EP0254077B1 - Internal gear pump - Google Patents

Description

The invention relates to an internal gear pump with a driving pinion and a ring gear, in which on the pressure side the trailing flanks of the pinion teeth (sealing flanks of the pinion) with the corresponding counter flanks of the ring gear (sealing flanks of the ring gear) in the area between the intersection of the tip circles and the Rolling point with a degree of coverage greater than 2 are engaged.

Internal gear pumps of this type serve as control pumps for hydraulic fluids. In this embodiment, they are provided with a large number of outlet openings, the pitch of which is smaller than or equal to the tooth pitch. These outlet openings all or in groups lead to a common pressure channel and - with one exception if any - all outlet openings of a group are closed by a check valve.

In this embodiment, the internal gear pump has a delivery characteristic that is speed-dependent only up to a certain speed. The delivery is constant above this speed. The threshold speed can be adjusted by adjusting a throttle in the inlet.

Such an internal gear pump is known from DE-OS 34 44 859. This internal gear pump has the peculiarity compared to conventional internal gear pumps that there is a degree of coverage of at least 2, so that the internal gear pump has at least two, but preferably three or more tooth rows which are closed off from one another on the suction and printing page forms.

Compared to all other known control pumps, the delivery characteristics of which show no speed-dependent delivery or whose delivery can be set independently of the speed, the known internal gear pump has the advantage of the robust construction, in which the delivery characteristic can be set without additional mechanical effort. Control pumps of this type are used with particular advantage for driving motor vehicle engines, the speed of which fluctuates greatly. They are used there as hydraulic pumps or lubricating oil pumps, since with these pumps the maximum delivery rate can be limited without loss of performance at a certain, relatively low speed.

The invention has for its object to further reduce the power requirement of the known internal gear pump.

For this purpose, it is proposed to design the tooth flanks so that the driving flanks have less coverage than the sealing flanks. The profile overlap of a toothing represents the ratio of the engagement length to the pitch. In a gear pump, the profile overlap is, among other factors, also decisive for the sealing effect of the sealing tooth flanks. Deviating from the technical term of profile overlap, in the context of this application the overlap indicates the number of tooth pairs on the suction or pressure side that are in engagement with one another on average, i.e. touch or face each other with little backlash that causes the seal.

This solution is particularly advantageous in the low pressure range - up to approx. 20 bar - and in particular in the automotive sector, where it is important to achieve a maximum delivery rate at relatively low speeds, while keeping the idle power and in particular the mechanical power consumption of the pump low. A preferred field of application are lubricating oil pumps which are arranged in the sump of the motor vehicle engine.

The proposed solution includes that the tooth flanks of the pinion and / or the tooth flanks of the ring gear on the driving side and the sealing side are not made mirror-symmetrical. It is essential that the tooth flanks of the driving side have only a relatively small area in which the driving flanks of the pinion and ring gear can come into engagement with one another (engagement area). This meshing area lies - for pinion and ring gear alike - between the pitch circle and the tip circle and begins at the pitch circle.

Outside this engagement area, the tooth flanks created by conventional toothing processes can be removed or deformed in such a way that no tooth engagement occurs. The asymmetrical tooth shape proposed here can advantageously also be produced in a sintering process, since a corresponding shaping is possible here without subsequent mechanical processing.

The engagement area is preferably chosen to be so large that the degree of coverage is between 1 and 2. On the one hand, this relatively low degree of coverage results in a considerable reduction in the mechanical power consumption. On the other hand, with this degree of coverage, in particular in the case of hydraulic pumps in the low-pressure range, there is no impermissible wear.

An essential part of the power consumption of the known internal gear pump is also based on the fact that the tooth cells become very narrow in the area of the dead center and there are very high flow rates. To solve this problem, it is therefore further proposed - preferably in combination with the aforementioned solution - to expand the cross section of the base of the tooth gaps of the ring gear between the root circle and the pitch circle considerably. The bottom of the tooth gap can have an essentially circular cross section in this area.

The reduction of the flow velocities and the power consumption caused thereby serves alone or in combination with the other measures of this invention that the outlet openings between the line of engagement and the outer circumference of the ring gear, preferably between the line of engagement and the root circle of the ring gear, are created, with only one towards the line of engagement narrow sealing web is retained. The cross section of the openings is essentially adapted to the cross section of the teeth of the ring gear minus a narrow sealing strip. The cross Cut of a tooth thus completely covers the outlet opening, but the area of the outlet opening comes as close as possible to the area of the tooth cross section. All of the outlet openings meshing with a tooth cell are therefore covered by the tooth cross section of the ring gear and are therefore always separated from one another, so that no short circuit can occur between the tooth cells via the outlet openings. On the other hand, the openings cover the resulting tooth cells over a large area. In order to further promote this large-area coverage, the outlet openings are guided beyond the base circle of the ring gear and the bottom of the tooth gaps is widened in a funnel shape by a corresponding bevel between the end face and the bottom of the tooth. This also results in a reduction in throttle losses.

The invention is described below using an exemplary embodiment.

Show it

• 1 shows the radial section of the embodiment with outlets which are arranged on both end faces of the pump housing, the outlet openings on one side being offset by half a division in each case with respect to the outlet openings on the other side;
• Figure 2 shows the axial section through the embodiment.
• Fig. 3 shows the axial section (partially) through the ring gear.

The outer wheel 1 is freely rotatably mounted in the housing 31. The outer wheel 1 has an internal toothing 2. The cylindrical housing 31 is closed on both sides by the covers 32 and 33. In the cover 32, the shaft 34 is rotatably supported and driven by the motor vehicle engine, not shown. The inner wheel 3 is rotatably mounted on the shaft 34. The inner wheel 3 has an external toothing 4 which is in engagement with the internal toothing 2 of the outer wheel 1. The interior of the pump, which lies outside the meshing of the teeth, can be filled with a sickle that largely conforms to the tip circles of the gears. In the cover 33 there is the inlet channel 35 (see also FIG. 2). The inlet channel 35 is connected to the sump 36 via a throttle 37. A pressure control valve 39 is located in a bypass 38, which is connected parallel to the throttle channel 37. The piston 40 of the pressure control valve controls the opening of the bypass channel 38 to the sump 36 with its control edge 41. The piston is on one side with a spring 42 charged. On the opposite side, the piston in control chamber 43 is acted upon by the outlet pressure in pressure channel 56 via control line 44. The outlet side of the pump will be discussed later. The function of the pressure control valve 39 as a function of the outlet pressure is described below. As long as there is no or only a low outlet pressure in the control line 44 and the control chamber 43, the piston with its control edge releases the flow from the inlet 45 to the outlet 46. An unlimited amount of lubricating oil can now flow from the sump 36 to the pump both via the throttle 37 and the bypass channel 38. When the pressure in the control chamber 43 increases and the spring force overcomes, the inlet 45 on the pressure control valve 39 is partially or completely closed off from the outlet 46. Now only a throttled flow of lubricating oil flows through the throttle 37 and possibly via the pressure control valve 39 from the sump 36 to the inlet 35 of the pump. If the outlet pressure rises further, the pressure control valve acts as a pressure relief valve. The spring 42 is compressed so far that the front control edge 47 opens the pressure line 44 with respect to the outlet 46 to the sump.

• Die Pumpe bildet - wie Fig. 1 zeigt - auf der Auslaßseite zwischen den miteinander kämmenden Zähren des Außenrades 1 und Innenrades 3 drei in Umfangsrichturg und Axialrichtung abgeschlossene Zellen, die über Einlaßkanal 35 mit ÖI ganz oder teilweise gefüllt worden sind. In den Deckel 33 sind drei Auslaßöffnungen 48.1, 48.3, 48.5 eingebracht. In den Deckel 32 sind zwei Auslaßöffnungen 48.2, 48.4 eingebracht. Die Auslaßöffnungen des Deckels 33 sind gegenüber den Auslaßöffnungen des Deckels 32 versetzt angeordnet. In der Projektion auf eine Normalebene überdecken sich die Auslaßöffnungen im Deckel 33 bzw. 32 nicht - wie Fig. 1 zeigt. Die Auslaßöffnungen schmiegen sich mit ihrer radial inneren Kante 27 (Innenkante) eng an die Eingriffslinie 11 an, und zwar derart, daß zwischen der Eingriffslinie 11 und der Innenkante 27 lediglich ein schmaler, jedoch für die Abdichtung ausreichend dichtender Dichtsteg 28 stehenbleibt. Die Breite der Auslaßöffnungen 48.1 bis 48.5 ist so gewählt, daß die Auslaßöffnungen von dem Querschnitt der Zähne 2 des Hohlrades 1 bei entsprechender Stellung der Zähne überdeckt werden, wobei in Umfangsrichtung ebenfalls ausreichende Dichtflächen stehenbleiben. In der radialen Höhe erstrecken sich die Auslaßöffnungen bis in den Bereich des Außenumfangs des Hohlrades und jedenfalls bis zum äußersten Bereich, mit dem der Grund der Zahnlücken des Hohlrades 1 auf der Stirnfläche der Deckel 32, 33 mündet.

To the outlet side of the pump:

• The pump forms - as shown in FIG. 1 - on the outlet side between the meshing teeth of the outer wheel 1 and the inner wheel 3 three cells which are closed in the circumferential direction and in the axial direction and which have been completely or partially filled with oil via inlet channel 35. In the cover 33, three outlet openings 48.1, 48.3, 48.5 are introduced. Two outlet openings 48.2, 48.4 are introduced into the cover 32. The outlet openings of the cover 33 are arranged offset with respect to the outlet openings of the cover 32. When projected onto a normal plane, the outlet openings in the cover 33 or 32 do not overlap - as shown in FIG. 1. The outlet openings nestle closely with their radially inner edge 27 (inner edge) to the line of engagement 11, in such a way that only a narrow, but sufficiently sealing sealing web 28 remains between the line of engagement 11 and the inner edge 27. The width of the outlet openings 48.1 to 48.5 is selected so that the outlet openings are covered by the cross section of the teeth 2 of the ring gear 1 with the teeth in a corresponding position, sufficient sealing surfaces also remaining in the circumferential direction. In the radial height, the outlet openings extend into the area of the outer circumference of the ring gear and in any case up to the outermost area with which the bottom of the tooth gaps of the ring gear 1 opens on the end face of the covers 32, 33.

• Die Zähne des Hohlrades werden nach einem Verzahnungsgesetz hergestellt, auf das später noch eingegangen wird. Dieser nach dem Verzahnungsgesetz entstehende ideale Zahnlückengrund ist für eine Zahnlücke punktiert eingezeichnet und mit 29 bezeichnet. Dieser Zahnlückengrund wird jedoch bei allen Zahnlücken und über die gesamte axiale Länge der Zahnlücken wesentlich erweitert und in den Ausführungsbeispielen durch Zahnlückengrund 30 gebildet. Zahnlückengrund 30 stellt in den Ausführungsbeispielen den halben Mantel eines Kreiszylinders dar, dessen Achse jeweils auf der Symmetrieebene der Zahnlücke und im wesentlichen auf dem Wälzkreis oder geringfügig radial außerhalb des Wälzkreises 7 des Hohlrades liegt. Darüber hinaus ist der Zahnlückengrund an seinen beiden Enden noch einmal mit einer trichterförmigen Erweiterung 26 versehen. Die trichterförmige Erweiterung 26 erstreckt sich radial bis nahezu an den Außenumfang des Hohlrades. Die trichterförmige Erweiterung 26 kann sich auch in Umfangsrichtung erstrecken. Sie liegt jedoch jedenfalls radial außerhalb des Wälzkreises 7 des Hohlrades 1. Wenn bei einer erfindungsgemäßen Pumpe der Ölaustritt nur einseitig vorgesehen wird, so befindet sich auch die trichterförmige Erweiterung nur an der betreffenden Seite.

The following results from FIGS. 1 and 3 for the design of the bottom of the tooth gaps in ring gear 1:

• The teeth of the ring gear are manufactured according to a toothing law, which will be discussed later. This ideal tooth space base, which is created according to the gearing law, is shown in dotted lines for a tooth space and designated 29. However, this tooth space base is significantly expanded for all tooth spaces and over the entire axial length of the tooth spaces and is formed by tooth space base 30 in the exemplary embodiments. In the exemplary embodiments, tooth space bottom 30 represents half the shell of a circular cylinder, the axis of which is in each case on the plane of symmetry of the tooth space and essentially Chen on the pitch circle or slightly radially outside of the pitch circle 7 of the ring gear. In addition, the bottom of the tooth space is again provided with a funnel-shaped extension 26 at both ends. The funnel-shaped extension 26 extends radially to almost the outer circumference of the ring gear. The funnel-shaped extension 26 can also extend in the circumferential direction. However, it is in any case radially outside of the pitch circle 7 of the ring gear 1. If the oil outlet is only provided on one side in a pump according to the invention, the funnel-shaped extension is also only on the relevant side.

The previously described outlet openings 48.1 to 48.5 now extend radially in any case so far outwards that they also cover the funnel-shaped extensions 26 on the end faces of the outer wheel 1.

2, only one of these outlet openings can be seen in each cover 32, 33. These outlet openings are designated there by 48. Each of the outlet openings is connected to an outlet channel 49 drilled in the cover 32, 33. The outlet channel is also directed radially outwards, as shown in FIG. 2. Therefore, each outlet channel 49 opens out on the outside of the cover 32 and 33 as close as possible to the housing 31. An outlet housing 50 is placed on each cover 32, 33 in a pressure-tight manner. Each outlet housing 50 forms an outlet chamber which is connected on one side to the outlet openings 48.1, 48.3, 48.5 and on the other side to the outlet openings 48.2, 48.4 each via a pressure channel 49 and a bore 52. The bores 52 (cf. FIG. 1) are each closed by a check valve, with the exception of the bore which is connected to the outlet opening 48.5. The outlet opening 48.5 is located at the end of the pressure zone immediately before the pitch point. Both outlet chambers are connected to the common pressure channel 56.

The check valves on both sides are formed by an n-shaped plate, which is screwed against the wall 53 of the outlet housing 50. The tongues protruding from the common crossbeam 55 of the check valve 54 cover the bores 52. Therefore, these tongues act as check valves. Each check valve only releases the connection from the respective pressure cell formed between the teeth via one of the outlet openings 48, pressure channels 49 and bores 52 if the pressure of the outlet cell is at least equal to the outlet pressure in the outlet chamber 51. The last and smallest pressure cell is directly connected to the outlet chamber via opening 48.5 and corresponding channels 49, 52.

Each outlet chamber 51 has an outlet which leads into the common pressure oil channel 56.

1, the teeth of the ring gear 1 are asymmetrical. First, both flanks of each tooth are formed according to a special toothing law. This interlocking law ensures that there is a high degree of coverage that is greater than 2, preferably greater than 3. This has the effect that the teeth are in engagement with one another in approximately the entire rotational range between the intersection of the two tip circles 5 and 9 and the pitch point and, as a result, more than two tooth cells are formed by two successive tooth pairs in each case. These tooth cells are mutually closed in the circumferential direction. This gearing law includes that the driving flanks of the inner wheel 3 and outer wheel 1 also have a correspondingly large degree of coverage. It is now provided that the degree of coverage is less on the driving side of the teeth than on the sealing side of the teeth. This means: The tooth flanks, which lie sealingly on top of each other in the pressure zone between the intersection of the tip circles and the pitch point and form the mutually closed tooth cells, are produced according to the toothing law described above. These flanks are referred to as sealing flanks in the context of this application.

However, the flanks of the teeth of ring gear 1 and pinion 3, which serve to transmit the torque between inner wheel 3 and ring gear 1 (driving flanks), are produced with a lower degree of coverage, which is preferably between 1 and 2. This is done in that only a partial area of the driving flanks of the outer wheel 1 and / or the inner wheel 3 is produced according to the toothing law (engagement area of the flank). The engagement area 64 of the drive flanks of the ring gear extends radially a little way inward from the pitch circle 7 of the ring gear. The cross-sectional area by which the driving flank of the ring gear deviates from the profile produced by toothing is designated by 65.

The engagement area 66 of the drive flanks of the inner wheel 1 extends radially a little outward from the pitch circle 8. The cross-sectional area of the tooth head by which the driving tooth flanks of the inner wheel 3 recede relative to the ideal tooth profile is designated by 67.

As mentioned, either the driving flanks of the ring gear or the driving flanks of the pinion or both can be provided with such cutouts 65 and 67, respectively. The latter solution has the advantage that only low flow velocities arise on the suction side of the pump. The engagement area 64 of the driving flanks of the ring gear and / or the inner wheel, which is formed according to the gearing law, is dimensioned such that on the one hand at least one pair of teeth of the ring gear and the inner wheel are always in engagement with one another, but on the other hand fewer tooth pairs are in engagement on the driving side than on the Sealing side. The degree of coverage on the engagement side is preferably not greater than 2 due to the correspondingly short design of the engagement areas.

• Bei niedrigem Druck in der Auslaßkammer 51 verschiebt die Feder 42 den Kolben 40 - in Fig. 2 - nach links. Die Pumpe wirkt nun wie eine normale Innenzahnradpumpe. Der Schmierölstrom fließt über Drossel 37 und Bypasskanal 38 zum Einlaß. Sämtliche Zahnlücken werden maximal gefüllt und auf der Auslaßseite wieder ausgedrückt. Der Grad der Füllung hängt davon ab, wie weit auch der Bypass 38 gedrosselt ist. Hierauf wird später noch eingegangen. Bei niedrigen Drehzahlen erfolgt jedenfalls eine vollständige Füllung.

For the function of the exemplary embodiment according to FIG. 2 (lubricating oil pump):

• At low pressure in the outlet chamber 51, the spring 42 displaces the piston 40 - in FIG. 2 - to the left. The pump now acts like a normal internal gear pump. The lubricating oil flow flows through throttle 37 and bypass channel 38 to the inlet. All tooth gaps are filled to the maximum and expressed again on the outlet side. The degree of filling depends on how far the bypass 38 is throttled. This will be discussed later. In any case, full filling takes place at low speeds.

This operating state is maintained at low speeds of the motor vehicle engine. The lubricating oil flow is therefore proportional to the demand according to the speed.

If the pressure in the pressure channel 56 increases as the speed increases, the bypass 38 is initially closed or at least severely throttled by the pressure control valve 39. There is now essentially only a throttled oil flow via throttle 37 on the inlet side. Therefore, the tooth gaps on the inlet side are only partially filled. There is also a vacuum in the tooth gaps. As a result, the pressure in the tooth cells on the outlet side is initially lower than the pressure in the outlet chamber 51. Therefore, the respective tongues of the check valve 54 remain closed. However, as the size of the cells on the outlet side shrinks, the pressure in the cells increases. It only opens the tongue of the check valve for which the pressure of the cell is greater than or equal to the pressure in the outlet chamber 51. The result of this is that the pump now only delivers a constant, independent oil quantity. It is therefore not necessary, even with increasing speed, to discharge an excessive amount of oil with corresponding power losses, as is the case with conventional systems. On the other hand, if the need for lubricating oil increases, e.g. due to wear, the threshold pressure in the control pressure chamber 43 is only reached at a higher speed. Therefore, the bypass 38 is also closed later. As a result, the lubricating oil pump automatically adapts to an increased demand. The lubricating oil pump therefore meets the increasing need for lubricating oil throughout the life of the motor vehicle engine. On the other hand, the lubricating oil pump works economically even with a new engine with a relatively low need for lubricating oil, since with this lubricating oil pump it is avoided that an unneeded feed portion must be returned to the sump with losses.

In addition, the lubricating oil pump also meets other requirements of special operating conditions. For example, occur that the lubricating oil heats up excessively or that engine parts have to be cooled by lubricating oil due to special performance requirements. In this case, as shown in FIG. 2, a further short-circuit channel 58 is provided between the inlet 35 of the pump and the oil sump 36. An electromagnetically switched valve 59 is located in this short-circuit channel. This valve is actuated via signal line 60 and amplifier 61 by a temperature sensor 62. The temperature sensor can e.g. the oil temperature or the temperature of a machine part, e.g. Pistons to be detected. It is also possible to use a different measuring instrument, e.g. Speed counter to use. The message line can also be used to record other extraordinary operating conditions. In any case, the valve 59 serves the purpose of meeting an extraordinary need. It is assumed here that the sum of the oil flow, which is conveyed by throttle 37 on the one hand and via bypass 38 on the other hand, is still throttled and therefore only partial filling of the cells of the internal toothing takes place even at open pressure control valve at speeds that exceed one certain threshold speed. Fig. 2 meets this requirement in that a further throttle 63 in the bypass 38 is indicated as a symbol.

To meet an extraordinary need, it is also possible to switch the spring side 42 of the pressure control valve 39 by a suitable valve from a low pressure, at which a relatively low outlet pressure is regulated on the outlet side of the pump via line 44, to a low pressure, at which the outlet pressure is increased accordingly. As shown in Fig. 3, e.g. the pressure relief valve through the valve 68 which is electromagnetically e.g. is switched by the temperature of a machine part, either to the pressure before the throttle 37 or to the pressure behind the throttle 37.

It has already been pointed out that the effectiveness of the pump depends on the toothing being designed in such a way that the teeth in the outlet area between the intersections of the head circles are in engagement with one another and - taking into account the viscosity of the hydraulic oil - form closed cells.

The configuration of the exemplary embodiment shown avoids that unnecessarily high power losses occur as a result of the cell formation and the emptying of the cells. On the one hand, this is achieved in that the degree of coverage on the drive side of the teeth is less than on the sealing side of the teeth. A balance must be made here between avoiding mechanical loss of performance on the one hand and increased wear on the other. This consideration depends on the application of the pump. Power losses play a smaller role in high-pressure hydraulic pumps. On the other hand, there is a considerable surface pressure between the tooth pairs with a correspondingly high risk of wear and therefore a relatively high degree of coverage will also be selected on the drive side of the teeth in high-pressure pumps. For pumps in the low pressure range, such as lubricating oil pumps in motor vehicles, hydraulic pumps for steering assistance or other consumers, you will be able to work with an overlap on the drive side of the teeth without increasing the wear, the two 1 and 2 is because, due to the low pressure, wear-promoting surface pressure cannot be expected.

By expanding the base of the tooth space, the flow rate of the oil to be pressed out of the tooth space can be very greatly reduced, especially in the area shortly before bottom dead center. In principle, the expansion of the tooth gap of the ring gear can be driven radially outside of the pitch circle 7 until the stability limit of the ring gear is reached. In one embodiment, the maximum flow rate when the oil was pressed out was reduced from 20 m / sec to 5 m / sec. This reduction in flow velocity also means a reduction in hydraulic power losses.

The same purpose serves on the one hand the funnel-shaped expansion of the tooth space base on the end faces of the ring gear and the corresponding dimensioning of the outlet openings.

The fact that the outlet openings are arranged radially outside the line of engagement while maintaining a narrow but sufficient sealing strip ensures that there is no short circuit between successive tooth cells via the outlet openings. On the other hand, this enables the outlet openings to be made over a very large area. The area of the outlet openings is selected so that it is covered by the tooth cross section of the ring gear with sufficiently wide sealing surfaces in the circumferential direction. In this context, however, the outlet openings can be chosen to be very large and the outlet openings can furthermore be arranged with a smaller pitch than the tooth pitch. This ensures that there is always a large connection cross-section between the tooth cells and the outlet.

REFERENCE SIGN LISTING

• 1 outer wheel, ring gear
• 2 internal teeth
• 3 inner wheel, pinion
• 4 external teeth
• 5 tip circle outer wheel
• 6 foot circle outer wheel
• 7 Outer wheel pitch circle
• 8 Internal gear pitch circle
• 9 Head circle inner wheel
• 10 foot circle inner wheel, base circle
• 11 line of engagement
• 12 pitch point
• 13 intersection of the head circles
• 14 tooth height
• 15 gear module, large section
• 16 small section
• 17 center point, outer wheel
• 18 circle of centers of curvature
• 19 center of curvature
• 20 radius of curvature of the line of engagement
• 21 Outer gear pitch radius
• 22 pitch circle radius inner wheel
• 23 direction of rotation, web
• 24 arrow direction
• 25 center of inner wheel
• 26 funnel-shaped extension
• 27 edge, inner edge
• 28 sealing web
• 29 ideal tooth space reason
• 30 tooth gap reason
• 31 housing
• 32 lids
• 33 cover
• 34 wave
• 35 inlet
• 36 tank
• 37 throttle
• 38 bypass
• 39 pressure control valve
• 40 pistons
• 41 control edge
• 42 spring
• 43 control room
• 44 control line
• 45 inlet
• 46 outlet
• 47 front steering edge
• 48 outlet kidney
• 49 outlet duct
• 50 outlet housing
• 51 outlet chamber
• 52 hole
• 53 wall
• 54 check valve
• 55 crossbars
• 56 pressure channel
• 58 short-circuit channel
• 59 valve
• 60 reporting line
• 61 amplifiers
• 62 temperature sensors
• 63 throttle
• 64 area of engagement of the drive flanks of the ring gear
• 65 Deviation cross section of ring gear
• 66 Area of engagement of the drive flanks of the gear
• 67 Deviation cross section

Claims (7)

1. Internal gear pump having a driving pinion (3) and an internal geared wheel (1), in which on the delivery side the lagging flanks, which form the sealing flanks of the pinion, of the teeth of the pinion are meshed with the corresponding mating flanks, which form the sealing flanks of the internal geared wheel, of the teeth of the internal geared wheel in the region between the point of intersection of the tip circles and the pitch point with a contact ratio factor equal to or greater than 2 in such a manner that a plurality of mutually sealed-off tooth tells are formed, with a plurality of said tooth cells each being connected by at least one outlet having a check valve to the common pressure channel, characterised in that the driving flanks of the teeth of the pinion (3) and the corresponding driving mating flanks of the teeth of the internal geared wheel (1) on the induction side of the pump have a lower contact ratio factor than the sealing flanks on the delivery side.

2. Internal gear pump according to claim 1, characterised in that the contact ratio factor of the sealing flanks is equal to or greater than 3, and in that the contact ratio factor of the driving flanks is between 1 and 2.

3. Internal gear pump according to claim 1 or 2, characterised in that, the tooth spaces of the internal geared wheel, insofar as they lie outside of the pitch circle, are substantially extended in cross-section relative to the envelope curve of a pinion tooth.

4. Internal gear pump according to one of claims 1 to 3, characterised in that the outlet openings, less a small sealing strip, are slightly smaller than the tooth cross-section of the internal geared wheel.

5. Internal gear pump according to one of the preceding claims, characterised in that the outlet openings are located between the line of action and the peripheral circle of the internal geared wheel and, except for a narrow sealing surface, project closely up to the line of action.

6. Internal gear pump according to one of claims 3 to 5, characterised in that the base of the tooth spaces of the internal geared wheel at the face facing the outlet openings is extended in a funnel-like manner almost up to the periphery of the internal geared wheel.

7. Internal gear pump according to one of the preceding claims, characterised in that the outlet openings radially overlap the cross-section of the tooth spaces of the internal geared wheel.

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