JPH0137599B2 - - Google Patents

Description

[Detailed description of the invention]

The present invention has a working chamber in the form of an annular hollow chamber, which is bounded in the radial direction by a rigid working chamber outer wall and a deformable working chamber wall constituted by a ring diaphragm, and in which the working chamber is radially bounded by a ring diaphragm. A rolling piston inside the diaphragm is configured to press the ring diaphragm against the outer wall of the working chamber in a circulating sealing and crimping area, and a rolling piston drive is engaged with the rolling piston. Regarding type diaphragm pumps. A diaphragm pump of the type described above has the advantage that the drive mechanism of the pump and the part of the pump for conveying the conveyed medium are completely separated and almost continuous conveyance is carried out. It is thus avoided that lubricants, lubricant gases or the like enter the transport area from the drive area. However, known diaphragm pumps still have many drawbacks. For example, according to British Patent No. 562,409, the rolling piston is guided rigidly and eccentrically in the radially outermost point of the rolling piston drive. Diaphragm pumps of the type mentioned at the outset are known, in which a positive connection is provided between the rolling piston drive, the rolling piston, the diaphragm and the outer wall of the working chamber. On the one hand, this allows the necessary pressing force to be applied to the ring diaphragm to seal it between the working chamber on the suction side and the working chamber on the discharge side, but at the same time it is ensured that the ring diaphragm is not continuously pressed during operation. This results in a significant transformation. Since such a ring diaphragm consists of an elastic molded part that can only be manufactured with relatively large manufacturing tolerances and is a worn part that must be frequently replaced, this known device does not allow for the transformation of the ring diaphragm. Movement is particularly disadvantageous and causes premature wear of the ring diaphragm. Hose pumps are already known which have an approximately cylindrical inner wall, along which a hose is guided in a loop, and in which a pressure body rolls on the inside of a hose rope. (West German Patent Application No. 2128487). However, in such hose pumps the working chamber is not formed by a rigid outer wall and a ring diaphragm, but from a hose of consistent cross-section. The hollow space of this hose cannot be completely sealed at the point where the crimping force is applied. The sealing properties achieved with hose pumps and thus the suction capacity are therefore limited. However, the transformation loads on the hoses of hose pumps are known to be very high, resulting in rapid fatigue of the hoses. Furthermore, rolling piston pumps are already known which have an approximately sickle-shaped working chamber which is delimited and divided by an outer wall of the working chamber, a partition slide and a rolling piston. In this case, an eccentric drive with a roller mounted tiltably inside the rolling piston comes into play. In the operating state, the rolling piston is pressed against the outer wall of the working chamber in the circulating sealing area by the sealing force (FR 1 256 125). However, this rolling piston pump does not have a ring diaphragm that is crimped against the outer wall of the working chamber in the circulating sealing and crimping area, and a partition slider separates the suction side of the working chamber from the discharge side. . In this case, in addition to the rollers arranged at a constant radial distance from the axis of rotation, a further pivotable roller is provided which is connected to the drive flange via a spring-loaded ring. There is. This not only results in higher manufacturing costs, but also increases the chance of failure and reduces the functionality of the rolling piston pump as the spring fatigues. Moreover, in this case there is no working chamber which is closed in cross section and is formed by the outer working chamber wall and the ring diaphragm. The object of the invention is to improve a diaphragm pump of the type mentioned at the outset in order to reduce the load on the ring diaphragm without adversely affecting the sealing action in the crimping region. The problem is that in the diaphragm pump of the type mentioned at the beginning, the rolling piston is tiltable around the point of contact with the rolling piston drive device.
The radial dimensions of the rolling piston drive, the thickness of the rolling piston plus the thickness of the diaphragm are such that there is play between the ring diaphragm and the outer working chamber wall when the rolling piston is not tilted. This was resolved by setting it to . An advantage of the present invention is that the rolling piston is free from movement between the rolling piston and the rolling piston drive due to the play provided between the ring diaphragm and the outer wall of the working chamber when the rolling piston is not tilted. can be tilted around the contact point,
The thickness of the ring diaphragm can be adapted without substantially interfering with the sealing action of the ring diaphragm in the crimping area. Advantageous embodiments of the invention are defined in the second patent claim.
It is described below. Experiments have shown that it is particularly advantageous to select the play in the range between 0.1 mm and 1.5 mm. Furthermore, if it is possible to prevent almost no tension from being applied to the ring diaphragm no matter which rolling position the rolling piston is in, the load on the ring diaphragm can be further reduced, and it can also act against the desired tilting of the rolling piston. The force generated by the ring diaphragm can be almost eliminated. According to the structure described in claim 4, a mechanically advantageous structure of the rolling piston drive device is obtained. Furthermore, with the configuration as set forth in claim 5, the reduction in circumferential elongation of the ring diaphragm can be achieved with a ring diaphragm that can be easily and inexpensively manufactured. This is because ring diaphragms can be made from only one material. According to the structure described in claim 6, the single or plural raised parts can be arranged particularly advantageously in terms of space. Claim 7 describes a relatively simple configuration that has been effective in practice. With the arrangement set forth in claim 8, the diaphragm section that is not pressed by the rolling piston can be pulled away from the outer wall of the working chamber more reliably. Furthermore, the features defined in claim 9 facilitate the production of a rolling piston with at least one ring groove and a ring body with possibly a curved section. If injection or casting parts are used for the rolling piston and/or the ring body, the slight dimensional errors that usually occur when manufacturing with this method are easily compensated for by tilting the rolling piston. be able to. The invention will now be explained with reference to the drawings. In FIG. 1, a diaphragm pump 1 is shown flanged to a drive motor 2. As shown in FIG. This diaphragm pump will be simply referred to as pump 1 from now on. The pump 1 has a suction connection 3 and a discharge connection 4, which in this embodiment are arranged in the same transverse plane. Pump 1 and drive motor 2 constitute a motor pump unit. The motor 2 has a fixed foot 37. The internal structure of the pump 1 is shown in particular detail in FIGS. 2 and 3. The pump 1 has a working chamber 6 which is an annular hollow chamber. This working space 6 is bounded by a rigid working space outer wall 7 and a deformable working space wall 8. The interior wall 8 of the working chamber is constituted by a ring diaphragm 9. The ring diaphragm 9 is circulated and pressed against the outer wall 7 of the working chamber by a pressing member 11 mainly formed by a rolling piston 10. As a drive device for the rolling piston 10, an eccentric drive device 1 is used.
2 is used. The eccentric drive device 12 is disposed inside a rolling piston 10 whose interior is hollow and coaxial with the outer circumferential surface 13 . The eccentric drive device 12 has a rolling piston drive member 14 . The rolling piston drive member 14 has a pivot arm 16 connected to a drive shaft 15 (FIG. 3). This pivot arm 16 carries at its free end a roller or the like, preferably a ball bearing 17 as in the exemplary embodiment. When the ball bearing 17 rolls as indicated by the arrow Pf1, the rolling piston 10 is continuously loaded along the inner wall 18. Correspondingly, the ring diaphragm 9 is also continuously and tightly pressed against the outer wall 7 of the working chamber in the crimping range Pr.
is crimped. As a result, the conveyed medium in the working chamber 6, in particular in the working chamber 6b on the discharge side thereof, is pushed away. This conveying medium is indicated by a dot on the discharge side and by an X on the suction side. According to the invention, the rolling piston 10 is arranged so as to be tiltable both with respect to the outer wall of the working chamber and with respect to the rolling piston drive member 14 due to the above-described structure of the pump 1. During this possible tilting movement, the inner wall 18 of the rolling piston 10 rolls on the outer circumferential surface of the ball bearing 17, whereas the outer circumferential surface 13 of the rolling piston 10 rolls on the outer circumferential surface of the ball bearing 17.
rolls on the outer wall 7 of the working chamber with a ring diaphragm 9 interposed therebetween. The tilting movement of the rolling piston 10 in the rolling plane is achieved by the following geometrical measures. In particular, preferably in the entire cross-sectional plane of the pump, the respective outer radius r1 (arrow pf4 in FIG. 2) of the eccentric drive 12 is in rotation with the respective wall thickness w of the ring diaphragm 9. dynamic piston 10
is approximately equal to the respective inner radius R1 of the outer wall 7 of the working chamber,
It is advantageously kept slightly smaller than that.
As a result, a play S is created between the pump parts 12, 17, 10, 9 and the outer working chamber wall 7, which is clearly visible in the assembled state shown in FIG. In the assembled state shown in FIG. 12, the rolling piston is in a neutral position in which it is not tilted relative to the rolling piston drive member 14, in order to clarify the dimensional relationship. . Correspondingly, if a certain play S is provided, there is also no crimping region Pr with a contact point between the rolling piston 10 or the diaphragm 9 and the outer wall 7 of the working chamber. Therefore, the position of the rolling piston shown in FIG. 12 is not the operating state of the pump, but is merely a position for explaining the type of pump operation. The play S is defined by the rolling piston 10 on the one hand and the ball bearing 1 on the other hand.
7, on the other hand, allows for the above-mentioned tilting relative to the working chamber outer wall 7. Even if the aforementioned play between the pump parts is actually zero, the compressive elasticity of the diaphragm 9 makes it possible for the rolling piston 10 to tilt slightly. It is clear from this that the amount of tilting movement of the rolling piston 10 is related to the play S mentioned above. In this case, the play S is the inner radius _R1 minus the outer radius r1, the ring thickness d, and the wall thickness w of the ring diaphragm 9 (FIG. 12). In practice, this play S is about 0 mm to 1.5 mm. As a result, the rolling piston drive member 14 (FIG. 2),
In particular, the contact point F between the ball bearing 17 and the inner wall of the rolling piston, and the rolling piston 10 against the outer wall 7 of the working chamber.
Alternatively, a delay angle A (second
) occurs. The tilting of the rolling piston 10 into the retarded position with the retardation angle A shown in FIG. It is given by the pressure ratio between 6a and 6b. Depending on this pressure ratio, the sealing pressure in the crimping range Pr is also affected. If other conditions do not change, the sealing force increases in the crimping range Pr with the pressure difference between the suction side working chamber 6a and the discharge side working chamber 6b. This makes it possible for the pump 1 to adapt the sealing force in the crimping range Pr to the respective operating state of the pump 1 without any structural changes. If the pressure difference between the suction side and the discharge side is small, a sealing pressure adapted to the reduced pressure ratio will normally act, whereas if the pressure difference is large, it will be correspondingly increased. Seal pressure will be applied. A seal adapted to the conditions relevant to the respective operation also has the advantageous effect of reducing wear on the ring diaphragm 9. This is because the ring diaphragm is only subjected to pressure loads commensurate with the respective sealing requirements. Furthermore, the required drive power of the pump, in particular its idle power, is thereby also reduced. The sealing pressure indicated by the arrow Pf2 in the crimping range Pr is also related to the magnitude of the delay angle A. In this case, when the delay angle A changes, the operating lever arm changes accordingly, so that the sealing pressure changes in inverse proportion to the delay angle A. Therefore, when the delay angle A is small, the sealing pressure (arrow Pf2) becomes larger than when the delay angle A is large, unless other conditions change.
This will be explained in detail later with reference to FIG. 13. In order to make the drawing easier to read, a larger delay angle A is shown in FIG. 2 than actually occurs. In practice, a delay angle A in the range of approximately 10 DEG has proven advantageous. however,
Other corresponding delay angles A can also be selected in order to achieve different sealing pressures in relation to the conveyed medium. The smallest possible delay angle A is the rolling piston 10.
This also works advantageously in the range of the top dead center OT where the suction connection 3 and the discharge connection 4 are provided.
In other words, in this top dead center range, the rolling piston 1
Since a change in the applied force occurs, the rolling piston attempts to carry out a corresponding compensating tilting movement. However, by virtue of the delay angle A, which is preferably chosen small, this tilting movement is kept within a narrow range. In addition, in the region of top dead center OT, the special shape of the ring diaphragm 9, as will be explained later, ensures that the rolling piston is guided perfectly, thereby ensuring the above-mentioned compensation. Tilting movements are at least damped. To achieve the same purpose, the circumferential angle B between the outflow opening 19 and the tightening piece 20 must be equal to the retardation angle A.
It is advantageous if it is smaller than that. In this case, angle B is advantageously smaller than the circumferential length section of angle C in which the clamping piece 20 is arranged.
As a result, the action of the force on the rolling piston, which may cause a counterbalancing tilting movement, is shifted to the area behind the outlet opening 19, in particular to the area of the clamping piece 20. The circumferential angle D between the inflow opening 21 and the outflow opening 19 is advantageously as small as possible, taking into account the circumferential dimensions of the clamping piece 20. Correspondingly, the effective working space of the diaphragm pump 1 is relatively large. Optionally, the circumferential angle D can also be adapted depending on the desired final pressure of the pump 1. The outer working chamber wall 7 is curved outwards in longitudinal section (FIG. 3). Correspondingly, the rolling piston outer circumferential surface 13 also has a vertical cross-sectional shape having a convex surface 13a matched to the curvature of the working chamber outer wall 7, taking into consideration the wall thickness w of the ring diaphragm 9. The respective coaxial cross-section of the outer working chamber wall 7 is of circular design, with respect to which the rolling piston drive member 14 is mounted concentrically. As can be seen from Figure 3,
The ring diaphragm 9 has a wall thickness w approximately equal to its effective length range 22 - viewed in the axial direction. In FIG. 3, the rolling piston 10 is located at a position offset by approximately 90 DEG relative to FIG. According to an important feature of the invention, the shape of the diaphragm and the diameter of the diaphragm are selected, at least with respect to its cross section, such that in at least substantially all rotational positions of the rolling piston 10 there is at least little tension on the diaphragm. It is. The shape required for this is shown in FIG. In this case, the ring diaphragm 9, in particular its effective length range 22, in the relaxed state has an approximately pear-shaped contour in the respective cross-sectional plane, in this case a radius of curvature K. The small, generally arc-shaped outer peripheral area of ₁ is provided approximately symmetrically with respect to the tightening piece 20 and adjacent to the tightening piece 20. A circular arc section with a large radius of curvature K ₂ is connected to this, preferably via a transition section 23 that extends approximately tangentially on both sides. The radius of curvature K ₂ of the diaphragm 9 is smaller than the inner radius R ₁ of the outer wall 7 of the working chamber and larger than the radius of curvature R ₂ of the outer peripheral surface 13 of the rolling piston 10 . The smaller radius of curvature K ₁ of the ring diaphragm 9 corresponds at most to the radius of curvature R ₂ of the outer circumferential surface 13 of the rolling piston 10 and is advantageously slightly smaller than this. The shape and diameter ratio described above means that the diaphragm 9
It was chosen so that no tension is applied to the pump, giving the pump as large a working volume as possible. The circumferential section (angle E) with a small radius of curvature K ₁ extends over an angle of approximately 70°.
The transition 23 forms a connecting transition to a section with a large radius of curvature K ₂ . Since the radius of curvature K ₁ of the diaphragm 9 is smaller than the inner radius R ₁ , the diaphragm 9 is easily separated from the outer wall 7 of the working chamber even in the working chamber 6 a on the suction side. On the other hand, the large radius of curvature K ₂ of the diaphragm 9 is selected taking into consideration that almost no tension is applied in the circumferential direction of the diaphragm and that a predetermined working volume is provided to the pump 1. The ring diaphragm 9 - also seen in the axial direction - has an effective length range ₂ with a large radius of curvature K 2
2 has a cylindrical inner surface (FIG. 6) in the relaxed state (FIG. 8). In contrast, the circumferential extent (angle E) with a small radius of curvature K ₁
are curved approximately symmetrically with respect to the tightening piece 20 in accordance with the curved portion 39 of the outer wall 7 of the working chamber, and even in the undeformed state, they gradually curve in the radial direction toward the tightening piece 20, and when tightened. The piece 20 has a curved section 24 that at least approximately corresponds to the curved outer wall 7 of the working chamber. In Fig. 4, the curved part 2
The curved chamber belonging to 4 is designated by 24a, and the wall that limits it is designated by 24b. In FIG. 7, the center lines of other portions of the effective axial length range 22 of the ring diaphragm, which are at different circumferential distances from the tightening piece 20, are indicated by dash-dotted lines. In this case, the center line M ₁ is that of the diaphragm part located inside the circumferential section with radius of curvature K ₂ , and the center line M ₂ is that of the tightening piece 2
This is the part of the diaphragm near zero.
The center line _M3 indicates the curvature at the center of the curved chamber 24a of the tightening piece 20. Due to the high stability of the diaphragm in this region, the curved chamber 24a in the region of the clamping piece 20 deforms only slightly and adapts to the curved shape of the outer circumferential surface 13 of the rolling piston 10. It is provided so that it can be used. 4 and 5 show the exact configuration of the unmodified ring diaphragm 9. In this case, the ring diaphragm 9 is shown in FIG. 4 with one side cut away. The tightening pieces 20 are oriented generally in the radial direction and protrude beyond the outer peripheral surface of the diaphragm 9. The tightening piece 20 extends over the entire width of the ring diaphragm 9 in the axial direction (FIG. 5). The ring diaphragm 9 has an annular lateral clamping flange 25 arranged approximately radially to the side of the effective length range 22 . This lateral tightening flange 25 also projects beyond the tightening piece 20. The tightening piece 20 has a curved chamber 2 in the range.
4a and a wall 24b delimiting the curved chamber 24a (FIG. 4). Advantageously, the inner circumferential surface 26 of the diaphragm forms a surface with at least approximately no bends relative to the rolling surface of the rolling piston 10, or a rolling surface on which the rolling piston rolls without impact. The already mentioned geometric relationships of the diaphragm, in particular its shape and diameter, and possibly its material properties,
The corresponding diaphragm section is selected in such a way that even if a negative pressure is applied at least in the working chamber 6a on the suction side, the corresponding diaphragm part is separated from the working chamber outer wall 7. For this purpose, it is advantageous if the outer circumference of the ring diaphragm 9 is smaller in each plane than the outer circumference of the working chamber or the inner wall 8 of the working chamber in the corresponding plane.
In this case, it is expedient for the ring diaphragm 9 to be virtually undeformable in the circumferential direction, but to be able to stretch somewhat in the axial direction. In addition to this, it is advantageous to provide inserts or similar reinforcements consisting of reinforcing threads 38, fabrics or the like, which are less flexible at least in the circumferential direction than the diaphragm material. This is shown in dash-dotted lines in FIG. 4 and in cross-section in FIG. In practice it has proven advantageous to design the diaphragm so that it has little circumferential expansion. On the other hand, a slight axial extension of the diaphragm 9 is caused by the diaphragm 9 being pushed into the curved part 39 by the rolling piston. However, it is advantageous if the diaphragm 9 is constructed in such a way that the maximum expansion in this direction is typically less than approximately 4.0 without deformation. Optionally, it is possible to design the ring diaphragm 9 to be somewhat longer in axial view than corresponds to the curved section 24 of the outer wall 7 of the working chamber. As a result, by a suitable design of the diaphragm 9, a diaphragm 9 can be obtained which is substantially free from tension in the axial direction even when the rolling piston 10 pushes the diaphragm 9 into the curved portion 24 of the outer wall 7 of the working chamber. The ring diaphragm 9 is made of an elastic material, preferably an elastomer, and preferably has a thickness of 1 mm to 4 mm. In order to additionally protect the diaphragm 9 against wear, at least the outer wall 7 of the working chamber and possibly the outer circumferential surface 1 of the rolling piston 10 are provided.
3 preferably has a covering layer of plastic or the like, which preferably has high sliding properties. 9 and 10 schematically show the pump 1 from the side with the rolling piston in different positions. In this case, the respective positions of the ring diaphragm 9 and the working chamber 6a on the suction side and the working chamber 6b on the discharge side are shown together. rolling piston 1
The rolling direction of 0 is indicated by arrow Pf3. The diaphragm pump 1 according to the invention can be constructed without valves. In FIG. 2, on the other hand, a valve 27 is additionally arranged at the outlet connection 4. The inlet opening 21 and the outlet opening 19 are arranged close to the clamping piece 20 of the diaphragm 9 and separated from each other by this or by the diaphragm extension. In FIG. 6, the axial course of the diaphragm 9, which is theoretically obtained when considering that no tension is applied to the diaphragm in the circumferential direction in practice, is indicated by a broken line. However, this shape differs from the cylindrical shape by at most a few tenths of a millimeter. Due to the symmetrical design of the diaphragm pump 1 according to the invention, the diaphragm pump 1 can be adjusted in its direction of rotation without special measures.
In addition, the conveyance direction can be switched. For this purpose, a drive motor 2 (first
It is sufficient to simply change the direction of rotation in Figure). For this purpose, it is advantageous to use a commutator motor as the drive motor. In the case of liquid conveying media, for example with a conveying efficiency of 1 to 100 lmin, the pump 1 operates in a speed range of approximately 500 to 4000 revolutions per minute, preferably in a speed range of 1000 to 1800 revolutions per minute. have. On the other hand, gaseous materials, for example, have a transport efficiency of about 5
~250l/min, then per minute
500 to 4000 revolutions, advantageously 3000 to 1 minute
A speed range of 3600 rpm is selected. If appropriate, the outlet opening can have relatively small dimensions in the circumferential direction of the pump and a relatively large internal diameter width in the axial direction. In contrast, it is advantageous if the inflow opening 21 has a relatively large inner diameter width or circumferential dimension in the circumferential direction of the pump. This aforementioned measure makes it possible to have a positive influence on the efficient operation of the pump 1. FIG. 11 shows a somewhat modified diaphragm pump 1a, in particular in which the eccentric drive 12a differs from, for example, that of FIG. This eccentric drive 12a for the rolling piston 10 has an auxiliary pressure element 28 which loads the rolling piston in the direction of the tilting movement in the rolling plane. The arrangement of the auxiliary pressing member 28 is such that the rolling piston 10 is aligned in the rolling direction (arrow) of the rolling piston drive device.
Pf3). That is, the rolling piston 10 tilts in the direction of arrow Pf4 about the tilting point Z (FIG. 11). As a result, an increased pressing force is generated in the pressing area Pf' of the rolling piston 10 or the ring diaphragm 9 against the outer wall 7 of the working chamber. Auxiliary pressing member 28
is the rotation direction of the rolling piston drive device 14a (arrow
A roller, in particular a ball bearing 2, which is placed behind this when viewed from Pf3) and loads the inner wall 18 of the rolling piston 10.
It has 9. The ball bearing 29 is arranged at the end of a coupling lever 30, which is pivotably connected at its other end to the pivot arm 16, which in this case is designed as a segment plate 31. The coupling lever 30 is loaded by a spring 32 connected to the pivoting arm 16 so as to pivot the ball bearing 29 towards the inner wall 18 of the rolling piston. The rolling piston 10 is subjected to different magnitudes of forces depending on its rolling position. In this case, the maximum force load occurs at bottom dead center, which corresponds approximately to FIG. 2, particularly in the direction of tilting. This is especially true for work chamber 6 on the discharge side.
b and the pressure ratio in the working chamber 6a on the suction side. As already mentioned, in the region of top dead center a significant change in the force load on the rolling piston 10 occurs. In the diaphragm pump 1a shown in FIG. 11, the preload that causes the rolling piston 10 to tilt is selected such that the rolling piston 10 is always maintained at a predetermined delay position. This avoids a tendency for the rolling piston 10 to oscillate, especially in the region of top dead center. The preload obtained by the auxiliary pressure member 28 can be adjusted by means of the spring 32 or the leverage or engagement point of the coupling lever 30 or the segment plate 31. In addition to its advantageous circular shape, the outer wall 7 of the working chamber can, in a given field of use, for example
Pf2), it can have a shape different from this circular shape, for example a partially helical shape or a shape similar thereto. This is made possible because the movement of the rolling piston 10 is automatically compensated for in this different manner. In this case, the delay angle A or A' of the rolling piston 10, and thus the sealing pressure, also changes. A small distance of the outer working chamber wall 7 from the center point of the drive shaft 15 results in a higher sealing pressure than if this distance were somewhat larger. The ring diaphragm 9 seals the working chamber 6 both to the rolling piston 10 and to the eccentric drive 12 or 12a, so that there is no special space between the eccentric drive 12, 12a and the pump housing 33. Seals are no longer necessary. A completely closed and tight working space is thus obtained. If necessary, the rolling piston drive member 14, in particular the ball bearing 17, can be constructed in an encapsulation, from which, for example, grease or the like can flow out and possibly be prevented from coming into contact with the diaphragm. Can be done. The pump according to the invention also works completely without oil, especially in the area of the working chamber. The pump casing 33 is composed of a ring body 34 having the working chamber outer wall 7 inside thereof and two casing plates 35. The casing plate 35 and the ring body 34 have a notch 36. This recess 36 serves to receive the lateral clamping flange 25 of the ring diaphragm 9. In this case, the recess 36 is designed in such a way that the lateral clamping flange 25 is tightly clamped, at least in the axial direction. Furthermore, from FIG. 2, the pump 1 itself is attached to the installation foot 37a.
It is clear that it is equipped with As already mentioned, the crimping force in the crimping range Pr between the ring diaphragm 9 and the rolling piston 10 is also related to the magnitude of the delay angle A (FIG. 2). 13th
By means of two different delay angles A ₁ or A 2, shown as two apex angles in the figure, the crimp point _Pr1 and
The pressure ratio in the corresponding crimp range, designated as Pr2, is shown. The delay angle is determined between the contact point F where the rolling piston drive member 14, in this case the ball bearing 17, contacts the inner wall 18 of the rolling piston 10 and the center of each crimping range or crimping point Pr1 or
limited by Pr2. In FIG. 13, individual components are omitted for the sake of simplicity.
Only the contact point and its analogues are filled in. Work chamber 6b on the discharge side and work chamber 6 on the suction side
The separation from a (FIG. 2) is effected by a ring diaphragm 9. In this case, one sealing point 20a is formed by the tightening piece 20, and the other sealing point is formed by a crimping area, for example Pr.
Separation is done at the sealing point 20a and each crimp point Pr1 or
13 by the separating straight line 40 and 40a between Pr2 and Pr2. First, we will discuss the pressure ratio when the delay angle A ₁ is relatively large. In this case, the dashed lines relate to each other. For standardization, half the length of the separation line 40 is the center of the separation line 40.
It is written as the pressure resultant line Q ₁ acting perpendicular to ml. This pressure resultant line Q ₁ is generated from the pressure difference between the working chamber 6b on the discharge side and the working chamber 6a on the suction side. In Fig. 13, the rolling direction (arrow) is the same as that shown in Fig. 2.
Pf1) is selected. The pressure resultant line Q ₁ cooperates with the lever arm 41 between the contact point F and the center point ml to form a rotational moment which causes the rolling piston 10 to tilt. The symbol Q′ ₁ indicates the force component produced by the pressure resultant line Q ₁ . Depending on the lever arm 41 and the force component Q' ₁ , a rotational moment acting around the contact point F is applied. For the pressing force at the crimping point Pr1, the same rotational moment occurs with the short lever arm. The lever arm 42 effective for this purpose is located between the contact point F and the crimping point Pr1. When a predetermined rotational moment is applied by the lever arm 41 and the force component Q' ₁ , a force corresponding to the force component N ₁ is generated perpendicularly to the lever arm 42 . From this, the sealing force _P1 acting perpendicularly to the working chamber outer wall 7 at the crimp point Pr1 can be derived. In comparison with the above-mentioned force relationship when the delay angle A ₁ is relatively large, the force relationship that occurs when the delay angle A ₂ is small will be described below. The line related to this small delay angle A ₂ is shown as a dash-dotted line. The separating straight line between the sealing point 20a and the crimping point Pr2 has the symbol 4.
It is indicated by 0a. For the purpose of simplification, the contact points F are written in the same place. Q ₂ is the pressure resultant line, from which the force component Q′ ₂ can be obtained. This force component Q' ₂ cooperates with the effective lever arm 41a to form a rotational moment about the contact point F. The lever arm 42a is relatively short because the delay angle A ₂ is small, and the force component Q' ₂ and the rotational moment from the lever arm 41a also act on the lever arm 42a, causing a force component N ₂ of a corresponding magnitude. bring about. This force component _N2 is transmitted to a pressure straight line located vertically at the crimp point Pr2, thereby generating a crimp force Pr2. This crimp force Pr2 is the crimp force that acts when the delay angle _A1 is large.
It is significantly larger than PR1. As can be seen, the effective crimping force in the crimping range Pr1 or Pr2 is inversely proportional to the associated delay angle A ₁ or A ₂ .
Reference numeral 43 designates the circle formed by the center point of the rolling piston 10, which rotates during operation of the pump. In FIG. 14, a diagram is shown in which the horizontal axis represents the delay angle A and the vertical axis represents the pressing force. This diagram is completely qualitative, so that the relationship between the delay angle A and the respective crimping force can be seen at a glance. The transition range, which has not been investigated in detail, is indicated by a dashed line, whereas the relationship, which is also explained in FIG. 13 from a predetermined delay angle Ax, is qualitatively indicated by a solid line. As already mentioned, a feature of a particularly advantageous embodiment of the diaphragm pump 1, 1a according to the invention is that the circumferential length of the individual circumferential sections of the ring diaphragm 9 is such that the circumferential extension of the diaphragm 9 The value is selected so that the value is at least approximately zero no matter where the piston 10 is located. As already mentioned, the advantage of such an arrangement of the ring diaphragm is that, on the one hand, the operating load on the ring diaphragm and the no-load operation required of the pump drive are kept relatively small, and on the other hand, the ring diaphragm 9 is This makes it possible to reinforce with the reinforcing threads 38 in the direction. The latter is not only advantageous for the stability and service life of the ring diaphragm 9, but also improves the operating conditions by ensuring a relatively large working volume. In addition, undesirable rocking of the ring diaphragm 9 can be reduced or eliminated by suitably designing the outer circumferential length U of the individual diaphragms. In order to provide the ring diaphragm 9 with such dimensions according to the invention, the individual circumferential length U of the diaphragm can be calculated in different transverse planes. For this purpose, 11 transverse planes are drawn in FIG. 6 as an example, Q0, Q, Q, Q, Q,
It is indicated by Q. In this case, the transverse plane Q0 lies in the plane of symmetry of the axis of the ring diaphragm 9. The transverse planes Q, Q, Q, Q, etc. located to the right and left of the transverse plane Q0, respectively, are the transverse plane Q as well as the curved part of the working room outer wall 7 shown in FIG.
It is arranged symmetrically with respect to 0. In this case, FIG. 15 shows the reinforcing thread 38 of the actual diaphragm 9.
The geometry of the central zone in which (Figs. 4 and 6) is located is shown with the diaphragm untensioned, i.e. without the rolling piston installed yet. . In this case, the length of the outer circumference Uu of the unstressed diaphragm in FIG. 15 is Uu=le+2t+lu. In this case, as can be seen from FIG. 15, the following geometric relationships are given. le=2π/360Rk・2γ t=√(−) ² −(−) ² lu=2π/360Ru(360−2γ) tanγ=t/Ru−Rk Note that Rk is the radius selected from the rolling piston 10 side and, taking into account the proportions of the wall thickness w, represent the aforementioned central zone marked by reinforcing threads 38. The angle γ can be selected from the viewpoints already mentioned and can be varied if necessary. This also applies to the radii Ra and Ru. FIG. 16 shows the circumferential course Uπ of the central zone of the diaphragm 9 shown in FIG. 15, which is displaced by the rolling piston 10. In this case, the rolling piston 10 has an angular position π
It is located in In this case, the circumferential progress Uπ
is given the following relational expression Uπ=b ₁ +b ₂ +2(b ₃ +t ₁ +t ₂ ). In this case, b ₁ =2πRk／3602α ₁ t ₁ =√(−) ² −(−) ² b ₂ =2πr _R ／360・2α ₂ t ₂ =√ ² −(− _R ) ² b ₃ =2πRi/ 360・180−(α ₁ +α ₂ ) Sinα ₁ =t ₁ /Ra−Rk Sinα ₂ =t ₂ /a. Such a relationship can be derived from FIG. 17 and 18 show the circumferential profile of the central zone of the ring diaphragm 9 for a rolling piston position of π/2 or 3/2π. In this case, the following relationship is given. U〓 _/2 = ³ 〓 ^/2 =b ₁ +b ₂ +b ₃ +b ₄ +b ₅ +t ₁ +t ₂ +t ₃ b ₁ =2πRk／360α ₁ t ₁ =√(-) ² -(-) ² b ₂ = 2πRi /360270 - (α ₁ + α ₂ ) t ₂ = √ ² - ( _i - _R ) ² t ₃ = √ (-) ² + ² - (- _R ) ² b ₃ = 2πr _R /360α ₂ b ₄ = 2πr _R /36090―(β ₃ +β ₆ ) b ₅ =2πRk／360180―(α ₃ +α ₆ ) Sinα ₁ = t ₁ /Ra―Rk Sinα ₂ = t ₂ /a tanα ₃ = a/Ra―Rk α ₄ = 90 - (β ₃ + β ₆ ) α ₅ = 180 - (α ₃ + α ₆ )

[Formula] β ₃ =90−α ₂ β ₆ =90−α _6Such a relational expression can be derived from FIGS. 17 and 18. In this case, if the radii Ra and Ru are fixed, different transverse planes Q0, Q, etc. (Figure 6)
By carrying out calculations and back calculations for this purpose, and by calculating the deformation, it is possible to determine that the circumferential course U of the central zone of the diaphragm 9 at any changing value is equal to It is possible to calculate whether the planes Q0, Q, etc. are substantially equal. The fact that the circumferential courses of the transverse planes Q0, Q, etc. of the diaphragm 9 are equal means that the diaphragm 9
This means that the extension or load is approximately zero no matter where the rolling piston 9 is located. In this case, in FIGS. 17 and 18 mentioned above, the geometric representation for the associated equations is shown in the same way as in FIGS. 15 and 16. However, in this case, the angle and distance display in the center of Figure 17 is
In FIG. 8, it is enlarged for ease of viewing. In particular, the curved portion 24 (for example, FIGS. 4, 5, and 7)
The special configuration of the diaphragm shape, as provided by (see figure), is ignored in the more geometrical mode of observation described above. Furthermore, Figures 15 to 18
The symbols U, R, b, t, etc. used in the equations in the preceding figures relate only to the equations associated with FIGS. 15 to 18. The embodiments of the present invention described above and shown in FIGS. 15 to 18 can theoretically provide a ring diaphragm 9 that is not subject to tension in the circumferential direction no matter where the rolling piston 10 is located. This shows that. This means that it is in practice possible to produce ring diaphragms that are subject to only small tension loads in the circumferential direction.
This diaphragm may be provided with reinforcing threads in the central zone, as in the features of the invention already described. Furthermore, referring to FIG. 2, the contact point of the rolling piston 10, that is, the contact point F between the rolling piston inner wall 18 and the ball bearing 17 on the one hand, and the crimping area on the other hand, is the same as that of the pivot arm 16 in FIG. Located on one side (left side) and on the other side (right side) in an extended position. This movement of the contact point with respect to the installation position shown in FIG. 12 is caused by the tilting of the rolling piston 10 according to the invention. The contact point 17 is somewhat offset in the direction of rotation of the swivel arm 16, whereas the crimping range P _R is
to follow the movement of. 19 and 20 show a ring diaphragm 9 which has been modified somewhat with respect to FIGS. 4 and 5.
a is shown. This ring diaphragm 9a
has an annular raised portion 44 as a reinforcing portion in the circumferential direction. This annular raised portion 44 protrudes from the inner circumferential surface 26 of the ring diaphragm 9a, and engages in the ring groove 45 provided on the rolling roller 10 when the ring diaphragm 9a is loaded by the rolling roller 10. (Figure 21). This annular groove 45 has a cross section that corresponds at least to the cross section of the raised portion 44 . In the embodiment shown in FIGS. 19 to 21, the ring diaphragm 9a has only one ridge 44 located approximately in the center of its radial width. However, if appropriate, a plurality of raised portions may also be provided. In this case, the cross-section of the bulge 44 and its bending resistance are selected such that the bulge 44 pulls away the diaphragm sections not loaded by the rolling piston 10 together from the outer wall 7 of the working chamber. In other words, if the raised portion 44 has a sufficiently large cross section and bending resistance, the length of the raised portion 44 in the circumferential direction will hardly change even when loaded by the rolling piston 10. The diaphragm sections not loaded by the piston 10 become pulled away from the working chamber outer wall 7 on both sides of the clamping piece 20 of the ring diaphragm 9a, so that a sufficiently large working chamber 6a, 6b is maintained. The raised portions 44 therefore have a similar function to the reinforcing threads 38 shown in FIG. It can further be seen from FIGS. 20 and 21 that the ridge 44 has a generally trapezoidal cross-section that tapers toward its inner free end. This cross section facilitates the penetration of the ridge into the ring groove 45 of the rolling piston 10. The ring diaphragm 9a with the bulge 44 ensures that the deformation work is as small as possible and that the corresponding diaphragm section does not displace the working chamber outer wall 7 under operating conditions, i.e. when suctioning from the vacuum.
It is becoming more and more separated from the world. The ring diaphragm 9a without reinforcing threads 38 is easy to manufacture. The rolling piston 10 is tiltably arranged and the tilting movement compensates for dimensional tolerances between the rolling piston 10 and the ring body 34, so that the pump housing 33 or the ring body 3 can be moved in a particularly advantageous manner.
4 and/or the rolling piston 10, preferably both pump parts 33 or 34, 10, can be constructed as injection molded or cast bodies. In particular, the component can consist of a metal casting, preferably a die-casting without further processing. Dimensional tolerances, particularly deviations from circularity, which often occur in such materials or in such manufacturing methods, are compensated for by the tilting movement of the rolling piston 10. A particularly inexpensive production is therefore obtained. This is because in this case expensive post-processing of the pump parts is avoided. It should be noted that the aforementioned pump parts, namely the rolling piston 10 and the pump casing 33, in particular its ring body 34, may also be made of plastic. In this case it is advantageous to produce these parts by injection molding. Such pumps can be used in particular with small dimensions, low operating temperatures and light loads. The invention is not limited to the illustrated embodiments, but can be implemented in various different ways and the individual features can be combined in any desired manner.

[Brief explanation of drawings]

The drawings show a plurality of embodiments of the invention, in which FIG. 1 is a side view of a pump motor unit with a pump shown partially cut away, and FIG. 2 is a side view of the pump of FIG. 3. Sectional view along the line, 3rd
4 is a front view with one side of the ring diaphragm cut away in an undeformed state, and FIG. 5 shows one side of the pump in FIG. 4. 6 is an enlarged vertical sectional view of the diaphragm in the lower half of the pump of FIG. 2; FIG. 7 is a side view of the ring diaphragm sectioned along a line; An enlarged vertical cross-sectional view of the area, Figure 8 is a schematic diagram showing the shape in one transverse plane of an unloaded ring diaphragm taken out from the pump, and Figure 9 is a case where the rolling piston is at the bottom dead center. FIG. 10 is a schematic diagram showing the shape of the ring diaphragm in one transverse plane when the rolling piston is rolled by 90 degrees with respect to FIG. 9. Schematic diagram shown, No. 11
The figure shows a diaphragm pump with an auxiliary pressing member, Figure 12 is a partial cross-sectional view of the pump in the assembled state, and Figure 13 shows the crimped state in the sealing range when the delay angle of the rolling piston is different. FIG. 14 is a diagram showing the relationship between crimp force and delay angle in the pressure range, and FIGS. 15, 16, 17, and 18 are based on the formulas described in the specification. 19 and 20 are diagrams corresponding to FIGS. 4 and 5 of a ring diaphragm with a protuberance,
FIG. 21 is an enlarged view of a portion Q surrounded by a chain line in FIG. 3 (the diaphragm has a raised portion). 1: Diaphragm pump, 2: Drive motor,
3: Suction connection part, 4: Discharge connection part, 6: Working chamber,
7: Working chamber outer wall, 8: Working chamber wall, 9: Ring diaphragm, 10: Rolling piston, 11: Pressing member, 12: Eccentric drive device, 13: Outer peripheral surface, 1
4: Rolling piston drive member, 15: Drive shaft, 1
6: Swivel arm, 17: Ball bearing, 18: Inner wall, 1
9: Outflow opening, 20: Tightening piece, 21: Inflow opening, 22: Length range, 23: Transition part, 24: Curved part, 25: Side tightening flange, 26: Inner peripheral surface,
27: Valve, 28: Auxiliary pressing member, 29: Ball bearing,
30: Combination lever, 31: Segment plate, 32:
Spring, 333: Pump casing, 34: Ring body, 35: Casing plate, 36: Notch, 37:
Fixed leg, 38: Reinforcement thread, 39: Curved part, 40, 4
0a: Separation straight line, 41: Leverage arm, 42: Leverage arm, 4
3: circle, 44: ridge, 45: ring groove.

Claims (1)

[Scope of Claims] 1. It has a working chamber in the form of an annular hollow chamber, which is limited in the radial direction by a rigid outer wall of the working chamber and a deformable working chamber wall constituted by a ring diaphragm, A rolling piston located inside the ring diaphragm in the radial direction is configured to press the ring diaphragm against the outer wall of the working chamber in a circulating sealing and crimping area, and a rolling piston drive device is engaged with the rolling piston. In the diaphragm pump of the type in which the rolling piston 10 is tiltable about the contact point F with the rolling piston drive device 14, the radial dimension r1 of the rolling piston drive device 14 and the rolling piston are The sum of the thickness d of ring diaphragm 10 and the thickness W of diaphragm 9 is set such that play S occurs between ring diaphragm 9 and working chamber outer wall 7 when rolling piston 10 is not tilted. A diaphragm pump characterized by: 2. The diaphragm pump according to claim 1, wherein the play S is approximately 0.1 mm to 1.5 mm. 3 The shape of the diaphragm 9 is similar to that of the rolling piston 10
The diaphragm pump according to claim 1 or 2, wherein the diaphragm pump is selected so that no tension is applied to the diaphragm 9 at almost any rolling position. 4 has a hollow cylindrical rolling piston and an eccentric drive for the rolling piston, the eccentric drive having a pivoting arm, the free end of the pivoting arm 16 being connected to the inner wall 18 of the rolling piston 10; Diaphragm pump according to any one of claims 1 to 3, characterized in that engaging rollers or ball bearings 17 are provided. 5. The diaphragm pump according to claim 1, wherein the ring diaphragm 9a has an annular raised portion 44 as a reinforcing portion in the circumferential direction. 6 The raised portions 44 protrude from the inner circumferential surface 26 of the ring diaphragm 9a, and the rolling piston 10 has at least one raised portion 4 as a corresponding member of each raised portion 44.
Diaphragm pump according to claim 5, characterized in that it has an annular groove (45) with a cross section suitable for receiving a ring groove (45). 7. One ring groove 4 in which the ring diaphragm 9a has only one ridge 44 located approximately in the center of the axial width of the ring diaphragm 9a, and the rolling piston 10a is aligned with said ridge 44;
Claim 5 or 6, having
Diaphragm pump as described in section. 8. The cross-section of the ridge 44 and its bending resistance are selected in such a way that the ridge 44 pulls together the diaphragm sections 24a, 24b which are not loaded by the rolling piston 10a away from the working chamber outer wall 7; A diaphragm pump according to any one of claims 5 to 7.