the summation planetary gear. The carrier shaft 181 of the planetary gear stage PR can be connected to the housing via the clutch KR for the reverse group. The ring gears 171 and 170 can be connected to the housing in the first and second shift ranges via clutches Kl and K2 respectively. The planet gears of the planetary stages PVl and PV2 have a common carrier shaft 172, which is connected to the ring gear of the planetary stage PR and the output shaft 2c. In the first shift range, the accumulated power flows via the first output shaft Al with the clutch or brake Kl closed via the planetary gear stage PVl, in the second shift range via the second output shaft A2 with the clutch or brake K2 closed, in the third shift range via the third output shaft A3 with the clutch K3 closed and in the fourth shift range via the second output shaft A2 with the clutch K4 closed to the output 2c. A shift range is provided for the reverse range, which holds the carrier shaft of the planetary gear stage PR to the housing when the clutch or brake KR is engaged, with the power flowing via the first output shaft Al via this planetary gear stage PR to the output. This gearbox design according to Fig. 27 has four forward and one reverse ranges. The functional sequence is such that when the direction of travel is preselected, the second input shaft E2 connected to the second hydrostatic unit B rotates in the opposite direction to the first input shaft El, with the first output shaft Al standing still, so that the clutch Kl can be engaged when the ring gear of the planetary gear stage PVl is stationary. The hydrostatic unit is now regulated within the first switching range up to its other adjustment end point, after which the second output shaft A2 and both sun gears 168 and 167 of the planetary gear stages PV2 and PVl have reached synchronization, whereby the ring gear 170 of the planetary stage PV2 is stationary in order to be able to close the second range clutch K2 and open the first range clutch Kl. The second range is now passed through by regulating the hydrostatic unit back to the starting point, with the third output shaft A3 of the summation planetary gear having reached synchronization with the output shaft 2c at the end of the second switching range. After closing the clutch K3 and opening the clutch K2, the third area is now connected by correspondingly repeated back-regulation of the hydrostatic drive to its other end point, after which the second output shaft A2 and thus the clutch elements of the clutch K4 are running synchronously. After closing the clutch K4 and opening the clutch K3, the fourth area can now be connected by repeated back-regulation of the hydrostatic drive.

«26

The summation planetary gear 157 has intermeshing first planet gears 174

and second planetary gears 173, which are arranged on the carrier shaft 179 connected to the third output shaft A3. A ring gear 177 connected to the first input shaft El and a sun gear 175 connected to the second input shaft E2 engage with first planetary gears 174. Second planetary gears 173 mesh with a ring gear 178 connected to the first output shaft Al and a sun gear 176 connected to the second output shaft A2.

To improve efficiency, the invention provides that the second hydrostatic unit B is held in place in a hydrostatically powerless operating state, which for example corresponds to the main operating range of a tractor at around 8 km/h, e.g. by closing a clutch or brake BR. According to the invention, a less expensive brake band BR is preferably used here, which preferably closes automatically at the corresponding transmission point. The control and regulation device is programmed for this purpose in such a way that, for example, after a certain dwell time at this preferred ratio point or in its vicinity, it closes automatically, whereby, depending on the given conditions, an automatic ratio correction or adjustment of the hydrostatic adjustment to the most precise setting possible to delivery volume "zero" takes place and/or a bypass function of a known type can become effective so that the hydrostatic pressure or differential pressure can be reduced to "zero" in order to largely eliminate the hydrostatic losses. This fixed ratio point corresponding to the fixed hydrostatic unit B can be maintained within a pre-programmed speed range, whereby the two end points - upper or lower end point of the specified engine speed in the corresponding gear range of the transmission - determine the switching on and off of the holding device. This engine speed range can be a fixed or variable operating variable, in particular dependent on the efficiency values of the transmission and the engine, which are preferably determined experimentally. The hydrostat 4c can also be used in this function, with the appropriate control device, as a brake retarder, as described in more detail in DE 43 39864.

According to the invention, the continuously variable transmission HVG or MVG also contains the complete control and regulating device 5; Ie (see Fig. 10; 26). This means that in addition to the control block and the entire adjustment device, the electronic components are also included, so that in the event of damage, the entire transmission can be replaced very quickly and the assembly effort, downtime and assembly errors can be reduced to a minimum.

Claims (15)

DE 122 N 35 DE 122 H 35 PCT 122 E 35 f ä I e irtan claims i Ji &eegr; s &rgr; r ü 1. Continuously variable hydrostatic or mechanical transmission with power split, in particular for tractors, with a first hydrostatic unit (A) with adjustable volume or primary unit (4dA) and a second hydrostatic unit (B) with constant or adjustable volume or secondary unit (4dB) with a summation planetary gear for summing the hydraulic and mechanical power or the continuously variable and non-continuous power branch with or without clutches (Kl, K2) thereby
characterized in that the hydrostatic units (A and B), the summation planetary gear (PSl; PSl';PSl";PS2;PS2';PS3;301;201; 101 et al.) and optionally clutches/brakes (Kl, K2, KR or Kl, K2, K3, K4, K5, KV, KR) are arranged coaxially to one another or the mechanical converter (4d) has a drive shaft (Ic) coaxial to the summation planetary gear (Fig.26 to 28; 38; 39; 43; 44), whereby - the output shaft (A"; 106) of the transmission is connected or can be connected to a drive shaft (2c"') of a differential gear coaxial to the branching gear (HVG / MVG) (Fig. 13; 14; 26 to 28) or - that the output shaft (A') is connected or can be connected via a gear stage (2d) to an axially offset drive shaft (2c"), whereby this gear design is equipped with or without a centrally leading through the gear unit connected to the input shaft (Ic) for the connection of a power take-off shaft or other serving shaft (2e) (Fig.7 to 11; 21; 29; 30 to 33 and others). 2. Continuously variable transmission with a mechanical continuously variable converter (4d) with power split, in which the input power is divided into a continuously variable power branch and a power branch with a constant ratio, which are added up in a summation transmission (5d), characterized in that the summation transmission (5d or 301; 201; 101) is arranged on the same axis as the drive shaft (Ic) and the continuously variable mechanical converter (4d) and the summation transmission (5d; 301; 201; 101 Fig.35 and others) form a common structural unit as a mechanical branching transmission (MVG), which is integrated into a main housing (1), in particular of a tractor, by a housing Opening (Ia') can be installed and removed, whereby the drive shaft or the output shaft of the gearbox is connected via connecting elements (Fl and F3) and the gearbox itself is fastened, e.g. on the gearbox output side via a connecting device (screw connection F4) (Fig.29) or can be installed in a vehicle frame (frame of a tractor not shown) or via a frame fastening (F6, F5 Fig.8, 9) on the gearbox housing (1). 3. Transmission according to claim 1, characterized in that the hydrostatic units (A and B) and the transmission unit (5c), which consists of summation planetary gears and if necessary clutches for switching one or more switching ranges, wherein a housing or carrier element (G) is provided which carries or connects the hydrostatic transmission (4c) with hydrostatic units A and B and transmission (5c) with summation planetary gears and if necessary clutches, such that a common assembly (HVG) is formed, wherein the hydrostatic units (A and B) are combined to form a common assembly (4c) which is mounted with or without a device (4e, 4h) for reducing noise and vibrations relative to the housing or carrier element (G), or that the components - hydrostatic components (A and B), the summation planetary gear or coupling gear (5c) with or without clutches (Kl, K2, KR) have a common housing (G), whereby the structural unit (HVG) is designed as a compact gearbox and can be installed or attached in or to a main housing (1), preferably of a tractor, via a fastening device (screw connection F4) (Fig. 10; 11; 13; 14; 8; 19; 26). 4. Transmission according to one of the above claims, characterized in that the power split transmission (HVG; MVG) can be combined with a high-drive transmission (HT Fig. 31) to adapt the speed of the transmission input shaft (Ic) to the engine speed, the transmission (HT) being designed as a planetary transmission and the sun gear (S) being connected to the housing (1), the carrier shaft (St) to the drive shaft (lca) and the ring gear (H) to the transmission input shaft (Ic) and/or that the power split transmission (HVG; MVG) is connected to a downstream group transmission (GR) which enables two or more group shifts (L/H or road group S, field group A). (Fig.32; 33) 5. Transmission with or without power splitting with a continuously variable converter (4d) which consists of a primary unit (4dA) and a secondary unit (4dB), with or without a summation transmission (5d), characterized in that the continuously variable converter (4d) is designed as a mechanical converter, in particular as a belt transmission, wherein the both converter elements (4dA and 4dB) are axially offset from the drive shaft (Ic) and the primary converter element (4dA) is connected to the drive shaft (Ic) via a spur gear stage (dA) and the secondary unit (4dB) is connected to the summation gear (5d) via a spur gear stage (dB) and that the converter gear (4d), the summation gear (5d) are arranged coaxially to each other and form a common structural unit or mechanical branching gear (MVG) which can be installed in a main housing (1; Fig.29) or on a gearbox (IA Fig.27) via a connecting device (F4) and wherein the gearbox output shaft (2; 2") is connected to an input shaft (2"') coaxial to the drive shaft (Fig.27) or that the gearbox output shaft (2") is connected via a spur gear stage (2d') and an axially offset drive shaft, preferably a pinion shaft (2"'), whereby the gearbox input shaft (Ic) preferably runs centrally through the gearbox and offers a connection option for PTO or auxiliary drives on the gearbox output side. 6. Continuously variable mechanical power split transmission with a continuously variable converter (4d) with a primary unit (4dA) and a secondary unit (4dB), a summation planetary gear (301; 201; 101; 401) for summing up the power divided at the transmission input into a transmission branch with a constant gear ratio and a continuously variable transmission branch and two or more shift ranges, characterized in that the summation planetary gear (301; 201; 101; 401) is designed with four shafts and has two input shafts (El and E2) and two output shafts (Al and A2), that the first input shaft (El) is in drive connection with the drive shaft (Ic) and the primary unit (4dA), the second input shaft (E2) is drive-connected with the secondary unit (4dB) of the converter (4d) and the first output shaft (Al) in the first shift range and the second output shaft (A2) in the second shift range transfers the power accumulated in the summation planetary gear directly or via intermediate elements alternately to an output shaft (106) and that the first output shaft (Al) has the speed "zero" when the drive shaft (Ic) is rotating (Fig.35 to 44). 7. Transmission according to claim 6, characterized in that the secondary unit (4dB) has an increasing speed from driving speed "zero" to the end of the first shift range and a decreasing speed after the start of the second shift range, that when changing ranges from range 1 to range 2 and vice versa, all members of the summation planetary gear (301; 201; 101; 401) and the clutch members to be switched (K2 or Kl) have synchronous operation. 8. Transmission according to claim 6 and 7, characterized in that the summation planetary gear (301; 201; 101; 401) is assigned a transmission gear (112; 112a; 113) and that in the first and second shift ranges the power is conducted via a transmission stage (146) or a planetary stage (115) of this assigned transmission (112; 112a; 113) and in a possible third shift range the second input shaft (E2) of the summation planetary gear transmits the power to the output shaft (106) without power splitting and that in a possible fourth shift range the second output shaft (A2) transmits the power summed up in the summation planetary gear to the output shaft (106) when the clutch (K4) is switched accordingly. 9. Transmission according to claims 6 to 7, characterized in that a planetary gear stage (PR, Figure 38; PR4; PR5; PR6; PR6', Fig. 15 to 17) is assigned for reverse travel, the first output shaft (Al) of the summation planetary gear being connected to a member (153; S4; H3'; S6) of the planetary stage (PR) and a further member (sun gear 154; ring gear H4; sun gear S6') being connected to an output shaft (106) and a third member (web 155; St4; St5; St6) being connectable to the housing (1; G) via a clutch (KR) or that an output shaft (106; 109) is in drive connection with a spur gear stage (147) formed with an intermediate gear (147a) or that an output shaft or intermediate shaft (109) is connected to a planetary gear stage (PR), whose carrier shaft (118a) can be connected to the housing (1; G) via a coupling (KR) (Fig. 44; 42; 41). 10. Transmission according to one of claims 6 to 9, characterized in that the summation planetary gear (301, Fig.35) is designed with two planetary gear stages (Pl and P2), the first input shaft (El) being connected to the carrier shaft of the first planetary gear stage (Pl) and the ring gear (127) of the second planetary gear stage and the second input shaft (E2) being connected to the ring gear (125), the first output shaft (Al) being connected to the carrier shaft (128) of the second planetary gear stage (P2) and the second output shaft (A2) being connected to the sun gear of the first and the sun gear of the second planetary gear stage and that the second planetary gear stage (P2) has two intermeshing planet gears (123 and 124) or that the summation planetary gear (401 Fig.39) has a carrier shaft (133) on which intermeshing first planetary gears (131), second planetary gears (130) and third planetary gears (132) are arranged, wherein a ring gear (134) connected to the second input shaft (E2) engages in the first planetary gears and a ring gear connected to the third planetary gears (135) engages in the third planetary gears (135). the first output shaft (Al) meshes with a ring gear connected to it, that a sun gear (136) connected to the second output shaft (A2) meshes with third ring gears (132), that the web shaft (133) is drive-connected to the drive shaft (Ic) and the primary unit (4dA) of the converter or variator (4d) or that the summation planetary gear (101, Fig.40) has a carrier shaft connected to the second input shaft (E2) on which intermeshing first planetary gears (107) and second planetary gears (108) are arranged, wherein a ring gear (102) connected to the first input shaft (101) engages in first planetary gears (107) and a ring gear (104) connected to the first output shaft (Al) and a sun gear (105) connected to the second output shaft (A2) engage in second planetary gears (108) or that the summation planetary gear (201, Fig.42) has a carrier shaft (144) connected to the second input shaft (E2), on which intermeshing first planetary gears (139) and second planetary gears (140) are arranged and a ring gear (141) connected to the first input shaft (El) is in engagement with first planetary gears (139), a ring gear (142) connected to the first output shaft (El) is in engagement with second planetary gears (140) and a sun gear (143) connected to the second output shaft (A2) is in engagement (Fig.42). 11. Continuously variable hydrostatic transmission with power split consisting of a first hydrostatic unit (A) with adjustable volume and a second hydrostatic unit (B) with constant or adjustable volume with a summing planetary gear for summing up the hydraulic and mechanical power, characterized in that the summing planetary gear (PSl; PSl'; PSl") is designed with four shafts and has two input shafts (El' and E2') and two output shafts (Al' and A2'), the first input shaft (El') being connected to the drive shaft (Ic) and the first hydrostatic unit (A) and the second input shaft (E2') being connected to the second hydrostatic unit (B) and the first output shaft (Al') being connected via a clutch (Kl) and the second output shaft (A2') being connected via a second clutch (K2) to an output shaft (A') alternately. that a further three-shaft planetary gear (PRl'; PR4; PR5; PR6 Fig. 15 to 18) is assigned, which is connected to the first output shaft (Al') of the summation planetary gear (PSl; PSl'; PSl") with a planetary gear element (ring gear H3' or sun gear S4; S6) and the second shaft (sun gear SR' or ring gear H4; H3") with the output shaft (2c; 2c') and that the third shaft (web PT2'; St5; St6) can be coupled to the housing via a clutch or brake (KR) (Fig.l; 2; 5; 7; 8; 10; 15; 16; 17 and others). • t ·· 12. Transmission according to claim 11, characterized in that the summation planetary gear (PSl; PSl') has intermeshing planetary gears (PU' and P12'), whereby first planetary gears (P12') engage in a ring gear (HEl) that is in drive connection with the drive shaft (lc; lc") and second planetary gears (PIl') engage in a ring gear (HE2) that is in drive connection with the second hydrostatic unit (B), and the first output shaft (Al') represents the web shaft, which is connected to the output shaft (Al') via a clutch (Kl), and the second output shaft (A2') of the summation planetary gear represents a sun gear (SA; S3'), which is connectable to the output shaft (A') and output shaft (2c) via a second clutch (K2). 13. Transmission according to claim 11 and 12, characterized in that the hydrostatic transmission (4c) and the drive shaft (Ic) are arranged offset from the summation planetary transmission (PSl) or that the summation planetary gear (PSl" according to Fig.5) is designed with four shafts and has two input shafts (El' and E2') and two output shafts (Al' and A2'), and consists of two planetary gear stages, whereby the first input shaft (El') is connected to the carrier shaft (StI") of the first planetary stage, the second input shaft (E2') is connected to the ring gear (H2") of the second planetary stage, the first output shaft (Al') is connected to the ring gear (Hl") of the first planetary gear stage and the carrier (St2") of the second planetary gear stage and the second output shaft (A2') is connected to the sun gear (Sl') of the first planetary gear stage and the sun gear (S2') of the second planetary gear stage (Fig.5; 7; 8; 10 and others) or that the hydrostatic transmission (4c'), the summation planetary gear (PSl') and the clutches (Kl and K2) are arranged coaxially to each other (Fig.l and 2), wherein preferably the drive shaft (Ic") and the PTO shaft or power take-off shaft represent a shaft leading through the gearbox, wherein the axle differential (DIF) is arranged parallel to the drive shaft (Ic") and is connected to the drive via an intermediate stage (2d) (Fig.2) or that the gearbox output shaft (2c') is designed as a hollow shaft and carries the bevel pinion for the bevel drive (KE) (Fig.3). 14. Transmission according to one or more of the preceding claims, characterized
characterized in that the hydrostatic transmission (4c) is placed parallel to the drive shaft (Ic"), the shift drum (5c), which contains the summation planetary gear and the shift clutches, being arranged on the drive shaft (Ic") and the first hydrostatic unit (A) being connected to the transmission output side via a spur gear stage (10b") and the second hydrostatic unit (B) being connected to the input side via a further spur gear stage (228'), and that the Output shaft (2c") and the axle differential (DIF) is arranged offset from the drive shaft (Ic) (Fig.4), or that the hydrostatic transmission (4c') is arranged parallel to the drive shaft (Ic), with both hydrostatic units (A and B) having the drive connection on the drive side via spur gear stages, with the first hydrostatic unit (A) being in drive connection with the drive shaft (Ic) via a spur gear stage (10b') and the second hydrostatic unit (B) being in drive connection with the shift drum (5c) via a spur gear stage (228'), and with the shift drum (5c) being arranged on the drive shaft (Ic) and the output shaft (2c") and the differential (DIF) being arranged axially offset or not axially offset (as shown in Fig.3) (Fig.6). 15. Transmission according to one of the preceding claims, characterized
that a holding device (clutch/brake BR) is provided which locks or holds the second hydrostatic unit (B) in operating states with a stationary output shaft, wherein this locking device is preferably designed as a brake with a brake band (Fig. 7) and that this holding function is triggered manually or automatically and also canceled again.