Patent Application
20240360829
The invention relates to an internal gear machine.
Internal gear machines of the generic type are known.
DE 198 26 367 A1, for example, discloses an internal gear pump in the form of a gear ring pump without a filler piece for pumping low-viscosity liquids. The gear ring pump comprises a pump mount and a split bearing body made of low-wear special material embedded therein, which forms a cavity. An internally toothed ring gear and an externally toothed pinion are arranged in the cavity, the toothings of which are in meshing engagement with each other in certain regions. The axes of rotation of the ring gear and pinion are arranged parallel to and spaced apart from each other. The toothings of the ring gear and pinion can be helically toothed. The split bearing body receiving the ring gear and pinion consists of a disc and a ring gear mount, which form an axial and radial bearing. A kidney-shaped aperture in the ring gear mount or the disc is congruent with a corresponding aperture in the pump mount or the disc and together they form the inlet or outlet channel of the pump on one side when viewed axially. On the side facing away from the aperture to the cavity, blind kidneys are provided in the corresponding disc or the ring gear mount that are congruent with the inlet or outlet channel and prevent the meshing toothing from squeezing oil in a known manner.
DE 20 2009 017 371 U1 and DE 41 02 162 A1 also disclose helically toothed gear ring pumps.
Compared to straight-toothed internal gear pumps, which are also known to those skilled in the art, the known helically toothed internal gear pumps offer the advantage of greater mechanical smooth running, as the helically toothed teeth mesh with one another with a continuous transition.
The known helical internal gear pumps are suitable for use in the low-pressure range.
A disadvantage of the known helically toothed internal gear pumps is that they have no hydraulic gap compensation. Under hydraulic pressure, this leads to high leakage as well as axial thrust and a tilting moment transverse to the axis of rotation on the hydraulically pressurised helical toothing. This leads to edge pressure and, as a result, to a high drive torque and wear. As a result, the known helically toothed solutions exhibit poor volumetric and/or hydraulic-mechanical efficiency as well as rapid degradation under higher pressure loads.
The known helically toothed solutions are therefore not suitable for operation at typically required 250 bar, 280 bar or 350 bar and/or necessary temperature spreads of up to −40° C. to 120° C. or with corresponding viscosity spreads during operation.
In addition, a significant noise advantage is only achieved if the helix angle is at least so large that, for a given gearbox width, the helix angle leads to a relative face section twist from the front of the gearbox to the rear of the gearbox of approximately one tooth pitch. Conventional toothings, especially those of gear ring pumps without a filler piece, have a relatively low number of teeth, typically between 6 and 15 teeth on the pinion and between 7 and 16 teeth on the ring gear. As a result, the helix angle must be very large to achieve a full pitch, which means that the engagement distance required for sealing in the tooth engagement is not long enough or the degree of overlap is too small. As a result, helical toothing cannot be sealed, particularly with regard to high pressure, and therefore cannot fulfil the purpose of being suitable for high pressure and quiet at the same time.
In other words, a relatively high helix angle could be implemented, thus achieving a purely mechanical noise advantage. However, this results in the pump being leaky in the tooth engagement, i.e. high leakage, sudden pressure reduction and cavitation occur, which ultimately means that such a pump is much louder and less efficient than a pump without helical toothing. Alternatively, a helix angle could be implemented which results in a hydraulic seal just being achieved through the engagement distance. In fact, the helix angle for conventional toothings is so small that the noise advantage over spur toothing is marginal and the associated manufacturing effort is not justified.
In contrast to the known helically toothed internal gear pumps, straight-toothed hydraulically gap-compensated internal gear pumps are known.
For example, DE 43 22 240 C2 describes an internal gear pump that has sealing discs arranged axially on the gearbox end faces, whereby the discs are designed with mirror symmetry to the gearbox centre plane. Axial gap compensation is achieved by corresponding axial pressure fields in the housing or housing parts that are symmetrical to the centre plane of the gearbox. The pressure fields are designed in terms of surface and surface centre of gravity in such a way that the pressure acting within the straight-toothed gearbox and its axial force effect are compensated for, thereby forcing the discs to contact at every operating point. This effectively prevents leakage between the gearbox end face and the sealing discs. The pressure fields are also designed with an appropriate sealing system to seal the sealing discs to the housing. Radial gap compensation between segment-shaped filler pieces and the gearbox tooth heads is achieved by the fact that the gap between two filler pieces in contact with the tooth heads has sealing elements and the gap is connected to the high-pressure side, thus forcing radial contact with the toothing.
The known compensated straight-toothed internal gear pumps are suitable for high-pressure applications. They are characterised by a high volumetric and hydraulic-mechanical efficiency. A disadvantage is that the known designs are noisy when in use.
The object of the invention is to create a quiet helically toothed internal gear machine that can be operated with low wear and high efficiency in the high-pressure range.
According to the invention, this problem is solved by an internal gear machine comprising a housing which forms a cavity in which an internally toothed ring gear and an externally toothed pinion are arranged, the toothings of which are in meshing engagement with one another in certain regions and the axes of rotation of which run parallel to and spaced apart from one another, at least one filler piece resting against the first and second toothings, which divides the cavity into two fluidically separate regions, and the toothing is designed as helical toothing or arrow toothing, it is advantageously possible to provide an internal gear machine suitable for high-pressure operation with helical toothing, which has a high degree of mechanical smooth running and a high degree of efficiency. This avoids both noise due to the flowing hydraulic fluid and wear minimization on the rotating parts or the housing parts adjacent to the rotating parts.
In a preferred embodiment, it is provided that the surfaces axially closing the cavity have non-congruent pressure fields whose control edges are rotated relative to each other in the circumferential direction, preferably by the face section twist between the front and rear sides of the gearbox specified by helical toothing. The advantage of this is that tilting moments on the gearbox formed by the ring gear and pinion due to the helical toothing can be optimally compensated. In addition to axial and radial gap sealing, which leads to a high degree of efficiency, this also significantly supports smooth running and wear-free operation of the internal gear machine.
Furthermore, in a preferred embodiment of the invention, it is provided that the axial boundaries of the cavity on both sides each have at least one mutually non-congruent hydrostatic pressure field, which are designed such that a thrust exerted by the helically toothed pinion and helically toothed ring gear in the region of the filler piece, acting axially on one side, and a thrust acting in the region of the tooth engagement of ring gear and pinion, acting axially in the opposite direction to the first acting axial thrust, are at least partially hydrostatically compensated in terms of area. In this way, the tilting moment resulting from the intended use of the internal gear machine according to the invention can be optimally balanced.
In particular, if the non-congruent hydrostatic pressure fields balance the respective axially opposite thrusts in the region of the filler piece and in the region of the tooth engagement by at least 20%, at least 30%, at least 40%, at least 50%, at least 60%, at least 70%, at least 80% or at least 90% in terms of area, the internal gear machine can be used as intended with very smooth running and high wear resistance.
It is preferably provided that the cavity accommodating the internally toothed ring gear and the externally toothed pinion is axially bounded by at least one axial disk, the axial disk has at least one fluid connection between the cavity and a pressure field provided on the side of the axial disk facing away from the cavity which is in connection with the fluid connection and the side of the axial disk facing the cavity has at least one hydrostatic surface which is arranged non-congruently with the side facing away from the cavity and which is in operative connection with the pressure field, it is advantageously possible to achieve axial and radial gap compensation in a simple manner, so that such helically toothed internal gear machines can also be operated with high efficiency in the high-pressure range. In addition, it is advantageously possible to compensate for a tilting moment acting on the gearbox consisting of ring gear and pinion. The pressure field facing the cavity is smaller than the pressure field facing away from the cavity.
In accordance with the invention, non-congruently arranged surfaces are understood to mean that the pressure surfaces on both sides of the axial disc or the axial discs are different and/or the pressure fields have sections that increase and/or decrease in size (for example pockets or the like) when viewed in the circumferential direction of the axial disc and/or their control edges are twisted relative to one another in the circumferential direction
In a preferred embodiment of the invention, it is provided that the opposing, non-congruent hydrostatic surfaces comprise relief grooves and/or pressure pockets. This makes it possible in a simple manner, through the arrangement and dimensioning of the relief grooves and/or pressure pockets, to form the opposing, non-congruent hydrostatic surfaces in such a way that hydrostatic relief of the tilting moments caused by the helically toothed pinion and ring gear can be absorbed depending on the operating point of the internal gear machine.
In a further preferred embodiment of the invention, it is provided that the toothings have a helix angle whose relative twist of the face section tooth contour from the front side of the gearbox to the rear side of the gearbox preferably corresponds to at least half a tooth pitch, in particular preferably a full tooth pitch. The front or rear side of the gearbox refers to the end faces of the pinion and ring gear meshing with each other. This makes it advantageous to use a relatively large helix angle. The internal gear machine can therefore be operated with a particularly high degree of efficiency, whereby the tilting moment emanating from the helical toothing with the helix angle can be absorbed by the opposing, non-congruent hydrostatic surfaces. In particular, this makes it possible to achieve very smooth running in the high-pressure range.
Furthermore, in a preferred embodiment of the invention, it is provided that a degree of overlap of the tooth engagements of the toothings is >=2. This makes it advantageously possible for the engagement distance in the tooth engagement to lead to complete sealing of the tooth engagement despite the full tooth pitch of the face section twist, with a relatively small helix angle.
Furthermore, in a preferred embodiment of the invention, it is provided that the number of teeth of the external toothing of the pinion is more than 15 and the number of teeth of the internal toothing of the ring gear is more than 20. Due to the relatively high number of teeth in conjunction with the helix angle and the full pitch of the cutting twist, in addition to a very smooth running of the internal gear machine, a large seal is simultaneously ensured in the engagement region between the pinion and ring gear.
In a further preferred embodiment of the invention, it is provided that the axial disc and/or the housing in the region of the axial disc comprises an axial recess which results in the pressure field and which is preferably surrounded by a sealing system, in particular a sealing ring. This makes it advantageously possible to achieve hydrostatic relief of the axial thrust exerted by the helical toothing and of the tilting moment acting transversely to the axis of rotation in co-operation with the non-congruently arranged hydrostatic surfaces provided on the side facing the cavity, with a high degree of effectiveness.
Furthermore, in a preferred embodiment of the invention, the internal gear machine comprises, in addition to the one axial disc on the opposite side of the cavity, at least one further axial disc, which preferably has a pressure field facing the cavity and a pressure field facing the housing, which are connected to each other via a fluid connection. This improves the axial and radial gap sealing, as this axial disc also comes to rest against the end face of the gearbox depending on the operating point.
Furthermore, in a preferred embodiment of the invention, it is provided that the pressure fields of one axial disc are not congruent with the pressure fields of the further axial disc. As a result, depending on the pressure conditions that occur when the internal gear machine is used as intended, very precise compensation of the tilting moment exerted by the helical toothing can be ensured, depending on the operating point, while guaranteeing effective axial and radial gap sealing.
Finally, in a further preferred embodiment of the invention, it is provided that pressure fields in the housing and/or pressure fields of at least one axial disc surrounding the end face of the ring gear are designed to be inversely symmetrical to one another with respect to the gearbox centre plane. A four-quadrant mode of the internal gear machine is possible by means of pressure fields designed in this way, so that compensation or counteraction of the tilting moment emanating from the helical toothing can take place at any time, even at high pressures, irrespective of the direction of rotation of the pinion and ring gear and an application-specific alternating pressure side and associated axially alternating thrust direction.
According to the invention, the internal gear machine is operated as a pump, as a hydraulic motor in reversing operation, in pure left-right operation or in four-quadrant mode, depending on the desired application. Due to the mutually non-congruent hydrostatic surfaces arranged on the gearbox end face, optimum compensation of the tilting moment is possible at all times with different operating pressures.
The invention is explained in more detail below in embodiments with reference to the associated Figures, showing:
Both the external toothing 24 of the pinion 16 and the internal toothing 26 of the ring gear 18 are helically toothed.
A filler piece 30 is arranged within a crescent-shaped free space 28 formed between pinion 16 and ring gear 18. The filler piece 30 is supported on a stop pin 32 and consists of an inner sealing segment 34 and an outer sealing segment 36. The gap between the inner sealing segment 34 and the outer sealing segment 36 is sealed by a sealing roller 38.
In the housing 12, there also are pressure pockets 40 and 42, each of which is connected to a fluid connection 44 or 46 of the internal gear machine 10.
The design and mode of operation of such an internal gear machine 10 are known to the skilled person, so that a more detailed description is omitted here. During operation of the internal gear machine 10, the pinion 16 is driven by a drive shaft 52 (
Fluid in the tooth gaps of the internal toothing 26 and the external toothing 24 then moves along the filler piece 30 with the tooth gaps and reaches the tooth engagement region of pinion 16 and ring gear 18. The fluid is displaced into the pressure pocket 42 and thus to the pressure connection 46 through the indicated radial bores 48 of the ring gear 18.
Also shown is the stop pin 32 against which the filler piece 30 rests with its inner sealing segment 34 and outer sealing segment 36. The sealing roller 38 is positioned between the sealing segments 34, 36.
The axial disc 48 delimiting the cavity 14 is also shown. The axial disc 48 has at least one fluid connection 54.
In the embodiment example shown, a total of 4 fluid connections 54 are provided, which are arranged spaced apart from one another in the circumferential direction of the axial disc 48 and have different diameters.
The fluid connections 54 are provided at the base of a pressure field 56 integrated into the axial disc 48. The pressure field 56 is formed by a kidney-shaped recess within the axial disc 48 on the side 58 of the axial disc 48 facing the cavity 14.
At least one relief groove 60 (also referred to as a control slot) extends from the pressure field 56 counterclockwise to the direction of rotation shown in
The end of the pressure field 56 opposite the relief groove 60 has at least one pressure pocket 62, which extends radially outwards over the circumference of the external toothing 24 of the pinion 16.
The axial disc 48 has an opening 66 by means of which the axial disc 48 is fixed in the internal gear machine 10 via the stop pin 32, which engages through the opening 66.
In the example shown, the axial disc 48 has only one fluid connection 54. Two relief grooves 60 and 60′ extend from the pressure field 56 on side 58. The pressure field 56 further comprises the radially outwardly directed pressure pocket 62 and an inwardly arranged pressure pocket 62′ on the opposite side.
The axial disc 48 also has a radially outwardly directed end groove 68 on its side 58.
The right-hand illustration in
The pressure field 70 is also formed by a trough-shaped recess in the axial disc 48. The pressure field 70 is surrounded by a sealing ring 72, via which the axial disc 48 rests against the housing 12.
The same parts as in
The axial disc 50 also has a fluid connection 80, which extends from the base of a pressure field 82 towards a pressure field 84 on the side 78 of the axial disc 50. The pressure fields 82 and 84 are each formed by trough-shaped recesses on the sides 76 and 78 of the axial disc 50.
The pressure field 82 has at least one pressure pocket 86, which extends in the opposite direction to the direction of rotation of the pinion 16.
The pressure pockets 86 and 86′ extend from the pressure field 82 on side 76. The pressure field 82 also has a pressure pocket 88 extending in the circumferential direction at the opposite end.
The axial disc 50 is also fixed to the stop pin 32 via the opening 90.
As
The illustrations in
The housing-side pressure fields 70 and 84 of both axial discs are preferably each designed with a surface centre of gravity congruent to the dynamically pressure-loaded surfaces defined by the pressure fields and the helical toothing in rotation on the respective end faces 58 and 76 of the axial disc in question, wherein preferably the overall pressure-loaded surfaces of pressure field 70 and 84 are each designed such that they exert a slightly increased pressure on both sides in the direction of the gearbox and thus the axial discs 48, 50 come to rest against the gearbox at every operating point. This effectively seals the end face of the gearbox.
The internal gear machine 10 shown in
By driving the drive shaft 52, a fluid, for example hydraulic oil, is drawn in via the fluid connection 44 and enters the cavity 14. The fluid is pumped past the filler piece 30 to the pressure connection 46 of the internal gear machine 10 via the meshing toothing of the pinion 16 and ring gear 18 in a manner known per se. The pressure fields 70 and 84 of both axial discs 48, 50 are supplied with pressure oil on the pressure side via the corresponding bores 54 and 80, so that both axial discs are brought to rest against the gearbox.
Due to the helical toothing of pinion 16 and ring gear 18, a tilting moment occurs transverse to the longitudinal axis 20 of pinion 16 and the longitudinal axis 22 of ring gear 18. This tilting moment is generated by the pressure oil present on the helical toothing in the region of the filler piece 30 and in the region of the tooth engagement by the drive torque of the pump and the pressure oil also present between the ring gear 18 and pinion 16. Both regions generate opposing radially offset axial thrusts, which cause the toothings to tilt
This tilting moment is compensated for by the design of the axial discs 48 and 50. This compensation is achieved by the hydrostatic surfaces 62, 62′ and 86, 86′ arranged on the axial discs in a mutually non-congruent manner at the level of the respective line of action of the opposing axial thrusts in the region of the filler piece 30 and in the tooth engagement, which are each connected to the pressure fields 56 and 82. Particularly in the case of internal gear machines 10 that are operated at high pressures, for example 250 to 350 bar, this results in optimum compensation of the tilting moment while simultaneously maintaining axial and radial gap compensation. Radial and axial sealing can therefore be achieved with high efficiency.
The different, asymmetrical design of the pressure fields 56 and 70, in particular the relief grooves 60, 60′ and the pressure pockets 62, 62′ as well as the arrangement of the control edges 57 and 71, can counteract the tilting moment of pinion 16 and ring gear 18 and compensate for the tilting moment.
In further embodiments not shown, the pressure fields 70 and/or 84 can also be formed in the wall of the adjacent housing 12 instead of in the axial disc 48 or 50. Proportional formations of the pressure field 70 and/or 84 in the axial discs 48 and/or 50 and the housing 12 are also possible.
The compensation options described above, in particular of the tilting moment of pinion 16 and ring gear 18 transverse to their longitudinal axes 20 and 22, make it possible to realise high-performance helically toothed gear fluid machines in the high-pressure range.
The external toothing of the pinion 24 and the internal toothing of the ring gear 26 can be used with a high number of teeth, for example more than 15 teeth for the pinion 16 and more than 20 teeth for the ring gear 18. A degree of overlap of the tooth engagement of pinion 16 and ring gear 18 can be at least two, i.e. the region in which pinion 16 and ring gear 18 mesh completely with one another can be at least 2 teeth.
For example, a face section twist (helix angle) of the toothing can be 22.5 degrees with a number of teeth 19 on the pinion 16 and a gearbox width of 20 mm.
The helix angle depends on the number of teeth and the width of ring gear 18 or pinion 16 and can therefore vary.
It is possible to provide an internal gear machine with a large number of teeth on both the pinion 16 and the ring gear 18, which is helically toothed and therefore runs very smoothly and is also suitable for conveying a fluid in the high-pressure range. As a result, the overlapping region of pinion 16 and ring gear 18 between the pressure region and the suction region within the cavity 14 is well sealed, since a larger number of fully meshing teeth of the internal toothing 26 of the ring gear 18 or the external toothing 24 of the pinion 16 is possible along the engagement distance between pinion 16 and ring gear 18.
As can be seen further in
The internal gear machine 10 described in the previous figures can be operated in reversing operation, i.e. both as a pump and as a motor. During pump operation, a fluid is sucked via the fluid connection 44 (suction side) and discharged under pressure at the fluid connection 46. For this purpose, the drive shaft 52 is driven in the manner described by an electric motor or in another suitable manner.
During motor operation, a pressurised fluid is fed into the fluid connection 46 so that the pinion 16 and ring gear 18 are set in rotation. The fluid is conveyed along the filler piece 30 to the fluid connection 44 via the pockets formed between the toothing. Due to the rotary movement of the pinion 16, an output torque can be tapped at its drive shaft 52 during motor operation.
The same parts as in the previous figures are designated with the same reference numerals and are not explained again.
The difference to the previous figures lies in the design of the axial discs 48′ and 50′ and the filler piece 30′. In this respect, only the differences are discussed here and reference is made to the previous description with regard to the other parts and functions.
The axial discs 48′ and 50′ are designed here as fully circumferential discs. This means that the axial disc 48′ has both a pressure field 56 on its side 58′ and a pressure field 82 opposite the drive axis 52. Correspondingly, the outer side of the axial disc 50′, which can be seen in the left-hand illustration in
In addition to the sealing segments 34 and 36, the filler piece 30′ also has an inner sealing segment 34′ and an outer sealing segment 36′ on the side opposite the stop pin 32, which are constructed and arranged as a mirror image of the sealing segments 34 and 36.
This is clearly shown in the right-hand illustration of
It can also be seen that the inner side 76′ of the axial disc 50′ has a pressure field 56 and a pressure field 82.
The outer side 64′ of the axial disc 48′ has a pressure field 84 and a pressure field 70.
The axial discs 48′ and 50′ are therefore inversely symmetrical to each other on their inner sides 58′ and 76′ and on their outer sides 78′ and 64.
With regard to the design and function of the pressure fields provided in the axial discs 48′ and 50′ with their relief grooves and pressure pockets as well as control edges, reference is made to the explanation of the preceding figures.
This design makes it possible to compensate for a tilting moment of pinion 16 and ring gear 18 at any time, even with an internal gear machine 10 that can be operated in so-called four-quadrant mode, regardless of its operating mode. At the same time, axial and radial gap sealing is guaranteed.
This means that such an internal gear machine 10 operating in four-quadrant mode can also be operated with high efficiency in the high-pressure range.
The internal gear machines 10 according to the invention can also be used according to the invention, for example, in the following ways: electrohydraulic/hydropneumatic chassis control systems, electrohydraulic steering systems, decentralised hydraulic applications in electrified vehicles.
| Number | Date | Country | Kind |
|---|---|---|---|
| 10 2021 120 395.3 | Aug 2021 | DE | national |
This application is the U.S. National Stage of International Application No. PCT/EP2022/070286, filed Jul. 20, 2022, which claims foreign priority benefit under 35 U.S.C. § 119 of German Patent Application Nos. 10 2021 120 395.3, the disclosures of which are incorporated herein by reference.
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/EP2022/070286 | 7/20/2022 | WO |
