EP0046779B1 - Eccentric screw pump with helical rotor - Google Patents

Abstract

The eccentric screw pump is subdivided into several stages by separation planes (18). The stator (1) is formed by distinct stator segments (3) which cooperate with the eccentrics of the rotor formed for example by distinct eccentric discs (22) mounted on a grooved shaft (20). Between the stator and the eccentric of each stage there is inserted an intermediary member comprised of an annular shrouding (240) pivotingly mounted on the eccentric, a piston ring or a wear ring fixed angularly and optionally adjustable longitudinally on the eccentric. The adjacent eccentrics (22) are offset angularly by one angular subdivision of the rotor (tr) about the axis of the shaft (15), the adjacent rings (3) of the stator are offset mutually by an angular shift of the stator (ts) the magnitude of which is twice smaller. The intermediary members (240) are guided by separation planes (18) of the stator. To allow the rotor to freely adapt to an axial thermal expansion different from that of the stator, there are provided free expansion spaces which, in the present case, are formed by annular grooves (67), but which may also be formed by crescent shaped and grooves. The eccentrics of adjacent stages may move in the grooves (67) in case of different thermal expansions.

Description

Field of application of the invention

The invention relates to an eccentric screw pump with a helically coiled rotor, which rolls like a planet in a hollow spiral of a stator with an elongated cross section.

State of the art

In eccentric screw pumps of the type mentioned, the hollow spiral of the stator always has one more gear than the rotor spiral. Usually, a single-start rotor is combined with a two-start stator, the rotor having a circular cross-section and the cross-section of the hollow spiral of the stator being elongated and delimited at both ends by semi-circles adapted to the rotor cross-section, which are delimited by two straight lines. The length of this straight line corresponds to four times the value of the eccentricity e with which the rotor rotates on the one hand around the pump axis and on the other hand each cylindrical eccentric part of the rotor around the rotor axis. With a continuous design of the interacting pitch curves on the rotor and stator, an eccentric moves in each radial plane of the pump between the two ends of the stationary, long, round hollow helix cross sections of the stator. As a result, cavities with the length of a rotor pitch on a screw path are successively conveyed from one end of the pump to the other.

Since, in the case of continuously curved curves, the seal between axially successive cavities only extends along circumferential lines and the rolling process is connected with grinding movements on the inner surface of the hollow helix, the seal and thus the achievable delivery pressure is reduced during operation by wear, so that from time to time Stator and / or rotor must be replaced.

From DE-Offenlegungsschrift 2 530 552 it is known to resolve the spatial pitch curves on the rotor and stator into individual pump stages, which are delimited on the end face by separating planes running radially to the pump axis and each of which has an eccentric that leads to the eccentrics of the adjacent pump stages a rotor pitch angle is arranged rotated and carries at least one intermediate member guided in the stator cavity of a pump stage between separating planes, which there has, for example, the shape of a cylindrical rolling ring which is rotatably mounted on the cylindrical outer surface of the eccentric and which rolls in the long round stator cavity of a pump stage . This rolling ring provided with uniform dimensions can also be produced inexpensively in large quantities from wear-resistant and expensive material and can therefore be replaced as required in the event of wear. Likewise, the stator and rotor can be composed of individual stator and rotor disks that are twisted and braced against each other by a stator pitch angle t _s or a rotor pitch angle t _{r that is} twice as large. This simplifies and reduces the cost of manufacturing and warehousing.

According to DE Offenlegungsschrift 2905917, instead of rolling rings, intermediate members in the form of piston plates are provided which move in the elongated stator cavity in the same direction at all times and thereby dissolve the conventional sliding rolling motion in a rotary motion on the eccentric and a sliding motion in the stator cavity. The stroke of the piston plates is dimensioned such that they still form radial clearances to the ends of the stator cavities even in their reversed positions. In this way, solid-state parts are to be deflected into spaces which remain free to the side without necessarily coming into direct contact with the piston plates, let alone having to be smashed by them.

According to US Pat. No. 3,274,944, a rotor with a single-start worm groove runs largely parallel to itself in a housing, a worm helix which is also supported thereon engaging in the worm groove of the rotor. The helical grooves of the rotor slide primarily in the radial direction on the largely identically shaped, multi-part screw webs, which results in constant spring adjustment if the seal is inadequate and the flow rate is small. Has. Elongation problems cannot occur there due to the constant spring adjustment.

This also applies to a worm pump according to US Pat. No. 3,229,643, in which a stationary-mounted rotor shaft carries individual eccentrics provided with axial spacing with rotatable intermediate rings, which are enclosed by a non-rotatably held flexible sleeve. Even the intermediate rings can adjust themselves axially to the extent necessary without any expansion stresses being able to build up.

In contrast to the last-mentioned versions, problems have arisen in progressive cavity step pumps with intermediate members designed as wear rings, for example, in that the intermediate members are supported in each operating direction with the outer shoulder on the stator and with an inner shoulder on the rotor. The frictional forces that occur on these shoulder surfaces and the resulting local wear are largely due to the different thermal expansion of the rotor and stator.

Purpose of the invention

It is an object of the invention to design eccentric screw pumps of the type mentioned at the outset as simply as possible so that friction and wear in the area of the intermediate links are reduced. In particular, the invention aims to prevent an increase in the system forces of parts that can be moved relative to one another by the different thermal expansion on the rotor and stator. Another purpose of the invention is to reduce friction and wear with at least the same or improved sealing in the individual pump stages.

Discussion of the invention

According to the invention, an eccentric screw pump with a helically coiled rotor, which rolls like a planet in a hollow helix of a stator with an elongated cross section, serves to solve the above-mentioned problem, the spatial pitch curves on the rotor and stator being resolved into individual pump stages which are provided on the end face by separating planes running radially to the pump axis are limited and each of which has an eccentric, which is arranged rotated by a rotor pitch angle to the eccentrics of the adjacent pump stages and carries at least one intermediate member guided in the stator cavity of a pump stage between parting planes, free spaces being formed in the area of use of each pump stage Eccentric screw pump is characterized in that the free spaces are designed as expansion free spaces arranged between the rotor and stator, the depth of which in the axial direction corresponds at least to the different thermal expansion of the rotor and stator ht.

The expansion clearances according to the invention ensure that different expansions of the rotor and stator are not transmitted, for example, from the rotor to the stator via the intermediate members. This is particularly important for such eccentric screw pumps, the rotor of which is driven by a relatively long cardan shaft. The common point of reference for the expansion of both parts is the mounting of a drive shaft upstream of the drive shaft. The distance from this bearing to the free end of the rotor is approx. 0.60 to 2.50 m in common pumps. Even if the materials of the stator and rotor are carefully coordinated, it cannot be avoided that the rotor part of a pump stage, i.e. H. the eccentric is shifted to different degrees relative to the associated stator part depending on the operating time and stress. Usually there are only sliding paths in the order of magnitude of 0.15 to 0.30 mm, but the addition of machining tolerances can result in larger deviations. Due to the expansion spaces now proposed, however, the fixed axial integration of the intermediate links between the rotor and the stator is eliminated and a movement is created which, although practically does not influence the sealing, achieves a largely compulsory alignment and guidance of the intermediate links. It is of secondary importance where these intermediate links experience their axial alignment. This is usually done on the stator, but can also be done the other way around on the rotor, or one can provide free space parts with respect to the rotor and stator and thereby achieve a limited floating guidance of the intermediate members.

The expansion shifts between the rotor and stator are of different sizes in the individual pump stages and are determined by the respective distance from the axial support of these parts. However, the depth of the expansion clearances is expediently designed to be the same in all pump stages and adapted to the greatest local expansion displacement of the rotor and stator, as usually occurs at the free rotor end.

The expansion clearances are preferably each formed by at least one recess formed on the end face in the eccentric and / or intermediate member and / or stator.

The recess in the eccentric can be designed as an edge groove and can be offset by the rotor pitch angle t _{r from} the longitudinal center plane of the eccentric. In particular, it has the shape of a crescent-shaped countersink, the inner boundary of which is adapted to the inner surface of the immediately adjacent intermediate member, so that this intermediate member can move into the recess.

According to a proposal of the invention, the intermediate member can be designed as a wear ring that can be coupled to the eccentric in the direction of rotation of the drive, but can be adjusted in length on the wear ring. The fixed connection in the drive direction between the eccentric and the wear ring leads again to grinding rolling, but has the advantage that the wear ring is guided exactly on the eccentric and cannot jam due to the bearing play or otherwise set in an uncontrolled manner. In principle, the connection can be brought about by any suitable coupling means that does not hinder the longitudinal adjustment, for example a longitudinal wedge.

It is not necessary that in such cases the edge groove runs parallel to the end face of the eccentric or extends in a radial plane. Rather, it can also be limited by an inclined surface inclined to the rotor axis, such as a conical surface, which is arranged with an axial expansion distance from a complementarily shaped inclined surface of the intermediate member.

A projection of the other part (for example the eccentric) which engages in it can be assigned to the recess of one part (for example the wear ring) forming the expansion space be. This engagement element can be a special component or a shape of the eccentric. If the wear ring is constantly held completely non-rotatably on the eccentric, its inner surface does not have to be centered on the cylindrical outer surface. The interface between the eccentric and the wear ring can in principle be designed as desired, taking further aspects into account.

According to a further invention proposal, the recess can have a depth that is greater than the thermal expansion to be compensated for, for. B. form a driver clutch with the constantly engaging projection. Elongation space and coupling are then formed by the same means. The easiest way to accomplish this is that the entire cross-sectional area of the eccentric protrudes into the correspondingly shaped recess of the wear ring of an adjacent pump stage.

According to another proposal of the invention, the wear ring is rotatably supported on its eccentric over at least limited angles, and the coupling means connecting the two parts have an effective play of movement when the load changes. This play of movement is then abolished as soon as grinding resistance occurs on the outer surface. Depending on the design of the coupling means, the eccentric and the wear ring can be firmly clamped together, while this tension can be released again when the load is released.

In one embodiment of this type, where separate eccentrics are seated on a splined shaft, the splined shaft is arranged as a coupling means in the interface between the wear ring and the eccentric, which is crescent-shaped in cross section. The meshing in conjunction with the change in cross-section of the wear ring and the crescent shape of the eccentric allows immediate tensioning when the load is on and an interim relaxation after the load has been reduced.

According to another proposal of the invention, the coupling means can have a coupling element which engages frictionally between the eccentric and the wear ring, which can possibly yield if the grinding resistance exceeds a certain value.

Such a coupling element can be loaded by a spring and be guided so that it can be moved radially outwards laterally of the longitudinal center plane in the eccentric. The lateral arrangement then results in different coupling torques depending on the direction of rotation, for example stronger in the drive direction than in the reverse direction.

The coupling means can also have at least one indexing element for intermittently changing the relative rotational position of the shooting ring on the eccentric. As a result, the engagement position changes constantly, which leads to an equalization of the intervention over the entire circumference and prevents the wear-out or shaping of locally limited engagement points. A freewheel clutch arranged between the eccentric and the wear ring has proven to be particularly useful.

It is understood that primarily a sliding engagement between the end faces of the eccentrics fixedly attached to the rotor and the stator must be prevented. This is ensured if the radial dimension of an intermediate member rotatably mounted on the eccentric is greater than the maximum radial projection of an eccentric over an adjacent eccentric. Instead of providing approximately crescent-shaped free spaces in the eccentric, an inner annular groove can be formed in at least one end face of the intermediate members as an expansion free space, the radial dimension of which is at least equal to the maximum radial projection of an eccentric over an adjacent eccentric. Geometrically, this is the simplest and clearest execution.

In order to enable adequate sealing and guidance on the stator, the annular grooves should each be surrounded by a contact surface aligned in the parting plane, the radial dimension of which is greater on the entire circumference than the largest radial projection of an eccentric cam over an adjacent eccentric cam. In this way, the external dimensions of the intermediate links and thus also the pump diameter are enlarged, but at the same time the delivery volume and delivery pressure increase with the operating values remaining unchanged.

According to a further invention proposal, the expansion clearances can also be provided in the area of the parting planes on the stator.

The expansion clearances can also be formed by the axial adjustability of the intermediate members or eccentrics.

According to another proposal of the invention, in each pump stage on the rotor two axially opposed intermediate members are provided axially adjustable with respect to one another, have sealingly engaging, circumferentially alternating projections and grooves on their mutually facing end faces and can be pressed apart in the direction of their axis.

In this way, wear occurring on the end faces of the intermediate links is largely avoided. The seal between adjacent pump stages is kept constant to such an extent that even higher delivery pressures can be achieved even after a long period of operation.

In this way, a uniform contact pressure of the two intermediate elements of each pump stage is initially exerted on the parts of the adjacent pump stages which adjoin the end face, and a sealing effect which is practically independent of the state of wear is achieved in the parting planes. One then only has to ensure that the relative Movability of the two intermediate members of a pump stage in the circumferential direction only limited flow channels are formed, which can not create pressure compensation between peripheral points of the intermediate members which are further apart, or that such channels are closed to the outside. Even if the axial adjustability and increased circumferential clearance create direct connections between the inside and outside of the intermediate links, there is at least one seal in the peripheral direction between adjacent projections of the two intermediate links. The pressure ratios, which change from level to level, result in small torque differences which ensure that they rest on one of the two side surfaces of a projection.

The greater the number of projections and grooves, the greater the number of sealing points and the smaller the circumferential angle of a clearance section, which u. This could cause a leak. Since such margins on the inside are closed off by the eccentric outer surface, the only issue is a flow short circuit between adjacent locations on the outside of the intermediate members. In the case of rolling rings, for example, this is only of importance for brief moments. In practice, these times are less than 0.1 seconds with opening cross sections of less than 0.1 mm ² , so that liquids can hardly react even at very high pressures.

The system pressure required for the sealing in the parting levels usually arises automatically during operation by building up a small overpressure between the two intermediate members. Expediently, however, separate spring elements are inserted between the two intermediate members, which can be mounted, for example, as compression springs, in particular axially symmetrically, in cavities in the projections.

For the use of intermediate members designed as a piston plate, it is recommended that the projections and grooves run only transversely to the longitudinal or sliding direction of the piston plates, since pressure differences can only occur there in the sliding direction. In addition, the projections and grooves designed as ribs should be distributed over their entire width only over the length of the lateral guide surfaces of the piston plates.

Figure lineup

• Figur 1 einen Längsschnitt durch eine erfindungsgemäße Exzenterschneckenpumpe mit Wälzringen, nach der Linie I/I in Fig. 2 - die unteren Flansche längs la - geschnitten,
• Figur 2 einen Schnitt durch diese Pumpe nach der Linie 11/11 in Fig. 1,
• Figur 3 eine Ansicht eines dort eingesetzten Wälzringes,
• Figur 4 einen Schnitt durch diesen Wälzring nach der Linie IV/IV in Fig. 3,
• Figur 5 einen Teil einer Abwicklung zweier ineinandergreifender Wälzringe der Ausführung nach den Fig. 3 und 4,
• Figur 6 eine Abwandlung der Darstellung aus Fig. 5,
• Figur 7 eine Abwicklung zweier Wälzringe mit trapezförmigen Vorsprüngen und Nuten in einer Anlage-Endstellung,
• Figur 8 dieselbe Ausführung mit axialem Einstell-Spiel,
• Figur 9 eine der Fig. 3 entsprechende Ansicht eines Wälzringes mit parallelen Nutscharen,
• Figur 10 eine Stirnansicht einer Exzenterscheibe und
• Figur 11 einen Mittenschnitt dieser Scheibe von links in Fig. 10 gesehen,
• Figur 12 einen der Fig. 1 entsprechenden Längsschnitt durch eine erfindungsgemäße Exzenterschneckenpumpe mit als Kolbenplatten ausgebildeten Zwischengliedern,
• Figur 13 einen Schnitt durch diese Pumpe nach der Linie X/X in Fig. 12,
• Figur 14 eine Stirnansicht einer Kolbenplatte von der Innenseite her gesehen,
• Figur 15 eine Ansicht zweier mit Axialspiel ineinandergreifender Kolbenplatten,
• Figur 16 einen Querschnitt nach der Linie XVI/XVI in Fig. 17 einer derzeit als beste Ausführung angesehenen Exzenterschneckenpumpe mit zentrischen Ringnuten in Wälzringen,
• Figur 17 einen Teilschnitt nach der Linie XVII/ XVII in Fig. 16,
• Figur 18 eine teilweise geschnittene Ansicht eines mit Wälzringen versehenen Rotors von der Linie XVIII/XVIII in Fig. 19 her gesehen,
• Figur 19 einen Teilschnitt nach der Linie XIX/XIX in Fig. 18,
• Figuren 20 bis 23 als Teil-Längsschnitt Abwandlungen des Rotors aus Fig. 19,
• Figur 24 die Stirnansicht einer Rotorstufe, deren Schleißring durch einen Längskeil undrehbar auf der zugehörigen Exzenterscheibe festgelegt ist,
• Figur 25 im Querschnitt die Anordnung der als Vielkeilwelle ausgebildeten Rotorwelle in der Trennlinie zwischen Exzenter und Schleißring,
• Figur 26 eine zwischen Exzenter und Schleißring angebrachte Freilaufkupplung und
• Figur 27 eine im Exzenter angebrachte federbelastete Reibungskupplung zur Verbindung mit dem äußeren Schleißring.

Show in the drawing

• FIG. 1 shows a longitudinal section through an eccentric screw pump according to the invention with rolling rings, according to the line I / I in FIG. 2 — the lower flanges are cut longitudinally,
• FIG. 2 shows a section through this pump along the line 11/11 in FIG. 1,
• FIG. 3 shows a view of a rolling ring used there,
• 4 shows a section through this rolling ring along the line IV / IV in FIG. 3,
• 5 shows a part of a development of two interlocking roller rings of the embodiment according to FIGS. 3 and 4,
• FIG. 6 shows a modification of the representation from FIG. 5,
• 7 shows a development of two rolling rings with trapezoidal projections and grooves in an end position of the system,
• FIG. 8 the same version with axial adjustment play,
• FIG. 9 shows a view corresponding to FIG. 3 of a rolling ring with parallel groove shares,
• Figure 10 is an end view of an eccentric and
• FIG. 11 shows a central section of this disk from the left in FIG. 10,
• FIG. 12 shows a longitudinal section corresponding to FIG. 1 through an eccentric screw pump according to the invention with intermediate members designed as piston plates,
• FIG. 13 shows a section through this pump along the line X / X in FIG. 12,
• FIG. 14 shows an end view of a piston plate from the inside,
• FIG. 15 shows a view of two piston plates which engage with one another with axial play,
• 16 shows a cross section along the line XVI / XVI in FIG. 17 of an eccentric screw pump currently regarded as the best version with central annular grooves in rolling rings,
• FIG. 17 shows a partial section along the line XVII / XVII in FIG. 16,
• FIG. 18 is a partially sectioned view of a rotor provided with rolling rings, seen from line XVIII / XVIII in FIG. 19,
• FIG. 19 shows a partial section along the line XIX / XIX in FIG. 18,
• 20 to 23 as partial longitudinal section of modifications of the rotor from FIG. 19,
• FIG. 24 shows the end view of a rotor stage, the wear ring of which is non-rotatably fixed on the associated eccentric disk by a longitudinal wedge,
• FIG. 25 shows in cross section the arrangement of the rotor shaft designed as a splined shaft in the dividing line between the eccentric and the wear ring,
• Figure 26 shows a one-way clutch fitted between the eccentric and the wear ring
• Figure 27 is a spring-loaded friction clutch mounted in the eccentric for connection to the outer wear ring.

Description of preferred embodiments

In the eccentric screw pump shown in FIGS. 1 and 2, the stator is designated by 1, the rotor by 2. The stator basically consists of 14 identical stator disks 3, which are clamped between flanges 4 and 5 by tie rods 6 by means of nuts 7. The two flanges are integrally formed on connecting pipes 8, 9, which can also be part of certain housing parts. Depending on the direction of rotation of the rotor 2, the connecting pipes guide the suction or pressure flow. According to the drawing, the connecting pipe 8 serves as a drive shaft housing on the suction side, the connecting pipe 9 as a pressure pipe.

The end-side stator disks are aligned in the flanges -4, 5 by centering lugs 10 and sealed by ring seals 11 embedded in the flanges. Each stator disc 3 has, close to its cylindrical outer surface 31, an annular groove 32 with a sealing ring 33 mounted therein for sealing against the adjacent stator disc.

As can be seen above all from FIG. 2, each stator disk has a long round cavity 34 which is central to the pump axis 12 and is delimited by two semi-cylindrical end faces 35 and two flat side faces 36. Since the axes of the two end faces 35, of which the upper axis coincides with the eccentric axis 13 in FIG. 2, are at a distance 2e from the pump axis 12, a length 1 of the cavity 34 with 1 = w + results for a clear width w 4e. Here e is the eccentricity that is decisive for the rotor and its eccentric, which will be discussed later.

Fig. 2 also shows that adjacent stator disks are rotated relative to each other by a pitch angle ts. In order to secure this rotational position, two cylindrical fitting bores 37 are provided in each stator disk at the same pitch angle ts, diametrically opposed to each other. Each of these fitting bores is covered with one of the stator disks adjoining the right and left and is covered by a dowel pin 14 which extends through two stator disks. Dowel holes 37 and dowel pins 14 are used to center each stator disk on the neighboring disks and to secure the exact pitch angle ts.

The stator disks 3 have the width b and lie against one another in the parting planes 18. With a pitch angle ts of 30 °, an entire slope ss of the step helix formed by the individual cavities 34 requires the width of twelve stator plates. Two additional stator plates are provided for additional sealing.

The stator plates expediently consist of hard, dimensionally stable material, for example steel or sintered material. However, they can also have a lining made of another material on the cavity wall, for example made of a rubber-like compliant material or made of a particularly abrasion-resistant material, such as hard metal powder with a ceramic or sintered bond.

The main component of the rotor 2 is the eccentric shaft 20, which is designed as a multi-spline shaft and has individual wedge ribs 21 offset from one another at the pitch angle unit ir = 30 ° and carries an eccentric disk 22 in each pump stage. These eccentric disks 22 are clamped between two end disks 23 by means of a pressure disk 24 by means of a nut 25 against the shaft head 26, which is connected by an articulated coupling 27 to an articulated shaft 28 which transmits the drive torque for the rotor. The rotor can also be driven in other known ways.

The eccentric disks 22, which are again identical and have the same width b of the stator disks or individual pump stages, have an outer cylindrical eccentric surface 221 and the same polygonal profile on the inside as the eccentric shaft 20. However, in the opposite sense to the stator disks, they are at a pitch angle tr rotated against each other, which at 60 ° is twice the pitch angle unit ir and the pitch angle ts of the stator disks 3. In any case, the angle adjustment can be made reliably by the multi-spline profile.

On each eccentric surface 221 there are rotatably as intermediate elements two identical, but oppositely arranged, rolling rings 40 which, in addition to their cylindrical inner surface, have a cylindrical outer surface 41 adapted to the width w and a flat end surface 42 which is directed outwards and lies in a parting plane 18. On the mutually facing sides, however, end teeth 43 made of alternating identical segment-shaped projections 44 and grooves 45 are attached over the entire thickness of the rolling rings. A projection 44 therefore engages in a practically identical groove 45 of the associated rolling ring.

In addition, cylindrical recesses 46 for accommodating helical compression springs 47 are provided on the toothing side in six projections which are symmetrical with respect to the ring axis, which press the two rolling rings apart and bring their end faces 42 into contact with the associated surfaces in the parting planes 18 between adjacent steps. This contact force remains practically unchanged if wear occurs on the end faces 42 or changes in the step width b on the rotor and stator result from different thermal expansion. The end faces are always ground in a little better, resulting in a reliable seal between the adjacent pump stages.

The trapezoidal or wedge-shaped projections or teeth in the front view according to FIG. 3 have a rectangular cross-section in the development according to FIG. 5. With a correspondingly small circumferential clearance, sealing is carried out on both side surfaces 48. The clearance 51 formed by the compression spring 47 between the head surface 49 of the projection 44 and the groove 45 is closed inwards by the eccentric surface 221. It extends over a circumferential angle of only 10 °, so it cannot bring about pressure equalization between adjacent locations of the rotor and stator in the circumferential direction.

In Fig. 6, the top edges of the Protrusions 44 are provided with roundings 52. The base corners 53 of the grooves 45 can also be rounded. This only slightly shortens the width of the side surfaces 48, but the seal itself is largely retained.

7, the projections 44a and the grooves 45a have a trapezoidal cross section. There, therefore, the sealing according to FIGS. 5 and 6 can take place on two side surfaces only in the drawn end position. However, if the two rolling rings 40 are axially pushed apart somewhat by the springs, then a zigzag-shaped channel could extend over the entire circumference along the entire contour of projections and grooves. In practice, however, depending on the pressure curve, one of the two rolling rings is subjected to a greater driving torque than the other. Therefore, one of two side surfaces 48a always comes to rest on the corresponding side surface of the associated projection of the other rolling ring, and only Z-shaped free spaces are formed over an angle of approximately 15 °. A limited pressure equalization could therefore only take place in this area. The trapezoidal shape should therefore only be selected with a very flat angle of inclination in order to maintain the seal on both side surfaces 48a during the adjustment, which experience has shown to extend over less than 1 mm.

When the rotor is rotated, it rolls in the helix of the stator in such a way that each pair of rolling rings 40 rotates on the eccentric surface 221 of the associated eccentric disc 22. Sliding movements on the wall of the screw spiral or the individual stator cavities 34 are avoided in this way. The shaft axis 15 of the eccentric shaft runs on the circular path 16 around the pump axis 12, while the eccentric axis 13 of each pump stage rotates again about the shaft axis 15.

In the end position shown in FIG. 2, the eccentric axis 13 coincides with one of the axes (13) of the end surfaces 35, and the upper displacement is completely emptied, the lower one completely filled. In all intermediate positions, however, the rolling rings lie tangentially in the region of the flat side surfaces 36 on the wall of the cavity 34. In principle, it is about a line seal. However, there are no significant pressure differences between the two opposite rooms, since two conveying paths run along the outer ends of the screw spiral or along the end faces. Therefore, a slightly reduced circumferential seal according to FIG. 8 cannot result in a significant reduction in performance.

On each conveying path of the screw spiral, however, bubble-like and largely closed cavities are conveyed helically from the input end on the right in FIG. 1 to the output end. Usually only a single pump stage seals between adjacent “bubbles”. This sealing is effected on the one hand by the contact of the rolling rings on an end surface 35, which can be seen in FIG. 2, on the other hand on the sealing guidance of the two rolling rings in the separating surfaces 18 on stator disks, rolling rings and eccentric disks of the adjacent pump stages.

FIG. 9 shows a rolling ring 140, in the end face of which groups of grooves 141 to 143 are machined by means of a set of disk cutters 64 mounted on a shaft. All side milling cutters and grooves have the same spacing Z, which results in projections 144 of the same width. In contrast, the side milling cutter shown at the top left in FIG. 9 and the corresponding groove 143 have only half the width of the other two side milling cutters or grooves 141, 142. There are three groove groups milled continuously at a double share pitch angle 2tn = 120 °, so that an effective share pitch angle tn = 60 ° results. Two grooves 143 each intersect and form an X-groove 147, while wedge-shaped projections 145 are formed opposite each other by 180 °, which, as indicated by dash-dot lines 146, fit into a corresponding X-groove 147 of the associated rolling ring. Small grooves with a triangular cross section remain, but these are only open to the outside and cannot create any noticeable pressure compensation due to their small spacing. Here, too, cylindrical recesses 46 for corresponding helical compression springs are provided in six projections 144.

2 shows that the maximum distance between the contours of two adjacent cavities 34 of the stator that are rotated by the pitch angle ts is greater than the radial dimension of a rolling ring 40 or 140. Each eccentric disc 22 engages in the area of a parting plane by the remaining excess 18 beyond the edge of the cavity 34 of the adjacent stator disk. Since the lateral end faces of the stator and rotor disks are practically never in one plane, there is increased wear with a depth of wear that is equal to the displacement that actually occurs in the respective parting plane, as is primarily caused by different thermal expansion.

In order to counteract this, according to FIGS. 10 and 11, a crescent-shaped edge groove 65 is formed on both sides of the eccentric discs 22 and forms an expansion clearance, which extends at least over the overlap area. Both edge grooves are offset by the rotor pitch angle tr = 60 ⁰ to the axial center plane 66 of the eccentric disc.

The relative longitudinal displacement and thus the necessary groove depth nt can change from step to step and is dependent, among other things, on the location of the axial alignment of the eccentric shaft 20 in the stator, in particular the available starting length for the expansion and the adjustment play occurring from the support point. This value must therefore be determined anew for each pump design, but can be adjusted uniformly to the largest displacement occurring in all pump stages, since the peripheral grooves are closed on all sides and cannot bring about any undesired pressure compensation within the pump.

The eccentric screw pump shown in FIGS. 12 and 13 differs from the one described above with regard to the stator 1 only in that the stator disks 3a are provided with a stator cavities 34a approximated to a rectangle, in each of which two piston plates 55 are longitudinally displaceable as shown in FIG. 14 and 15 are guided. These piston plates, which are more likely to be attributed to the stator, sit rotatably on the eccentric surfaces 221 of the eccentric discs 22 instead of the rolling rings. The actual rotor is therefore identical to the embodiment in FIG. 1.

The two piston plates 55 are again identical and are inserted into one another rotated by 180 °. They have perpendicular to the lateral guide surfaces 56 two rectangular grooves 57 and an angular groove 58 which extends to a plate end. Between the rectangular grooves 57 and the corresponding projections 59, which are also rectangular in cross-section, free spaces 60 are formed during the transverse adjustment between the intermeshing rolling rings 40, which, however, are located in the lateral guide surfaces 56 and thus on the corresponding side surfaces 36a of the stator cavity 34a in the middle of a pump stage end and are thereby completed. However, the free space 61 occurring in the area of the angular groove 58 is open outside the guide surfaces 56 up to the plate end 62. However, this free space 61 only slightly increases the respective step space and has the effect that this step space cannot be reduced to zero. In all the projections 59 and in the thickened plate part opposite or opposite the angular groove 58, cylindrical recesses 46 with compression springs 47 are again provided.

The plate ends 62 have the same partial cylindrical shape of the cavity end faces 35a. Its radius is approximately 60% larger than in the case of a cylinder surface which is located centrally to the eccentric axis 13. The minimum length 1a of the cavity 34a results from

The stroke h = 4e also corresponds to the adjustment movement of the rolling rings in the stator cavity according to FIG. 1. With the same width w, the same amount is therefore also displaced. However, by dissolving the rolling movement into a purely linear sliding movement and a rotary movement, the wear is further reduced, and the seal between opposite sides of the intermediate members is improved, which enables higher pressures. In addition, a much larger flow rate can be displaced with the same eccentric dimensions.

Instead of positively moving the piston plates, as shown in FIG. 13, to one end of the stator cavity 34a, one can leave a slight distance there, attach damping means to the stator or piston plates, or, for example, manufacture the piston plates themselves from resilient material.

The roller rings 240 of the eccentric screw pump shown in FIGS. 16 and 17 have a width corresponding to the full width b of each pump stage and a radial thickness which is more than three times greater than that of the aforementioned roller rings with respect to the eccentricity e.

In this way, it is first reliably avoided that the eccentric discs 22 can come into contact with the stator discs 3. Axial thermal expansions can therefore only occur to overlap between the rotating eccentric disks and the rolling rings 240 that are rotatable relative to them and between them and the stator disks 3. Due to the relative rotatability of the rolling rings in both engagement areas, wear is reduced, but is still a nuisance. In order to avoid this forced friction as well, annular grooves 67 are formed on the inside of an end face of the rolling rings as free spaces, the groove depth 2nt of which is shown as exaggeratedly large as in FIG. 11 and the figures described below.

The radial dimension rr of the annular groove is at least equal to the radial projection rü (FIG. 16) of each eccentric over an adjacent eccentric. The annular groove 67 is also surrounded by a flat contact surface 70, the radial dimension ra of which is greater than the aforementioned projection rü of each eccentric over an adjacent eccentric. In this way it is ensured that the outer edge of the annular groove 67 always comes to lie within the outer edge of each adjacent rolling ring and that there is still sufficient radial sealing distance. The bending stiffness, dimensional stability and guiding accuracy of the rolling ring is improved in this way, as is the operational reliability of the pump.

The annular groove 67 is here in only one end face of the rolling ring. This one-sided arrangement is fundamentally also possible with sickle grooves, for example according to FIG. 10. It is of less importance whether the displacement shift always occurs in a certain direction, according to FIG. 17 a shift of the rotor from left to right. If the groove depth 2nt is sufficient, the rotor can be shifted from the start with respect to the stator, i. H. the end faces of the rotor are already displaced in the direction of the free spaces in relation to the parting planes 18.

The use of circular ring grooves 67 still has crescent-shaped ring grooves the advantage that a friction attack that cannot be completely ruled out is distributed over the entire surface and does not result in significant wear even in this case. The production of the rolling rings is easier and the volume conveyed increases in proportion to the outside diameter.

18 and 19, the rolling rings 340 guided in the parting planes 18 of the stator, which is otherwise not shown, are smooth on all sides with a rectangular cross-section, while the free spaces are again designed as crescent-shaped edge grooves 651 and are attached to the eccentric discs 222 on both sides. This also prevents the end faces of the rolling rings 340 from being pulled along the adjacent eccentric discs. The reduction in the radial dimension of the edge groove 651, which corresponds to the radial overhang of adjacent eccentrics relative to one another, is due to an increase in the number of stages per rotor revolution corresponding to a reduced rotor pitch angle tr. With the same step width b, the pump is correspondingly longer.

If the pitch angle tr according to FIG. 18 is designed so that it corresponds to a rotor pitch 8, 5, then the same opening width is obtained as with a seventeen pitch, a rotor pitch 9, 5 results in the same opening width as a nine-pitch pitch, and a 6.5 pitch the same as a thirteen division and thus a significant improvement in the sealing than when choosing a nine division.

The stator pitch angle is half as large and therefore always corresponds to twice the pitch of the rotor, that is 17, 19 or 13 and is only determined by the distance between the two bores 37 (FIG. 2). Even in the case of the eccentric shaft, a division 6, 5 for the V-ribs would in any case not result in any problems if a division angle unit ir of half the size of the rotor division angle tr is always used. The division of the V-ribs 21 on the eccentric shaft 22 then corresponds exactly to the stator division.

According to FIG. 20, eccentric disks 223 are assigned to the rolling rings adopted unchanged, which are provided with a one-sided crescent-shaped edge groove 652, the groove depth nt _{1 of which is} many times greater than the groove depth 2nt required for expansion compensation. Here, the eccentric discs 223 can always be arranged relatively far from their alignment position to the parting planes 18 and thereby enable long displacement paths to the friction system.

According to FIG. 21, it is assumed that wear rings 68 which are non-rotatable as intermediate members are seated on eccentric disks 22 and are adjustable in length. Correspondingly shaped sickle grooves 653 with the outer diameter of the eccentric discs are drawn into the wear rings 68 on the inside on both sides. Due to this fixed arrangement, an improved fit on the eccentric discs is achieved during the engagement process.

22 again shows an embodiment with wear rings 681 non-rotatably mounted on eccentric disks 224. The required expansion clearances are distributed over eccentric disks and wear rings. Each free space is formed there by a first crescent-shaped edge groove 654 in the eccentric disk and a second crescent-shaped edge groove 655 in the wear ring. The boundary surfaces of these edge grooves are also inclined and are formed by cone-shaped or cone-shaped surfaces.

According to FIG. 23, edge grooves 656 in circumferential shapes comparable to those in FIG. 21 are provided in the wear rings 682. The depth of these grooves again largely corresponds to dimension nt ₁ , ie the eccentric disks can be substantially offset from the parting planes 18 of the stator and engage far into the edge dimensions 656, in order to thereby form a positive driving coupling.

24, a cylindrical pin serves as a longitudinal wedge 69 between the wear ring 68 and the eccentric disc 22. The longitudinal wedge couples in the circumferential direction and also provides a longitudinal guide for the relative axial adjustment of the wear ring on the eccentric disc.

25, the eccentric surface 221 of the eccentric 225 is reduced to such an extent that the eccentric shaft 20 penetrates the eccentric surface 221 and also engages with its ribs in a toothed segment 71 of the wear ring 683. The eccentric disc then has a crescent-shaped cross section, and there is also a limited circumferential play between the eccentric shaft and the wear ring, which results in a small rotational displacement after the load has been removed by springback. However, all three parts are always firmly wedged together in the same position under load.

The eccentric disc 226 according to FIG. 26 has a flattened portion 72 in the thickened part of the eccentric, between which and the inner surface of a rolling ring 340 a cylindrical pin 73 is guided such that it forms a one-way clutch between the eccentric disc and the rolling ring which is effective in both directions. The rolling ring, which is rotatably guided on the eccentric surface 221, therefore becomes a grinding and rolling wear ring with each rotation. However, a relative displacement in the circumferential direction is always achieved during the breaks, so that the eccentric surface 221 is approximately uniformly stressed over the entire circumference.

The embodiment shown in FIG. 27 also aims at a gradual shifting of the coupling position between rolling ring 340 and eccentric disc 22. There, parallel and at a lateral distance a to the central plane 66 of the eccentric disc, a radially outward bore with a brake plug is made in the thickened part thereof is loaded by a compression spring 75. In this way, a braking force is constantly exerted on the rolling ring 340, but this braking force is speed-dependent and in the intended manner Driving direction clockwise according to FIG. 27 stronger than in the opposite direction. In any case, the braking force must be such that the rolling ring can only rotate in the drive direction in the event of an overload and / or is rotated over a small distance if the drive torque is removed in the meantime.

In principle, expansion clearances can also be provided in the support areas of the stator, for example the stator plates 3 to the side of the parting planes to adjacent stator cavities 34. In some cases, a limited compensation is created even by providing recesses in the circumference of a rolling ring or wear ring, for example in such a way that axially parallel grooves lead to the opposite end face of the intermediate member from a lateral annular collar which provides the seal. This has the further advantage that the comminution necessary for the enforcement of this good is improved when conveying granular goods.

Claims (24)

1. Eccentric screw pump with a helical rotor (2), which rolls on a planetary path in a cross- sectionally elongated hollow cavity of a stator (1), in which the spatial contact curves on the rotor and stator are broken down into individual pump stages which are limited on their front face by separation planes (18) running radial to the pump axis (12) and of which each shows an eccentric (222) which is arranged angularly off-set by one angular sub-division (tr) of the rotor relative to the eccentrics of the adjacent pump stages and which carries at least one intermediary member (40, 140, 55, 240, 340, 68, 681, 682) guided in the stator cavity (34) of a pump stage between separation planes (18), free spaces being formed in the contact zone of each pump stage, characterized in that the free spaces are formed as free expansion spaces (51, 65, 67, 651-656) arranged between rotor (2) and stator (1), whose depth (nt) in an axial direction corresponds at least to the differential thermal expansion between rotor (2) and stator (1) effective in the individual pump stages.

2. Eccentric screw pump according to claim 1, characterised in that the depth (nt, 2nt) of the free expansion spaces (51 ... 656) is the same in all the pump stages and is adapted to the greatest local expansion difference of rotor (2) and stator (1).

3. Eccentric screw pump according to claim 1, characterised in that the free expansion spaces are formed by at least one recess formed at the front in the eccentric (22) and/or intermediary member and/or stator (1).

4. Eccentric screw pump according to claim 3, characterised in that the recess disposed in the eccentric (22) is formed as a peripheral groove (65) and is arranged displaced by the angular sub-division of the rotor (tr) relative to the longitudinal median plane (66) of the eccentric (22).

5. Eccentric screw pump according to claim 4, characterised in that the peripheral groove 65 of the eccentric has the form of a crescent-shaped cut-away (counter-sinking) whose inner limit is matched to the inner surfacen of the immediately adjacent intermediary member.

6. Eccentric screw pump according to claim 4, characterised in that the intermediary member is formed as a wear ring (68, 681, 682) which is at least couplable with the eccentric in the drive direction but which is longitudinally adjustable on the latter.

7. Eccentric screw pump according to claim 6, characterised in that the peripheral groove (654) is limited by an inclined surface such as a conical surface inclined to the rotor axis, which inclined surface is arranged with axial expansion distance from a complementary conical surface (655) of the wear ring (681).

8. Eccentric screw pump according to claim 6, characterised in that the recess (655) forming the free expansion space of the one part (e. g. wear ring 682) is arranged with a projection (656) of the other part (e. g. eccentric 22) projecting into it.

9. Eccentric screw pump according to claim 8, characterised in that the recess (656) has a depth (nt1) which is greater than the thermal expansion to be compensated and forms a take-up coupling with the projection which continuously projects into it.

10. Eccentric screw pump according to claim 9, characterised in that the eccentric (22) projects with its whole cross-sectional surface as the projection into the recess (656) of the wear ring (682) of an adjacent pump stage.

11. Eccentric screw pump according to claim 6, characterised in that the wear ring (683, 340) is rotationally adjustable at least over limited angles over its eccentric (22, 225, 226) and the coupling means linking the two parts show a « p)ay" of movement effective when there is a change of load.

12. Eccentric screw pump according to claim 11, with eccentrics disposed on a spline shaft, characterised in that the spline shaft is arranged as coupling means in the boundary surface (221) between the wear ring (683) and the cross-sectionally crescent-shaped eccentric (225).

13. Eccentric screw pump according to claim 11, characterised in that the coupling means show a coupling element (74) engaging by means of friction between eccentric (22) and wear ring (340).

14. Eccentric screw pump according to claim 13, characterised in that the coupling element (74) is loaded by a spring (75) and, to the side of the longitudinal median plane (66), is movable radially outwards in the eccentric (22).

15. Eccentric screw pump according to claim 11, characterised in that the coupling means show at least one stepping element for intermittently changing the relative rotational position of the wear ring on the eccentric.

16. Eccentric screw pump according to claim 15, characterised by a slip coupling arranged between eccentric (22) and wear ring (340).

17. Eccentric screw pump according to claim 1, whose intermediary members are rotationally mounted on their eccentric, characterised in that in at least one front surface of the intermediary members (240) an internal annular groove (67) is formed as the free expansion space, whose radial dimension (rr) is at least equal to the maximum radial projection of one eccentric (22) beyond an adjacent eccentric.

18. Eccentric screw pump according to claim 17, characterised in that the annular grooves (67) are each surrounded by a contact surface (70) aligned with the separation plane, whose radial dimension (ra) in total is greater than the greatest radial projection (rü) of one eccentric (22) beyond an adjacent eccentric.

19. Eccentric screw pump according to claim 1, characterised in that the free expansion spaces are located in the region of the separation planes (18) in the stator (1).

20. Eccentric screw pump according to claim 1, characterised in that in each pump stage two axially opposed intermediary members (40, 55) are arranged on the rotor (2) so as to be axially adjustable with respect to each other, and in that on their mutually facing front faces they show sealing, interprojecting projections (44, 59), which alternate in the peripheral direction, and grooves (45, 47), and they (the intermediary members) are urged apart in the direction of their axis (13).

21. Eccentric screw pump according to claim 20, characterised in that separate spring elements (47) are inserted between the two intermediary members (40, 55).

22. Eccentric screw pump according to claim 21, characterised in that the spring elements (47) are located as pressure springs, in particular distributed symmetrically about the axis, in recesses (46) of the projections.

23. Eccentric screw pump according to claim 20, with piston plates (55), displaceable in the stator cavity, as intermediary members, characterised in that the projections (59) and grooves (57) are exclusively transverse to the longitudinal or sliding direction of the piston plates (55).

24. Eccentric screw pump according to claim 23, characterised in that the projections (59) and grooves (57) formed as ribs are disposed exclusively over the length of the lateral guide surfaces (56) of the piston plates (55), distributed over their whole width.

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