EP0046779A1 - Exzenterschneckenpumpe mit einem schraubenförmig gewendelten rotor. - Google Patents
Exzenterschneckenpumpe mit einem schraubenförmig gewendelten rotor.Info
- Publication number
- EP0046779A1 EP0046779A1 EP81900526A EP81900526A EP0046779A1 EP 0046779 A1 EP0046779 A1 EP 0046779A1 EP 81900526 A EP81900526 A EP 81900526A EP 81900526 A EP81900526 A EP 81900526A EP 0046779 A1 EP0046779 A1 EP 0046779A1
- Authority
- EP
- European Patent Office
- Prior art keywords
- eccentric
- stator
- screw pump
- rotor
- pump according
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Granted
Links
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
- F04C2/00—Rotary-piston machines or pumps
- F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
- F04C2/10—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member
- F04C2/107—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member with helical teeth
Definitions
- 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.
- the hollow spiral of the stator always has one more gear than the rotor spiral.
- 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 helix of the stator being elongated and bounded 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.
- stator and rotor can be composed of stator and rotor disks rotated and braced against one another by individual stator pitch angles t s or a rotor pitch angle t r that is twice as large. This simplifies and reduces the cost of manufacturing and warehousing.
- intermediate members in the form of piston plates are provided instead of rolling rings, which are located in the elongated 'stator cavity always move in the same direction and thereby dissolve the conventional grinding rolling movement into a rotary movement on the eccentric and a sliding movement in the stator cavity.
- 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 at the rotor and stator.
- Another purpose of the invention is to achieve friction and wear reduction with at least the same or improved sealing in the individual pump stages.
- 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 broken down 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 that is too the eccentrics of the adjacent pump stages is arranged rotated by a rotor pitch angle and carries at least one intermediate member guided in the stator cavity of a pump stage between separating planes, which eccentric screw pump is characterized in that strain-free spaces are arranged between the rotor and stator in the individual pump stages, the Depth corresponds at least to the different thermal expansion of the rotor and stator.
- 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 eccentric screw pumps whose rotor is driven by a relatively long cardan shaft.
- the common reference point for the expansion of both parts is the mounting of one upstream of the cardan shaft. Drive shaft.
- the distance from this bearing to the free end of the rotor is approx. 0.60 to 2.50m with the common pumps.
- 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.
- 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 end of the rotor.
- 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 made 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.
- 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.
- the intermediate member can be designed as a coupling ring which is at least coupling-friendly but adjustable in length on this wear ring in the direction of rotation with the eccentric.
- the fixed connection in the drive direction between the eccentric and the wear ring leads again to grinding rolling, but it has the advantage that the wear ring is guided exactly on the eccentric and cannot tilt due to the bearing play or otherwise adjust in an uncontrolled manner.
- connection can be brought about by any suitable coupling means which does not hinder the longitudinal adjustment, for example a longitudinal wedge.
- edge groove runs parallel to the end face of the eccentric or extends in a radial plane. Rather, it can be limited by an inclined surface inclined to the rotor axis, such as a conical surface, which stood with axial expansion from a complementarily shaped inclined surface of the intermediate member.
- the recess of one part (e.g. the wear ring) forming the expansion clearance can be assigned a projection of the other part (e.g. the eccentric) which engages in it.
- 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 need not 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.
- the recess can have a depth which is greater than the thermal expansion to be compensated, e.g. form a drive coupling with the protrusion that constantly engages in them. 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.
- 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.
- the eccentric and the wear ring can be firmly clamped together, while this tension is released in the meantime when the load is released.
- the splined shaft has a separate eccentric
- 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 instant tensioning and an interim relaxation after reducing the load.
- 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 can be guided so that it can move 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 wear ring on the eccentric.
- the engagement position changes constantly, which leads to an equalization of the intervention over the entire circumference and prevents the "wearing 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.
- 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.
- the annular grooves should each be surrounded by a contact surface aligned in the parting plane, the radial dimension of which is larger 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 otherwise unchanged operating values.
- the expansion spaces can also be provided on the stator in the area of the parting planes.
- the expansion spaces can also be formed by the axial adjustability of the intermediate members or eccentrics.
- the two intermediate members are provided so as to be axially adjustable with respect to one another, have sealingly engaging, circumferentially alternating projections and on their mutually facing end faces Grooves open, and can be pushed apart in the direction of their axis. In this way, wear occurring on the end faces of the intermediate members 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.
- the projections and grooves run exclusively transversely to the longitudinal or sliding direction of the piston plates, since pressure differences can only occur there in the sliding direction.
- the projections and grooves designed as ribs should also be distributed over their entire width over the length of the lateral guide surfaces of the piston plates.
- FIG. 1 shows a longitudinal section through an eccentric screw pump according to the invention with rolling rings, along the line I / I in FIG. 2 - the lower flanges along Ia - cut,
- FIG. 5 shows 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 illustration from FIG. 5,
- FIG. 7 a development with two rolling rings trapezoidal projections and grooves in an end-of-line position
- FIG. 8 the same embodiment with an axial adjusting play
- FIG. 9 a view corresponding to FIG. 3 of a rolling ring with parallel groove shares
- FIG. 10 an end view of an eccentric disk
- FIG. 11 a Center section of this disk seen 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 interlocking with axial play
- FIG. 16 shows a cross section according to FIG 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 a partially sectioned view of a rotor provided with rolling rings 19 seen from line XVIII / XVIII in FIG. 19, a partial section along the line XIX / XIX in FIG. 18, FIGS. 20 to 23 as partial longitudinal section modifications of the rotor from FIG. 19,
- Fig. 24 is an end view of a rotor stage, the wear ring by a longitudinal wedge un is rotatably attached to the associated eccentric, Fig. 25 in cross section the arrangement of the
- Multi-spline shaft-formed rotor shaft in the dividing line between the eccentric and the wear ring
- FIG. 26 shows a one-way clutch fitted between the eccentric and the wear ring
- FIG. 27 shows a spring-loaded friction clutch mounted in the eccentric for connection to the outer wear ring.
- 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.
- the connecting pipes lead the suction or pressure flow.
- the connecting pipe 8 serves as a suction-side drive shaft housing, the connecting pipe 9 as a pressure pipe.
- stator disks are aligned in the flanges 4, 5 by centering projections 10 and sealed by ring seals 11 embedded in the flanges.
- Each stator disk 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 disk.
- each stator disk has a long round cavity 34 which is central to the pump axis 12 and which is defined by two semi-cylindrical end faces 35 and two flat side surfaces 36 is limited. 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 results for a clear width w
- Fig. 2 also shows that adjacent stator disks are rotated relative to each other by a pitch angle ts.
- 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 extending through two stator disks.
- Dowel holes 37 and dowel pins 14 serve to center each stator disk on the adjacent 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 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 a different 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.
- 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.
- Eccentric surface 22T and inside the same polygonal profile as the eccentric shaft 20 are rotated relative to each other by a pitch angle tr, 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-wedge profile.
- 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.
- 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.
- the trapezoidal or wedge-shaped projections or teeth in the end view according to FIG. 3 have a rectangular cross section in the development according to FIG. 5. With a correspondingly small circumferential clearance on both side surfaces
- Groove 45 formed free space 51 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.
- the head-side edges of the projections 44 are provided with rounded portions 52.
- the base corners 53 of the grooves 45 can also be rounded. This only slightly shortens the width of the side faces 48, but the seal itself is largely retained.
- the projections 44a and the grooves 45a have a trapezoidal cross section.
- the seal according to FIGS. 5 and 6 can therefore take place on two side surfaces only in the drawn end position.
- the two rolling rings 40 are axially pushed apart somewhat by the springs, then it could be along the whole Contour of projections and grooves a zigzag channel extend over the entire circumference.
- one of the two rolling rings is subject 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.
- 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.
- the eccentric axis 13 coincides with one of the axes (13) of the end faces 35, and the upper displacement is completely emptied, the lower one is completely filled.
- the rolling rings lie tangentially in the region of the flat side surfaces 36 on the wall of the cavity 34. In principle, this is about a line seal.
- each eccentric disc 22 extends in the area of a parting plane 18 beyond the edge of the cavity 34 of the adjacent stator disc. 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.
- the relative longitudinal displacement and thus the necessary groove depth nt can change from level to level and is, among other things, depending 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 sealed on all sides and cannot bring about any undesirable pressure compensation within the pump.
- the eccentric screw pump shown in FIGS. 12 and 13 differs from the one described above with respect to the stator 1 only in that the stator disks 3a have stator cavities 34a approximated by a rectangle are provided, in which two piston plates 55 are each longitudinally displaceable as shown in FIGS. 14 and 15. These piston plates, which are probably more attributable to the stator, are rotatably seated on the eccentric surfaces 221 of the eccentric disks 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 of identical design and are inserted into one another rotated by 180 ° to one another. They have two rectangular grooves 57 running perpendicular to the lateral guide surfaces 56 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 of the guide surfaces 56 to the plate end 62.
- this free space 61 only slightly increases the respective step space and has the effect that this step space cannot be reduced to zero.
- 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 601 larger than in the case of a cylinder surface lying centrally to the eccentric axis 13. The minimum length la of the cavity 34a results from
- damping means T5 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 T5 to the stator or piston plates, or, for example, manufacture the piston plates themselves from resilient material.
- the center screw pump has 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 rolling rings in relation to the eccentricity e. 5
- 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 are 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 wear 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. With a sufficient groove depth 2nt, the rotor can be shifted in relation to the stator from the start, i.e. the end faces of the rotor are already displaced in the direction of the free spaces in relation to the parting planes 18.
- circular ring grooves 67 has the advantage over sickle-shaped ring grooves that a frictional 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 required volume increases in proportion to the outside diameter.
- the rolling rings 340 guided in the parting planes 18 of the stator are smooth on all sides with a rectangular cross section formed, while the free spaces are again formed as crescent-shaped edge grooves 651 and 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 projection 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.
- the pitch angle 1 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 seventh-pitch pitch, a rotor pitch 9.5 gives 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 the double pitch of the rotor, i.e. 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 pitch 6.5 for the V-ribs would in any case not result in problems if one always works with a pitch angle unit ir of half the size of the rotor pitch angle tr.
- the division of the V-ribs 21 on the eccentric shaft 22 then corresponds exactly to the stator division.
- 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.
- the Ex centering disks 223 must always be arranged relatively far from their alignment position to the parting planes 18 and thereby enable long displacement distances up to the friction system.
- 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.
- Fig. 22 shows an embodiment with non-rotatable
- the required expansion clearances are distributed over eccentric discs and wear ranks.
- 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.
- edge grooves 656 are provided in the circumferential shapes comparable to those in FIG. 21 in the wear rings 682.
- the depth of these grooves again largely corresponds to the dimension nt-, i.e. the eccentric disks can be substantially offset from the separating planes 18 of the stator and can engage far into the edge dimensions 656, in order to thereby form a positive driving clutch.
- a cylindrical pin serves as a longitudinal wedge 69 between the wear ring 68 and the eccentric disk 22.
- the longitudinal wedge couples in the circumferential direction and additionally provides a longitudinal guide for the relative axial adjustment of the wear ring on the eccentric disk.
- 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 7T 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 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 cylinder pin 73 is guided so 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. During the breaks in operation, however, a relative displacement in the circumferential direction is always achieved, so that the eccentric surface 221 is stressed approximately uniformly over the entire circumference.
- the embodiment shown in FIG. 27 also aims to gradually shift the clutch pitch between
- Rolling ring 340 and eccentric disc 22 There is parallel and at a lateral distance a to the central plane 66 of the eccentric disc in its thickened part a radially outward bore with a brake plug, which is loaded by a compression spring 75.
- a braking force is constantly exerted on the rolling ring 340, but this braking force is speed-dependent and stronger in the intended drive direction in the clockwise direction according to FIG. 27 than in the opposite direction.
- the braking force must be such that the rolling ring can only rotate in the drive direction when overloaded and / or is rotated over a small distance if the drive torque is removed in the meantime.
- 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.
- 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.
Landscapes
- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Rotary Pumps (AREA)
- Details And Applications Of Rotary Liquid Pumps (AREA)
Description
Claims
Applications Claiming Priority (2)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
DE3007881 | 1980-03-01 | ||
DE19803007881 DE3007881A1 (de) | 1980-03-01 | 1980-03-01 | Exzenterschneckenpumpe |
Publications (2)
Publication Number | Publication Date |
---|---|
EP0046779A1 true EP0046779A1 (de) | 1982-03-10 |
EP0046779B1 EP0046779B1 (de) | 1985-01-09 |
Family
ID=6095987
Family Applications (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
EP81900526A Expired EP0046779B1 (de) | 1980-03-01 | 1981-02-27 | Exzenterschneckenpumpe mit einem schraubenförmig gewendelten rotor |
Country Status (3)
Country | Link |
---|---|
EP (1) | EP0046779B1 (de) |
DE (2) | DE3007881A1 (de) |
WO (1) | WO1981002447A1 (de) |
Families Citing this family (1)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
DE3231157A1 (de) * | 1982-08-21 | 1984-02-23 | Armatec FTS-Armaturen GmbH & Co KG, 7988 Wangen | Exzenterschnecken-stroemungsmaschine |
Family Cites Families (6)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
FR1335006A (fr) * | 1962-06-12 | 1963-08-16 | Machine rotative utilisable notamment comme pompe | |
US3274944A (en) * | 1965-09-30 | 1966-09-27 | Frederick L Parsons | Screw vane pump |
DE2712121A1 (de) * | 1977-03-19 | 1978-09-28 | Streicher Foerdertech | Exzenterschneckenpumpe |
DE2712120A1 (de) * | 1977-03-19 | 1978-09-28 | Streicher Foerdertech | Exzenterschneckenpumpe |
DE2712122A1 (de) * | 1977-03-19 | 1978-09-28 | Streicher Foerdertech | Exzenterschneckenpumpe |
DE2905917A1 (de) * | 1978-02-20 | 1979-08-23 | Streicher Foerdertech | Exzenterschneckenpumpe |
-
1980
- 1980-03-01 DE DE19803007881 patent/DE3007881A1/de not_active Withdrawn
-
1981
- 1981-02-27 EP EP81900526A patent/EP0046779B1/de not_active Expired
- 1981-02-27 DE DE8181900526T patent/DE3168099D1/de not_active Expired
- 1981-02-27 WO PCT/EP1981/000016 patent/WO1981002447A1/de active IP Right Grant
Non-Patent Citations (1)
Title |
---|
See references of WO8102447A1 * |
Also Published As
Publication number | Publication date |
---|---|
DE3168099D1 (en) | 1985-02-21 |
WO1981002447A1 (en) | 1981-09-03 |
EP0046779B1 (de) | 1985-01-09 |
DE3007881A1 (de) | 1981-09-10 |
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