EP0552443B1 - Gear machine - Google Patents

Description

The invention relates to a gear machine for liquids or gases with a housing which contains a gear chamber which has inlet and outlet openings, with an internally toothed toothed ring arranged in the gear chamber, and a pinion rotatably arranged within the toothed ring in the housing, which has one tooth less than the toothed ring has, is in engagement therewith and, when rotated between its teeth and the teeth of the toothed ring, forms circumferential enlarging and reducing liquid cells which conduct liquid from the inflow to the outflow, the tooth heads of the pinion and the tooth gaps of the toothed ring being in the form of epicycloids have, which are formed by rolling a first rolling circle (generating circle) on the pitch circle of the pinion or toothed ring, the tooth gaps of the pinion and the tooth tips of the toothed ring also having the shape of hypocycloids, which are formed by rolling a second rolling circle on the pitch circle of the pinion or toothed ring, and finally the radius of the first rolling circle is different from that of the second rolling circle.

The gear machine according to the invention can be used both as a pump for liquids or gases and as a motor driven by pressurized liquids or gases. However, the preferred field of application of the invention is the use as a liquid pump. In the following description and also in the claims, only liquid is spoken of for the sake of simplicity. In the claims, the term liquid should therefore also include gases.

The following explanation of the invention is made exclusively on the basis of a pump for liquids.

The gear machine according to the invention can be one in which the toothed ring is fixedly arranged in the housing, the pinion then rotating around the crank arm of a shaft, the latter being arranged centrally to the internal toothing of the pinion. However, the machine according to the invention is preferably one in which the toothed ring rotates in the gear chamber and the pinion mounted eccentrically to the axis of the toothed ring and the gear chamber rotates with a stationary shaft or about such an axis. A main area of application of the invention is the use of the machine designed as an internal ring gear pump as a lubricant and hydraulic fluid pump for internal combustion engines and automatic transmissions, where delivery pressures of up to 30 bar can occur. For this application, in which the pump pinion is preferably arranged in the extension of the crankshaft of the engine or the main shaft of the transmission or is carried by this shaft Internal gerotor pumps have proven themselves as quiet and low-vibration pumps. However, the requirements for smooth running of such pumps are constantly increasing due to the ever better smooth running of the motors and gears.

Most known and realized internal gear pumps or gerotor pumps for internal combustion engines and automatic motor vehicle transmissions work with trochoid gears, in which the tooth flanks of the ring gear or the pinion are delimited by circular arcs and the counter gear is defined by slip-free rolling in the toothing of the other wheel predetermined by the circular arcs.

Gear pumps of the type improved by the invention have long been known, for example from GB-PS 233 423 from 1925, or the publication "Kinematics of Gerotors" by Myron S. Hill, also from the 1920s. A modern application of the cycloid gearing for the above-mentioned use in internal combustion engines and automatic transmissions is described in DE-PS 39 38 346 by the applicant. The pump according to this German patent utilizes the excellent kinematic properties of the teeth and tooth gaps which have a complete cycloidal contour in an internal gear ring pump with the number of teeth difference one in order to support the gear ring with its toothing on that of the pinion, which is from the crankshaft of the engine or main shaft of the automatic transmission is carried. In this way, the relatively strong radial movement of the crankshaft can be compensated for by selecting the circumferential bearing of the toothed ring with air sufficient for this compensation. It is also possible to store the toothed ring with little play and then to provide a correspondingly large play between the shaft carrying the pinion and the pinion, the pinion with its toothing then being mounted in that of the toothed ring.

Such pumps represent a preferred field of application of the present invention.

From GB-A-223 257 a gear machine is also known in which the curve shape of tooth spaces and tooth heads have a flatter course than that of cycloid teeth, which consist of hypocycloids and epicycloids.

Pressure pulsations, i.e. flow pulsations, and the hammering of the teeth in the radial and tangential directions are primarily responsible for the undesirable noise development and the associated drop in efficiency of the known pumps. The flow pulsations are intensified by pinch oil pressure peaks that lead to vibrations in the gear wheel set. Gravitational noises act in the same sense, which arise primarily from the collapse of liquid vapor bubbles in the area of the pressure chamber of the pump.

The invention has in particular the object of making the known ring gear machines quieter, that is to say to reduce noise, which has a significant advantage when these machines are used as lubricating oil pumps in motor vehicle drive and transmission units. Another advantage achieved by this noise reduction is the improvement in efficiency and an increase in the service life of the gear ring machine.

The invention solves the problem according to the characterizing features of claim 1 in that the circumferential extent measured on the respective pitch circle of the pinion tooth gaps and toothed ring teeth limited by hypocycloids is 1.5 times to 3 times the circumferential extent measured on the respective pitch circle of the extent limited by epicycloids Sprocket teeth and tooth ring tooth gaps and that the epicycloids and hypocycloids are flattened towards their pitch circles to such an extent that the sum of the two flattenings corresponds to the required relatively large radial play between the tooth heads at the point opposite the point corresponds to the deepest meshing, while the gears mesh with each other at the point of deepest meshing with very little play.

The first-mentioned feature can also be formulated in such a way that the radius of the rolling circle producing the hypocycloids is equal to 1.5 times to 3 times the radius of the rolling circle producing the epicycloids.

In reducing the gerotor machine noise to a minimum, the invention is based on the fact that the course of the instantaneous displacement volume is primarily responsible for the flow pulsations - at least with precise manufacture and little play - in gerotor machines according to the preamble of claim 1. This in turn depends primarily on the position of the sealing points between the pressure chamber and the suction chamber of the machine over the angle of rotation of the pinion or the zan ring. Theoretically, that is, with a perfect toothing without play, the sealing points coincide with the intersection of the tooth flanks with the tooth engagement line. The sealing points in the area above the pressure and suction openings are irrelevant since the liquid cells separated by the sealing points are connected to one another by the suction and pressure openings. The only decisive factor is the position of the sealing points in the area of the deepest tooth engagement and in the area opposite this point. The theoretical line of engagement in gear ring machines is composed of three circles touching each other at the interface of the pitch circles and the connecting straight line of the two gear wheel centers, which are symmetrical to the connecting straight line of the two gear wheel centers and are halved by this straight line.

The cycloid gearing according to the preamble of claim 1 offers optimal engagement conditions in the most important area of deepest tooth engagement (in FIG. 1 above). However, this only applies if the play here is very small. The reduction of the backlash, however, is limited, among other things, by the fact that it is not possible for the series production to go below a certain degree of out-of-roundness of the toothed ring without excessive technical effort. The result of this is that, according to the prior art, the minimum play must still be large enough to prevent metallic contact between the pinion tooth tips and ring tooth tips opposite the point of deepest tooth engagement (in FIG. 1 below). The play necessary to ensure that the teeth are free from one another with respect to the point of deepest engagement in turn has the consequence that the "minimal tooth play" is still relatively large in the known toothings. This in turn has the consequence that the course of the path of the sealing point in the area of deepest tooth engagement deviates considerably from the theoretical course. In order to enable the least possible backlash in this area with a large backlash in the area opposite, according to the second characterizing feature of claim 1, either the interacting tooth spaces of the toothed ring and teeth of the pinion or the interacting teeth of the toothed ring and tooth spaces of the pinion are to such an extent flattened so that the tooth tips in the area opposite to the point of deepest tooth engagement will surely be free from each other. The flattening of the teeth therefore causes the relatively large tooth play in the area opposite the point of deepest tooth engagement. The flattening of the tooth gaps by the same amount compensates for the resulting increase in backlash in the area of deepest tooth engagement.

The flattening can of course also be distributed over the two cycloid groups mentioned above, i.e. the epicycloids and the hypocycloids. However, it is easier if you limit them to one of the two groups.

In this way, the gearwheels can mesh with each other in the area of deepest meshing with actually minimal play and very closely approximate theoretical maximum values. An unfavorable influence of the deviation of the sealing points between intermeshing teeth in the area of the deepest tooth engagement from the theoretical course is thereby minimized. The negative influence of such a deviation on the flow pulsation is thus reduced.

However, the delivery flow pulsation is reduced to a particularly high degree by the selected tooth thickness ratio according to the first feature of claim 1. As extensive tests have shown, the flow pulsation is therefore not the fluctuation of the throughput per unit of time regardless of the tooth profile selected, which can be particularly easily changed with a cycloid toothing by changing the ratio of the tooth thicknesses of the inner ring gear and pinion to each other, without thereby the advantages the cycloid teeth are lost. The first characteristic feature of claim 1 makes use of this fact. If one records the fluctuation of the instantaneous displacement volume, i.e. the quotient of the difference between the maximum displacement volume and the minimum displacement volume and the mean displacement volume over the ratio of the widths of the ring gear tooth and pinion tooth to one another, there is a minimum in the range between tooth width ratios of 1, 5 and 3 for the non-uniformity of the instantaneous displacement volume.

The training becomes even more favorable if, according to claim 2, the circumferential extent of the pinion gaps and toothed ring teeth is selected to be 1.75 to 2.25 times as large as the circumferential extent of the pinion teeth and toothed tooth gaps.

The conditions become optimal if, according to claim 3, the pinion teeth are selected half as thick as the toothed ring teeth, that is to say that the rolling circle producing the epicycloids is made half as large as the rolling circle producing the zypocycloids.

When the tooth profiles are flattened, only one of the two groups of cycloids, that is to say either the epicycloids or the hypocycloids, is preferably flattened to the full extent of the required play, while the flattening of the other cycloid group is zero. Here again it is preferred that the epicycloids are flattened.

In the case of flattening, it is of course essential that both the flattening of the tooth gaps and the flattening of the tooth heads interacting with these tooth gaps obey the same mathematical law. The flattening can be brought about, for example, by reducing the radial height of the teeth and the radial depth of the gaps of the counter gear interacting with these teeth by a small amount, which is continuous from the center of the tooth or the center of the gap to the intersection of the tooth contour with the pitch circle decreases to zero. However, this represents a deviation from the cycloid profile which is optimal per se. The easiest way to achieve the flattening according to claim 6 is by a slight radial displacement of the point describing the cycloid from the circumference of the rolling circle towards its center. In this way a cycloid contour is maintained.

This creates a small gap of the order of a tiny fraction of a millimeter between the starting point of the flattened cycloid and the corresponding base point of the non-flattened cycloid on the pitch circle. This gap is advantageously bridged according to claim 7 in that the starting point and the end point of the flattened cycloids are connected to the start and end point of the non-flattened cycloids on the pitch circle by a straight line.

Since the flattening of the cycloids is only a minimal correction to reduce the play, which is already kept as small as possible, it is sufficient if, according to claim 8, the sum of the two cycloid displacements (the one displacement, as mentioned above, is also zero) can and is preferably also) measured in the middle of the cycloid is the 2000th to 500th part of the pitch circle diameter of the toothed ring.

For relatively large gear ring diameters, the sum mentioned will be measured with a 1000th, while with small gear ring diameters one will go up to a 500th. It can be seen from this that, for example, with a gear ring pitch circle diameter of 100 mm, the sum of the two cycloid flattenings and thus also the distances of the starting points of the flattened cycloids from the associated pitch circle is only of the order of magnitude of 0.1 mm.

Nevertheless, this flattening ensures that in the area of deepest tooth engagement the two gears can mesh with one another almost without play, while a play of the order of magnitude of a maximum of 0.1 mm between the tooth tips is kept free, which in certain rotational positions the gears can compensate for out-of-roundness of the toothed ring and possibly also the pinion at the point of the smallest diameter of the toothed ring up to zero.

Even if, according to the invention, the tooth play at the point of deepest engagement can be extremely small, it naturally cannot be zero. Here, the required minimal tooth flank play in the circumferential direction can be brought about by an equidistant withdrawal of the tooth contour. The extent of this withdrawal can be, for example, 10 times the diameter of the toothed ring pitch circle. From this number you can see how small the backlash required in the invention.

In gear ring machines, the flow pulsation naturally decreases with an increasing number of teeth; but unfortunately also the flow rate itself. Therefore, the aim is to keep the number of teeth in the gear ring machine as low as possible without having to deal with excessive flow rate pulsation and other disadvantages due to the too low number of teeth. Accordingly, the number of teeth of the pinion is advantageously chosen between 7 and 11.

In order to prevent the influence of abrupt fluctuations in the pressure in the delivery flow of liquid pumps, which can result from the collapse of vapor bubbles caused by gravitation in the liquid delivery flow, a narrow axial groove is advantageously provided in the tooth gap base at least and preferably of the pinion.

According to claim 12, the grooves are advantageously about a quarter to a sixth of the rolling circle circumference, preferably a fifth of the same width.

According to claim 13, the grooves are advantageously 2 to 3 times as wide as deep.

The axial grooves in the base of the pinion tooth gaps ensure a certain dead space without, however, the optimal filling of the tooth gaps by the tooth heads of the toothed ring and thus also the optimal guidance of the gears to one another and thus the perfect seal between the teeth being impaired to a disruptive extent. In the dead space created in this way, gravitation bubbles and squeeze oil filled with vapor of the operating liquid can collect without the bubbles being forced to collapse by the function of the pump or the motor. Since the gravitational bubbles collect due to their low mass under the influence of centrifugal force near the tooth base of the pinion, the negative dead space effect of the grooves provided according to the invention is reduced to a negligible residual minimum.

The guidelines given above for the dimensioning of the grooves assume that grooves that are too narrow have too little absorption capacity, grooves that are too deep impair the strength of the pinion and grooves that are too wide in turn impair the interaction of the gear contours.

If you give the grooves a rectangular profile, this has the advantage of a relatively large capacity; if you give them a strongly rounded profile, such as a circular arc profile, this has the advantage of weakening the pinion strength as little as possible. In the case of rectangular grooves, the edges between the side walls and the base of the grooves are advantageously rounded in order to avoid notching effects. The edges between the side walls of the grooves and the subsequent tooth space base are advantageously angular in order to to maintain the full load-bearing capacity of the tooth space base as far as possible. However, these edges should not be sharp.

According to an advantageous development of the invention, the grooves are also provided in the tooth space base of the ring gear. Here, the grooves cannot absorb gravitational bubbles, but can squeeze oil, which is an advantage in some cases. These grooves can usually be kept smaller than those in the tooth gap base of the pinion.

The grooves can have a circular arc profile, for example, seen in axial section. For manufacturing reasons, however, it is preferred that the grooves run with a constant profile over the entire tooth width.

The groove arrangement described can of course also be used advantageously in gear machines other than those according to claim 1; they are even suitable for gear machines with a filler, in which the difference in the number of teeth is greater than 1.

Fig. 1

zeigt schematisch die Ansicht auf eine Zahnringpumpe nach der Erfindung, wobei der Deckel weggelassen ist, so daß die Zahnradkammer mit den Zahnrädern erkennbar ist.

Fig. 2

zeigt eine vorteilhafte geometrische Konstruktion für die Abflachung der Zykloiden in vergrößertem Maßstab.

Fig. 3a

zeigt die linke Hälfte einer idealen spielfreien Verzahnung gemäß der Erfindung an der Stelle des tiefsten Zahneingriffs in noch stärker vergrößertem Maßstab.

Fig. 3b

zeigt eine ein reales Spiel aufweisende Verzahnung gemäß der Erfindung in gleicher Darstellung wie Fig. 3a.

Fig. 4 und 5

zeigen die Zahnräder der Pumpe gemäß Fig. 1 in verschiedenen Umlaufpositionen.

Fig. 6

zeigt die Abhängigkeit der Ungleichförmigkeit des instantanen Fördervolumens in Abhängigkeit vom Verhältnis der Hohlradzahnbreite zur Ritzelzahnbreite für eine Pumpe mit dem Zähnezahlverhältnis 7 : 8.

The invention is explained in more detail below with reference to the drawings.

Fig. 1

shows schematically the view of a gerotor pump according to the invention, the cover being omitted, so that the gear chamber with the gears can be seen.

Fig. 2

shows an advantageous geometric design for flattening the cycloids on an enlarged scale.

Fig. 3a

shows the left half of an ideal backlash-free toothing according to the invention at the location of the deepest meshing on an even larger scale.

Fig. 3b

shows a real toothing according to the invention in the same representation as Fig. 3a.

4 and 5

show the gears of the pump according to FIG. 1 in different circulating positions.

Fig. 6

shows the dependence of the non-uniformity of the instantaneous delivery volume as a function of the ratio of the ring gear tooth width to the pinion tooth width for a pump with the tooth number ratio 7: 8.

The gerotor pump shown in FIG. 1 has a housing 1 in which a cylindrical gerotor chamber 2 is recessed. On the circumferential surface of the toothed ring chamber 2, the toothed ring 3 is rotatably supported with its cylindrical circumferential surface. The toothed ring 3 has eight teeth 4. These teeth mesh with the teeth 5 of the pinion 6, which is non-rotatably seated on a shaft 7 driving the pinion. The axis of rotation of the ring gear 3 is designated 8; that of pinion 6 with 9. The pump rotates clockwise as indicated by the arrow in FIG. 1. It has a suction opening 10 and an outlet opening 11. The contours of these two openings lie behind the gear wheels in FIG. 1 and are therefore shown in broken lines.

The supply and discharge channels to the suction opening 10 and from the outlet opening 11 are not shown in FIG. 1 for the sake of clarity.

As far as previously explained in the description of the figures, the pump is generally known.

Except for the flattening of the cycloids, the ratio of the pinion tooth width to the internal ring tooth tooth width and the grooves 12 at the bottom of the pinion tooth gaps, the pump shown corresponds to a pump according to German Patent No. 39 38 346 or US Patent Application S.N. 593 135 of October 5, 1990.

Furthermore, the width of the pinion teeth BE, which is measured in radians on the pinion rolling circle TR, and the width BH of the inner ring teeth, measured analogously on the internal ring gear rolling circle TH, are entered in FIG. 4. In Fig. 4, the theoretical line of action E is also entered. The upper part of this line of engagement in FIG. 4 can be seen again enlarged in FIG. 3a. As said, this line of engagement represents the path of the point at which the contours of the pinion teeth and inner ring teeth touch when the gears rotate.

Assuming the position of the gearwheels shown in FIGS. 3a and 5, the point of engagement is initially at the location EO (FIG. 3a). From there, the point of engagement moves along the semicircle E1 to the pitch point C, that is to say the point at which the two pitch circles TH and TR touch on the connecting line of the gear wheel centers 8 and 9. The point of engagement on the circle E3 moves from C in the direction of the arrow. If the point of engagement has reached the apex of this circle on the straight line through EO and C, then the center line of the pinion tooth which can be seen on the left in FIG. 3a is on the straight line EO-C. The point of engagement now moves further along the left half of the circle E3 to point C, at which the left flank of the pinion tooth shown on the left in FIG. 3a is now located. At the same time, the point of engagement between the epicycloid of the pinion tooth head and the hypocycloid of the inner tooth ring moves down along branch E2 between the two pitch circles into the area opposite the point deepest meshing and then back up to point C (see Fig. 4).

In practice, or more precisely, the path of the sealing point between two teeth, however, the line of engagement deviates considerably from this theoretical course of the line of engagement because of the play and the manufacturing inaccuracies.

In Fig. 3a you can also see that in the theoretical ideal case shown there, when the ring gear tooth is with its center line on the center line of the two gears, there is only one between the trailing tooth flank of the ring gear tooth and the driving flank of the pinion tooth in a cycloid toothing extraordinarily thin remaining volume strip VR. This must be displaced to the optimum position during the rotation angle range that follows, until the displacement maximum is reached.

In practice, however, a toothing cannot run completely free of play. In particular, a relatively large amount of play has hitherto been required, because in the area opposite the point of deepest tooth engagement in the sealing area between suction and pressure kidneys, which is necessary there, there must be an adequate tooth play, which is undesirable per se, so that there is no inhibition and no hammering Teeth occurs against each other. This required running play at the lower sealing point in FIG. 1 also leads to an undesirably large game in the known cycloid toothing at the sealing point in the area of the deepest tooth engagement. The invention now makes it possible to actually make do with minimal play at the point of deepest tooth engagement, without therefore increasing the required relatively large tooth play in the area opposite the point of deepest tooth engagement affect. The preferred possibility for producing the flattening of the cycloids forming the tooth gap and tooth contours required for this is exaggerated in FIG. 2. The pitch circle of the gear to be corrected is designated by T there. In the following it is assumed that this is the pitch circle of the pinion.

2 also shows the rolling circle RH. If it rolls from the point Z0 on the pitch circle on the inside of the pitch circle, point Y1 of the circumference of the pitch circle RH, which is initially located at point Z0, describes a cycloid FR, which here limits the tooth gap of the pinion. If you now move the point describing the cycloid a little bit on the radius rH of the rolling circle RH towards the center of the rolling circle RH to position X1, this point X1 is initially in the starting position, at which point Y1 in Z0 is in position Z1. If the rolling circle RH now rolls to the left on the pitch circle T, point X1 also describes a cycloid FR1, the end point of which, however, is at a short distance from the pitch circle. This distance corresponds to the distance Z1-Z0 in FIG. 2. In an analogous manner, the epicycloid FH, which delimits the tooth tip of the pinion, can be flattened by rolling off the rolling circle RE. In this case, the point X2 describing the flattened cycloid FH1 is in the starting position at Z2. In this way, both the large pinion tooth base located on the left was shifted radially outward to the pitch circle T, while the pinion tooth contour was flattened radially away from the cycloid FH toward the pitch circle T.

The teeth and tooth gaps of the inner tooth ring are flattened in the same way. The construction is carried out as just described, except that the pitch circle T is the pitch circle of the inner toothed ring and the rolling circle RH generates the tooth contour and the rolling circle RE generates the tooth gap contour. In the construction according to the invention, the flattened cycloids begin and end at a slight distance from the pitch circle T. In FIG. 2, this is the distance Z1-Z2. This distance can easily be bridged by a straight line, since it is very small compared to the greatly exaggerated illustration in FIG. 2. If you have constructed the teeth as just described, you first get an ideal toothing in the area of deepest tooth engagement, which corresponds to FIG. 3a, but has a tooth play SR compared to the area of deepest tooth engagement, which in the position of FIG. 5 is the sum of Distances Z0-Z1 and Z0-Z2 corresponds. When defining the tooth play in the area of deepest tooth engagement, it is no longer necessary to take into account the out-of-roundness of the inner toothed ring, as long as the sum of the two reductions in the tooth height of the pinion and inner toothed ring is large enough to cause metallic contact of the teeth in the area opposite the point to prevent deepest tooth meshing with certainty. In practice, of course, you will not flatten both the teeth of the pinion and the ring gear, but only one of these two groups of teeth. That is easier. Now you can actually get by with only a minimal tooth play, which is easiest to obtain by either the contour of the inner toothed ring or that of the pinion on an equidistant line lying one or a few hundredths of a millimeter behind the tooth contour FR1, FH1 obtained according to FIG. 2 is withdrawn. 5 shows the gearwheel pairing obtained in this way again. It can be seen here that the circumferential tooth play SU need only be a small fraction of the play SR between the tooth heads in the area opposite the point of the deepest tooth engagement.

3b shows the toothing created by the invention in the same representation as FIG. 3a. It can be seen here that the minimal tooth play caused by the withdrawal of a tooth contour by, for example, a thousandth of the pitch circle diameter is filled by the liquid volume VR. The play or gap thus generated between the two gearwheels in the position shown in FIG. 3b has the effect that the driving force exerted by the driven pinion is not transmitted, as in the theoretical case, at point E0, but rather is distributed over a fairly large area , which arises from the fact that the minimum gap is filled with fluid and this fluid fluid cushion transmits the driving force over a large width. With the previously required large tooth play, the conforming of the two tooth contours to each other was much worse, so that the liquid film was only of a much smaller width and the amount of squeezing oil was significantly larger. The contact between the driving pinion tooth and the driven ring gear tooth takes place in the invention over a large area, since the thickness differences of the thin fluid layer between the two tooth flanks are so small that the pressure required to squeeze the fluid out of the gap in Fig. 3b is sufficient to transmit the torque on the ring gear. The area of engagement E1 shown in FIG. 3a has now been replaced by the area shown in FIG. 3b covered by the bundle of curves E1 '.

Analogously, the same thing that was just described for the interaction of pinion tooth gap and ring gear tooth applies to the interaction of pinion tooth and ring gear tooth gap. Here the line of engagement E3 becomes the engagement surface E3 '.

A force-transmitting tooth contact no longer takes place in the area of the engagement line parts E4 and E5 in FIG. 3a.

This is prevented by the large tooth play in the circumferential area outside the area of deepest tooth engagement. Only the first part of branch E2 remains for a short while.

Fig. 3b finally shows that the inventive design with a minimum gap VR between the teeth in the position shown in Fig. 3 also ensures an excellent seal, since the remaining gap VR is extremely narrow over its entire length.

As can be seen from FIGS. 2 and 4, the circumferential extent of the tooth heads 4 or tooth spaces delimited by hypocycloids FR1 or tooth spaces delimited on the pitch circle T of the respective gear wheel 3, 6 is twice as large as the corresponding extent of the tooth spaces delimited by epicycloids FH1 or heads 5. In other words: the rolling circle RH, which describes the hypocycloid FR1, should have a diameter which is approximately twice as large as the rolling circle RE.

A particular advantage of the invention is that there are practically no radial and tangential accelerations and decelerations between the two gears.

In general, for the radial running clearance, the shortening of the tooth profiles, which is also effective in the area of deepest tooth engagement, to an equidistant to the cycloid or flattened cycloid, which is one or a few hundredths of a millimeter behind, is generally one sixth to one third of the running clearance in the area compared to the point of deepest meshing.

From the above, it can finally be seen that the play reduction according to the invention achieves a particular advantage in the gear machine according to German Patent No. 39 38 346, in which the toothings are mounted one inside the other.

As can be seen in FIG. 3b, the amount of residual squeezing oil in the invention, which when the toothing is rotated further from the position shown in FIG. 3b to a position in which the center line of the pinion tooth on the axis spacing line covers - at least in the case of an oil pump - is not essential more than the thin oil film, which cannot be removed from the surface at all without excessively high pressures. In other words, pinch oil hardly needs to be displaced, since the amount of oil remaining in the gap hardly exceeds the thin oil film that just fills the game.

Fig. 6 zeigt die Verhältnisse bei einem Zähnezahlverhältnis von 7 : 8, wie es in Fig. 1, 4 und 5 gezeigt ist. Fig. 6 zeigt mit der dort erkennbaren Kurve die Abhängigkeit des Ungleichförmigkeitsgrades des instantanen Fördervolumens vom Verhältnis der Zahnbreiten. Dieses Verhältnis hat bei $\text{BH/BE = 2}$

ein ausgeprägtes Minimum. Dort liegt der Ungleichförmigkeitsgrad nur noch bei etwa 2,5 %, während er bei gleich breiten Zähnen mehr als 5 % beträgt. Auf diese Weise trägt das gewählte Zahnbreitenverhältnis gemäß der Erfindung ganz erheblich zu einer Verringerung der Förderstrompulsation bei, die wiederum geräuschmindernd wirkt.This significantly reduces the flow pulsation. The different tooth head width according to the invention described above acts in the same sense. 6, the ratio of the tooth width of the ring gear to the tooth width of the pinion or, expressed mathematically, the ratio of the diameter of the rolling circle producing the hypocycloids to the diameter of the rolling circle producing the epicycloids is plotted on the abscissa. The ordinate shows the degree of non-uniformity of the instantaneous delivery volume A. The degree of non-uniformity is then given by the formula:

FIG. 6 shows the relationships with a tooth number ratio of 7: 8, as shown in FIGS. 1, 4 and 5. Fig. 6 shows with the curve recognizable there the dependence of the degree of non-uniformity of the instantaneous delivery volume on the ratio of the tooth widths. This relationship has $\text{BH / BE = 2}$

a pronounced minimum. There the degree of non-uniformity is only about 2.5%, while it is more than 5% for teeth of the same width. In this way, the selected tooth width ratio according to the invention contributes significantly to a reduction in the flow pulsation, which in turn has a noise-reducing effect.

In order to significantly reduce the noise development even at higher speeds in gear machines according to the invention, which are already characterized by low noise, the axial grooves 16 are provided in the center of the tooth space base of the pinion 6. As can be seen in the drawing, these grooves have a semicircular profile and merge angularly but not with sharp edges into the tooth space surface of the pinion.

If the gear machine now rotates clockwise, the gravitational bubbles created at higher speeds accumulate in the conveying liquid due to the centrifugal force in the grooves 16, where they are transported with only a minimal dead space effect over the point of deepest tooth engagement, i.e. the pitch point C, into the suction area. The grooves can also hold squeeze oil here. As tests have shown, this results in a very considerable reduction in noise and thus a corresponding improvement in efficiency.

Analog grooves can also be provided in the bottom of the tooth space of the internal gear at 17 for receiving pinch oil. These grooves are indicated by dashed lines in FIG. 5.

Claims (13)

1. A gear-type machine (pump or motor for fluids or gases)
   comprising a housing containing a gear chamber (2) having inlet and outlet openings (10, 11);
   and comprising an internally toothed ring gear (3) arranged in the gear chamber (2) and a pinion (6) which is rotatably arranged within the ring gear (3) in the housing (1) and which has one tooth (5) less than the ring gear, meshes with the latter and on rotation forms between its teeth (5) and the teeth (4) of the ring gear (3) revolving, expanding and diminishing fluid cells which conduct fluid from the inlet to the outlet;
   the teeth heads of the pinion (6) and the teeth gaps of the ring gear (3) having the form of epicycloids (FH) which are formed by rolling of a first pitch circle (RE) on the rolling circle (T) of the pinion (6) or ring gear (3) (Fig. 2);
   the teeth gaps of the pinion (6) and the teeth heads of the ring gear (3) having the form of hypocycloids (FR) which are formed by rolling of a second pitch circle (RH) on the rolling circle (T) of the pinion (6) or ring gear (3) respectively (Fig. 2);
   and the radius of the first pitch circle (RE) being different than the radius of the second pitch circle (RH),
   characterized in
   that the peripheral extent (BH), measured on the respective rolling circle (TH, TR), of the pinion teeth gaps defined by hypocycloids (FR1) and ring gear teeth (4) is 1.5 times to 3 times the peripheral extent (BE), measured on the respective rolling circle (TH, TR), of the pinion teeth (5) defined by epicycloids (FH1) and the ring gear teeth gaps; and
   that the epicycloids (FH1) and/or the hypocycloids (FR1) are flattened towards their rolling circles (TH, TR) to such an extent that the flattening or the sum of the two flattenings (Z0 - Z1; Z0 - Z2) corresponds to the necessary relatively large radial clearance (SR) between the teeth heads in the region opposite the point of deepest teeth engagement, whereas the gears (3,6) mesh with each other with a far smaller clearance at the point of deepest teeth engagement.
2. A gear-type machine according to claim 1, characterized in that the peripheral extent (BH) of the pinion teeth gaps and ring gear teeth (4) is 1.75 times to 2.25 times the peripheral extent (BE) of the pinion teeth (5) and ring gear teeth gaps.
3. A gear-type machine according to claim 2, characterized in that the peripheral extent (BH) of the pinion teeth gaps and the ring gear teeth is two times the peripheral extent (BH) of the pinion teeth and ring gear teeth gaps.
4. A gear-type machine according to one of the claims 1 to 3, characterized in that the cycloids of one of the two groups of cycloids (hypocycloids (FR) and epicycloids (FH)) are flattened by the full extent of the necessary clearance whilst the flattening of the cycloids of the other group is equal to zero.
5. A gear-type machine according to claim 4, characterized in that the epicycloids (FH) are flattened.
6. A gear-type machine according to one of the claims 1 to 5, characterized in that the flattening of the cycloids (FR, FH) is effected by a slight radial displacement of the point describing the respective cycloids from the periphery of the pitch circle (RH, RE) in the direction towards the centre thereof (Fig. 2).
7. A gear-type machine according to claim 6, characterized in that the starting point (Z1, Z2) and the end point of each flattened cycloid (FR1, FH1) is connected by a straight line with the starting and end point (Z0) respectively of the original unflattened cycloids (FR, FH) the rolling circle (T).
8. A gear-type machine according to one of the claims 1 to 7, characterized in that the flattening or the sum of the two cycloid flattenings measured in the cycloid centre is the 2000th to 500th part of the diameter of the rolling circle (TH) of the ring gear (3).
9. A gear-type machine according to one of the claims 1 to 8, characterized in that the minimum teeth flank clearance necessary at the point of deepest teeth engagement is effected by an equidistant reduction of the tooth contour.
10. A gear-type machine according to one of the claims 1 to 9, characterized in that the pinion (6) has 7 to 11 teeth.
11. A gear-type machine according to one of the claims 1 to 10 for liquids, characterized in that in the teeth gap bottom at least of the pinion (6) narrow axial grooves (16) are provided.
12. A gear-type machine according to claim 11, characterized in that the grooves (16) are about one quarter to one sixth as wide as the periphery of the pitch circle (RH, RE) generating the teeth gap, preferably one fifth thereof.
13. A gear-type machine according to claim 11 or 12, characterized in that the grooves are 2 to 3 times as wide as they are deep.

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