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
  • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
  • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
• • 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
  • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
  • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
  • F04C18/082—Details specially related to intermeshing engagement type pumps
  • F04C18/084—Toothed wheels
• • 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
  • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
  • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
  • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
  • F04C18/123—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with radially or approximately radially from the rotor body extending tooth-like elements, co-operating with recesses in the other rotor, e.g. one tooth
• • 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
  • F04C23/00—Combinations of two or more pumps, each being of rotary-piston or oscillating-piston type, specially adapted for elastic fluids; Pumping installations specially adapted for elastic fluids; Multi-stage pumps specially adapted for elastic fluids
  • F04C23/001—Combinations of two or more pumps, each being of rotary-piston or oscillating-piston type, specially adapted for elastic fluids; Pumping installations specially adapted for elastic fluids; Multi-stage pumps specially adapted for elastic fluids of similar working principle
• • 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
  • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
  • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
  • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
  • F04C18/14—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
  • F04C18/16—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type

Definitions

• the present invention relates to a rotary piston machine for compressible media, with at least two tightly enclosed rotary pistons which are mounted in a common housing and which are inevitably rotatable with one another, the rotary pistons having a plurality of disc-shaped sections which engage in pairs and whose thickness decreases in the direction of the pressure side, each of which Disc has at least one lateral surface and a core surface, which are formed by guidelines guided along a circular arc with a center on the axis of the respective rotary piston and are each connected by an intermediate surface.
• Rotary pistons for vacuum pumps or positive displacement pumps for gases are usually manufactured in the form of pairs of screw spindles.
• these screw spindles have a variable pitch.
• screw compressors for gases with two interlocking screws, the pitch of which decreases steadily towards the pressure side, are known.
• the manufacture of pairs of screw spindles with a variable pitch is technically difficult, in particular since the screws should mesh with one another with as little play as possible in order to keep pressure losses small.
• the manufacture of this type of screw compressor is expensive.
• Roots blowers are known, with two disc-shaped rotary pistons which are in engagement with one another.
• the air flow is transverse to the axes of rotation of the rotary pistons, so that such compressors are suitable for large amounts of air, but only for small degrees of compression.
• several compressor units of this type must be connected in series, or assembled to form a multi-stage Roots pump.
• DE-2934065 discloses such helical rotary lobes in a rotary lobe machine of the type mentioned in the introduction.
• the spindles have a pseudo-thread-like groove which is formed by step-shaped recesses which follow one another in the helical line and are provided with flat boundary surfaces which are perpendicular to the spindle axes.
• a correspondingly shaped thread-like comb of the counter spindle engages in this groove in the plane defined by the two spindle axes and closes a groove volume with each thread turn, so that when the spindles are rolled off, the comb displaces the groove volumes with compressible medium from the inlet to the outlet , whereby the groove volumes are changed and the desired pressure difference between inlet and outlet is achieved.
• the spindles In cross section, the spindles have a semicircular contour with a cutout which is delimited by the core surface and two step-forming intermediate surfaces. The sector angles of the outer lateral surfaces and inner core surfaces have the same value, namely 180 °.
• a disadvantage of this rotary piston machine is the large number of step-shaped boundary surfaces which are necessary to form the pseudo-thread-like groove, the manufacture of which requires a large number of chip-removing operations.
• Another disadvantage is the high precision of the intermediate surfaces, which is necessary in order to minimize the pressure losses from stage to stage.
• the document AT-261792 also describes a rotary lobe machine of this type in which the helical rotary lobes consist of individual disks which are identical in the end cut. Each disk has two diametrically opposed outer lateral surfaces and two diametrically opposed inner core surfaces, the sector angles of which are all the same (90 °). With this configuration of the disks and this offset arrangement in the rotor, the gap widths between opposing disks must be kept as small as possible. The jacket and core surfaces are therefore connected by intermediate surfaces which are designed as elongated epicycloids to effect the seal between the panes. Consequently, both their profile and the exterior
• Synchronizing device of the machine very precisely - and consequently consuming - can be produced.
• this publication provides for a rounded shape to reduce the thermal loads on the flank tips, these cannot be avoided in the case of gas reflux.
• the aim of the present invention is to produce a rotary lobe machine with a high degree of compression, in particular a vacuum pump, in which the ultimate vacuum is to be better than with rotary vane, approximately similar to that of multi-stage Roots pumps.
• the production should be less expensive than that of multi-stage pumps and also less expensive than that of screw pumps.
• an internal compression of the compressible medium or gas should take place in order to reduce energy consumption and Achieve operating temperature.
• the noise level during operation should be as low as possible.
• a stepped spiral gear with horizontal intermediate sections between two chambers is formed in the individual, unassembled rotary piston, with varying step heights.
• a chamber sequence is formed in the axial direction, with optionally variable volume, i.e. optionally variable internal compression, through selectable thickness variation of the disc-shaped sections.
• each rotary lobe can be manufactured in one piece, which significantly improves the dimensional stability during operation and is thermally less critical than a stack of individual disks. If the operating temperature of the rotary lobe machine is low due to the application, the rotary lobes can also be assembled from sequences of individual profile disks arranged axially one on top of the other, which saves manufacturing costs.
• the word "disk”, if not particularly specified, is used both for individual profiled disks and for disk-shaped sections of a one-piece piston.
• the displacement machine according to the invention is contactless and constantly rotating.
• the gaps between the two rotating lobes can be broken down into three types:
• Mantle surface / core surface of opposite disc-shaped sections these linear columns are determined by the accuracy of the manufacture of the cylindrical surfaces of the pistons and the distance between the two axes of rotation. Small gap values can be achieved with common manufacturing technology.
• these gap widths are not critical and can be in the millimeter range, which greatly facilitates the manufacture of the intermediate surfaces. Since these gap widths also determine the permissible angular play between the rotary pistons, this permissible angular play is very large, so that the requirements for the synchronizing device of the rotary piston machine are reduced and their selection or implementation is facilitated.
• the theoretically curved intermediate surfaces ie the parallelepiped surfaces, the connecting the outer surface and the core surface, ie outer cylinder and core cylinder, of a disc-shaped profile section, with the rotary pistons rotating in opposite directions, do not fulfill a critical, functionally essential sealing function and thus describe a theoretical maximum contour.
• a profile contour of the intermediate surface that is somewhat smaller or flattened than this theoretical maximum contour and is easier to produce, for example an undercut and / or almost straight contour, can therefore be preferred and is quite functional. This further increases the angular play permitted during operation.
• the difference between the sector angles of the outer lateral surface and the core surface of a disk-shaped section is expediently large.
• the sector angle of this lateral surface is preferably less than 90 ° and particularly preferably less than 60 °.
• Such a disk lies opposite a disk of the other rotary piston, with a sector angle of the outer lateral surface which is correspondingly greater than 270 ° or greater than 300 °.
• the chambers of a respective rotary piston are preferably designed in such a way that the intermediate surfaces of one disk each with an intermediate surface of an adjacent disk form a continuous intermediate surface with a common guideline.
• the synchronizing device of the rotary piston machine according to the invention can be selected in such a way that the two external-axis rotary pistons have an opposite direction of rotation.
• the outer diameter of the rotary lobes, the The diameter of the core cylinder and the transmission ratio can then be selected such that the pistons roll against one another without sliding, the outer surface of a disk-shaped section rolling on the core surface of the opposite section. If the numbers of the outer surfaces and core surfaces of a disk-shaped section are the same as those of the opposite section of the other rotary lobe, a ratio of 1: 1 should be selected. However, if these numbers are different, the translation must be selected accordingly.
• the two outer-axis rotary pistons have the same direction of rotation.
• the two rotary pistons are internally axis, i.e. designed as an outer rotor and inner rotor, with an additional G rotor.
• the disk-shaped sections of a respective rotary lobe have only two alternating end cut profile contours.
• the diameters of the jacket cylinders and core cylinders of external-axis rotary pistons can each be the same, with in a same plane perpendicular to the piston axis, the section of the first piston having an end-cut profile contour, while the opposite section of the second piston has the other end-cut profile contour having.
• the two rotary pistons can also be configured as main rotors and secondary rotors with different diameters and thus different shaft powers - up to 100: 0% - which offers advantages in the execution of the synchronization device.
• sequences of sections with different end cut profile contours alternate with circular locking disks, so that a respective piston has sections with three or more different profile contours.
• Figure 1 is a side view of a first embodiment of a rotary piston according to the invention, with 14 superimposed disks, numbered from 0 to 13;
• Figure 2 is a side view of the corresponding second rotary piston of the first embodiment
• Figure 3 is a top plan view, from the suction side, of the assembled rotary lobes of Figures 1 and 2, with section "0" of the rotary lobe of Figure 2 omitted;
• FIG. 4 is a section / sequence diagram or angle of rotation, which schematically shows the functioning of the first embodiment
• FIG. 5 is a side view of a pair of eleven-section lobes according to a third embodiment, with a main rotor having eleven sections, numbered 0 to 10;
• FIG. 6 is an end section through section 1 of the assembled rotary pistons of FIG. 5;
• Figure 7 is an end section through a section 2 of Figure 5;
• Figure 8 is a section / rotation angle diagram schematically showing the operation of the third embodiment;
• Figure 9 is a section / rotation angle diagram schematically showing the operation of a fourth embodiment
• Figure 10 is a section / rotation angle diagram schematically showing the operation of a fifth embodiment
• FIG. 11 is a side view of a pair of rotary pistons according to a sixth embodiment, with 17 sections, numbered from 0 to 16;
• FIG. 12 is an end section through section 1 of the assembled rotary pistons of FIG. 11;
• FIG. 13 is an end section through section 2 of the assembled rotary pistons of FIG. 11;
• Figure 14 is an end section through section 3 of the assembled rotary pistons of Figure 11;
• FIG. 15 is an end section through section 4 of the assembled rotary pistons of FIG. 11;
• Figure 16 is a section / process diagram, or rotation angle, schematically showing the operation of the sixth embodiment
• Figure 17 is a section / flowchart schematically showing the first nine sections of a seventh embodiment and their interaction;
• Figure 18 is an end section of section 1 of the outer rotor of the embodiment of Figure 17;
• Figure 19 is an end section of section 2 of the outer rotor of the embodiment of Figure 17;
• Figure 20 is an end section of section 1 of the inner rotor of the embodiment of Figure 17;
• Figure 21 is an end section of section 2 of the inner rotor of the embodiment of Figure 17;
• Figure 22 is an end section of the crescent-shaped G rotor of the embodiment of Figure 17;
• FIG. 23 is a partial view in axial section of a section of the inner rotor and the parts of the outer rotor surrounding it in an eighth embodiment.
• the rotary pistons are mounted on the outside and parallel axes in a housing (not shown) with two cylindrical bores with an external synchronizing device.
• the rotary lobes have an opposite direction of rotation.
• the rotary pistons have 14 disk-shaped sections, namely two end sections (0, 13) for the inlet and outlet of the medium, and profile sections (1-12) with two different, alternating profile contours, each with a section that has an outer lateral surface (ml) has with a small sector angle, alternates with a section which has a lateral surface (Ml) with a large sector angle.
• FIGS. 3 and 4 illustrate the progressively rotated angular position from one section to the next, ie 72 ° from one section to the identical next but one section, with an intermediate surface (zl) of a section in each case above or below, seen in the axial direction, an intermediate surface an adjacent section of the other Profile contour is arranged. A chamber is thereby formed, which (see FIG.
• the latter type of gap determines the permissible angular play and is not critical, i.e. can be in the millimeter range, which opens up many implementation options for the synchronization device.
• a compression rate of 1: 4 is realized, which leads to significant savings in energy consumption and heat development.
• the total number of profile sections is minimized.
• sections 1 and 2 have the same thickness. From section 2 to section 3 the thickness decreases by a factor of about 1.4; the thicknesses of sections 3 and 4 are again the same, etc. With this distribution of the thicknesses of the sections, where two successive and opposite sections of one and the other rotary lobes have the same thickness, the energy distribution is about 50:50% for each rotary pistons.
• the thickness of the sections could also decrease from each section to another, according to a selectable geometric rule.
• the five sections form pump sections P1-P5. They are separated and surrounded by six sections 0, 2, 4, 6, 8, 10, which only have a short-angled core surface section (k3), and each form a control section S, which forwards the gas to the next pumping section.
• the thickness of the five pump sections, from P1 to P5 can decrease by about one third from approximately 70 millimeters to a thickness of 13 millimeters, while each control section S has a thickness of 10 millimeters.
• the total length of the main rotor then measures around 240 millimeters.
• the diagram in FIG. 8 shows an embodiment in which the core diameter of the main rotor is the same as the outer diameter of the secondary rotor. With a gear ratio of 1: 1, the runners roll on each other without one sliding over the other. Under these circumstances, the energy distribution between the main and secondary rotor is around 75:25%.
• the diameters of the main rotor and the secondary rotor are also very different.
• the main rotor also has two alternating profile contours which are different in the end cut and similar to the third embodiment.
• the secondary runner has three different profile contours, namely in the following sequence: a section 1, which consists of a simple core disk, a section 2, in the form of an outer cylinder with a low-angle cutout, a section 3, which in turn consists of a core disk Section 4, which consists of a full outer cylindrical disk and forms a locking disk.
• FIG. 10 shows a fifth embodiment diagrammatically.
• the main rotor has two alternating different end profiles, each having two identical outer lateral surfaces and two identical core surfaces, each diametrically opposite.
• the relative sector angle dimensions of the jacket and core surfaces vary from section to section as in the previous embodiments.
• the secondary runner has only one outer surface and one core surface, alternating between large and small angles.
• the synchronization device is designed such that the number of revolutions of the secondary runner is twice the number of revolutions of the main runner. With this design, a strongly asymmetrical energy distribution is achieved, namely approximately 85% on the main rotor and approximately 15% on the secondary rotor.
• a rotary piston can be manufactured as a monoblock, which significantly improves the dimensional stability during operation; the large gap lengths along the flow between the rotary lobes ensure a good seal and thus a good final vacuum; the permissible large play facilitates the manufacture and assembly and use of the synchronization device.
• the intermediate surfaces of the main rotor are designed without an undercut, which simplifies individual work steps in production.
• the power components of the driving rotary piston and the driven rotary piston are very different, which is additionally the case with the Selection and design of the synchronization device offers advantages.
• a sixth embodiment the pair of rotary pistons of which is shown in FIGS. 11 to 15, consists of a non-contact, parallel-axis, two-axis, external-axis, constantly rotating displacement machine, with a housing with two cylindrical bores and an external synchronizing device, the two rotary pistons having the same direction of rotation to have.
• the different diameters of the rotary lobes are designed as main and secondary rotors. Both the main runner and the secondary runner have at least three different profile types. In the exemplary embodiment shown by FIGS. 12 to 15, both the main runner and the secondary runner have four different profile types, which form sequences of four different disk-shaped section pairs
• FIG. 12 a first section in which the main rotor has a large-angle lateral surface (M6); the sector angle of the core surface can be kept very small or, as shown in FIG. 12, it can even be left out entirely, so that the outer lateral surface of this section is only interrupted by a crescent-shaped asymmetrical cutout.
• This section serves as the first control disk S and lies opposite a first section of the secondary rotor, which simply consists of a core cylinder disk;
• a second section P of the main rotor has a core surface (K6) whose sector angle is greater than 180 °, an extremely short outer surface (m6) and two elongated intermediate surfaces (z6). Opposite this is a second section of the secondary rotor, with an outer lateral surface (M6 ') whose sector angle is greater than 180 °, with a minimal core area (k ⁇ 1 ) which, as can be seen in FIG. 13, is due to the continuous flow of the two am Intermediate surfaces (z6 ') tangent to the core cylinder can completely or almost completely disappear.
• This section forms the actual pumping stage of the sequence;
• the third section of the main rotor (FIG. 14) is identical in shape to the first section, but is arranged in a plan-symmetrical manner, as can be seen from FIGS. 12 and 14.
• This section serves as a second control disc.
• the opposite third section of the secondary rotor is designed as a simple core cylinder disk;
• the fourth section (FIG. 15) of the main rotor is a simple core disk and serves as a channel K for the compressible medium. Opposite this is a fourth section of the secondary rotor with an uninterrupted outer surface, which serves as a locking disk.
• FIG. 11 shows the full structure of an exemplary embodiment with 17 disk-shaped sections, namely two end disks (E), 0 and 16; three full sequences S-P-S-K, of the four sections just described, 1 to 4, 5 to 8, 9 to 12; and an incomplete sequence, S-P-S, i.e. with a first control disk 13, a pump stage 14 and a second control disk 15.
• the control disks S of the main rotor can all consist of thin disks, since they only serve to transfer the medium from one pump stage P to the following channel K and again to the next pump stage.
• the gradation of the axial extent of the pump stages and the duct stages can be subject to various computational rules for the intended purpose.
• the Taffei 1 shows an example of two levels, in which the Thickness of the thickest stage, namely pump stage 1, was arbitrarily used with 1.
• Example 1 the thickness of the stages decreases progressively in the order P1, K1, P2, K2, etc., while in Example 2 the thicknesses of the pump stages on the one hand and the channel stages on the other hand decrease but alternate in thickness.
• a thickness Pl 49 mm and a thickness of the control disks of 8 mm, with the gradation of example 2, the overall length of the main rotor is approximately 240 mm.
• This sixth embodiment results from the diagram in FIG. 16.
• an axial chamber sequence is realized in an external-axis displacement machine with pistons rotating in the same direction.
• the shaft powers of the pistons vary widely, ie the energy distribution is extremely asymmetrical, up to 100: 0%.
• This embodiment has the following advantages: the undercut-free contours allow very simple production; in particular, monoblock production is easy to carry out; the very large permissible play is advantageous for production and assembly; the large gap lengths along the flow allow a good final vacuum; the same spin and the large allowable game open up additional opportunities for the
• timing belts In view of the low power of the secondary rotor, even timing belts can be used.
• both rotary pistons are generally cylindrical, with parallel axes of rotation.
• the guidelines, the redirection of which form the lateral surfaces, core surfaces and intermediate surfaces of the disk-shaped sections, are cylindrical guidelines which are parallel to the axes of rotation.
• the rotary pistons can also be shaped conically, the guidelines, the deflection of which defines the peripheral surfaces of the disks, are the guidelines of a cone, so that the disks on their circumference are conical, and their diameters gradually decrease towards the pressure side.
• the axes of rotation of the two pistons are then not parallel, but have an intersection.
• the diameter variation causes internal compression. This diameter variation can be used in addition to the variation in the thickness of the disks or instead of the variation in the thickness of the disks.
• Figures 17 to 22 represent a seventh embodiment, namely a non-contact, parallel-axis, two-axis, inner-axis, constantly rotating displacement machine.
• the machine has a hollow outer rotor, an inner rotor and a crescent-shaped G-rotor, which lies between the outer and inner rotor.
• the rotors have the same direction of rotation, as shown in Figure 17.
• the outer rotor (A) and the inner rotor (I) have a plurality of mutually interlocking disc-shaped sections, the thickness of which decreases in the direction of the pressure side, each disc having at least one circumferential surface and a core surface which is formed by a circular arc centered on the axis of the respective rotor-led guidelines
• the gaps between end faces of two sections sliding over one another are gaps between two spherical surfaces (Ku, Ku '), as shown in FIG.
• the large gap lengths along the direction of flow also provide a good seal and a good final vacuum in this embodiment.
• the synchronizing device can be implemented as a simple lubrication-free coupling mechanism, for example as a universal joint, inside the displacement machine or vacuum pump.

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• Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Applications Or Details Of Rotary Compressors (AREA)
• Reciprocating Pumps (AREA)
• Compressors, Vaccum Pumps And Other Relevant Systems (AREA)
• Details Of Reciprocating Pumps (AREA)
• Hydraulic Motors (AREA)
• Pistons, Piston Rings, And Cylinders (AREA)

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