EP0147654A2 - Machine en particulier pour la compression et le déplacement des fluides - Google Patents

Machine en particulier pour la compression et le déplacement des fluides Download PDF

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Publication number
EP0147654A2
EP0147654A2 EP84114423A EP84114423A EP0147654A2 EP 0147654 A2 EP0147654 A2 EP 0147654A2 EP 84114423 A EP84114423 A EP 84114423A EP 84114423 A EP84114423 A EP 84114423A EP 0147654 A2 EP0147654 A2 EP 0147654A2
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EP
European Patent Office
Prior art keywords
annular piston
cylinder
piston
ring
machine 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
Application number
EP84114423A
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German (de)
English (en)
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EP0147654B1 (fr
EP0147654A3 (en
Inventor
Kurt Gerhard Fickelscher
Hans-Peter Schabert
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Individual
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Individual
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Publication of EP0147654A2 publication Critical patent/EP0147654A2/fr
Publication of EP0147654A3 publication Critical patent/EP0147654A3/de
Priority to JP26631285A priority Critical patent/JPS61192884A/ja
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Publication of EP0147654B1 publication Critical patent/EP0147654B1/fr
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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
    • F04C2/30—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members
    • F04C2/34—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in groups F04C2/08 or F04C2/22 and relative reciprocation between the co-operating members
    • F04C2/356—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in groups F04C2/08 or F04C2/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the outer member
    • F04C2/3562—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in groups F04C2/08 or F04C2/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the outer member the inner and outer member being in contact along one line or continuous surface substantially parallel to the axis of rotation
    • F04C2/3564—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in groups F04C2/08 or F04C2/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the outer member the inner and outer member being in contact along one line or continuous surface substantially parallel to the axis of rotation the surfaces of the inner and outer member, forming the working space, being surfaces of revolution
    • 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
    • F04C15/00—Component parts, details or accessories of machines, pumps or pumping installations, not provided for in groups F04C2/00 - F04C14/00
    • F04C15/0057—Driving elements, brakes, couplings, transmission specially adapted for machines or pumps
    • F04C15/0061—Means for transmitting movement from the prime mover to driven parts of the pump, e.g. clutches, couplings, transmissions
    • F04C15/0065—Means for transmitting movement from the prime mover to driven parts of the pump, e.g. clutches, couplings, transmissions for eccentric movement

Definitions

  • the invention relates to a machine, in particular a working machine for compressing and conveying fluids, with a cylinder, with a thin-walled annular piston which is arranged eccentrically with respect to the cylinder and which lies flat against a cylinder wall, with a separating element, by means of which a suction is provided between the cylinder and the annular piston - And a pressure chamber are separated from each other, and with a rotating body to transmit a rotary motion to the ring piston.
  • the ring piston is designed as a thin-walled, elastically deformable ring, which is arranged either within a circular cylinder or around a circular cylinder.
  • the ring piston is on or within a arranged at least two projections rotating body and it is pressed in the region of these projections against the inner wall or outer wall of the cylinder, so that the rotary piston rolls in a predetermined angular range on the wall of the cylinder during the rotational movement of the rotating body.
  • the ring piston which can also be referred to as a deformed sleeve or rolling membrane, is subject to high strength requirements.
  • a deformed sleeve or rolling membrane is subject to high strength requirements.
  • high fatigue strength stresses occur, so that with regard to an economical suction volume, large deformations, i.e. deviations from the circular ring, occur which can only be achieved in practice with very thin-walled sleeves . This limits the pressure load on the pumped medium to smaller values.
  • the task is based on the disadvantages of the invention, to develop the machine with little design effort so that the loads on the various components, in particular annular pistons and cylinders, are reduced and there is little wear is.
  • the machine should have a high mechanical and thermal efficiency and can be used for vacuum and high pressure conveying. High strength loads on the cylinder should be avoided and resilient restoring forces on the drive bearings should be reduced. A low surface pressure should occur between the annular piston and the cylinder so that inexpensive, simple materials can also be used. Low fatigue strength should be given.
  • the machine should run smoothly and have a long service life, and the housing construction should enable simple and reliable cooling, in particular optionally air or water.
  • the machine should be suitable for oil-free operation, grease lubrication or oil flooding.
  • the machine should be designed such that a material with lower strength properties and preferably good heat conduction can be used for the cylinder, reference being made here in particular to aluminum alloys and, in the case of aggressive gases, to bronzes and austenite.
  • the ring piston is essentially annular, the deviation from the circular shape being at most 5% of the ring piston diameter, that the center of the ring piston with the rotating body in the direction of the rolling region from the center of the cylinder by the sum the eccentricity and a deformation is arranged offset, the eccentricity being equal to half the diameter difference between the cylinder and the annular piston.
  • the deformation ensures a flat contact in the predeterminable rolling range, which is preferably greater than 10 angular degrees, the center of the Ring piston is rotated essentially on a circular path around the center of the cylinder.
  • the machine according to the invention is characterized by a simple construction and a high level of functional reliability.
  • the essentially annular annular piston according to the invention shows only a relatively small deviation from the circular shape, namely a maximum of 5%, so that a low deformation and a low material stress are achieved in a surprisingly simple manner .
  • the ring piston clings to the cylinder or to the inside or outside of the cylinder in the rolling area.
  • it is ensured that the predetermined friction torque is retained when the separating slot is rolled over and that the suction and pressure slots are closed in a sealing manner.
  • the deformation is in the range between 0.2 to 2%, preferably 0.5%, of the ring piston diameter.
  • the ring piston fits snugly against the cylinder wall over a large angle rich is ensured.
  • the configuration according to the invention keeps the ovalization, that is to say the deviation from the exact circular shape of the annular piston, less than 5% and preferably less than 3% of the outside diameter. Due to the reduction of the surface pressure and uniformity in the rolling area, it is also possible to use materials that have good thermal conductivity and have lower strength properties, in particular aluminum alloys and, moreover, bronze or austenite for aggressive gases.
  • the machine can be used for vacuum and high-pressure conveying without any significant changes to the essential components.
  • single-stage vacuums of up to 99% and final pressures of up to 16 bar could be achieved, a good heat-conducting material with low strength properties being used for the cylinder housing.
  • the special housing construction allows either air or water cooling with the same housing.
  • the machine is suitable for oil-free operation, grease lubrication or oil flooding while maintaining the essential design features.
  • a high mechanical and thermal efficiency is achieved, with great smoothness and low wear.
  • the ring piston can be arranged in the cylinder or on the outside around the cylinder.
  • a number of rotatable drive rollers are arranged offset in the circumferential direction relative to one another in such a way that the drive rollers closest to the rolling region are at a substantially greater angular distance from one another than the other angular distances between the drive rollers.
  • the drive rollers can have different radii be arranged with respect to the center of the rotating body and / or the roller diameters can be specified differently accordingly.
  • the drive rollers can also have different wall thicknesses. It should be noted that the increase in eccentricity known as deformation also serves to compensate for thermally induced out-of-roundness.
  • the described embodiment is particularly suitable for relatively high pressure ranges.
  • the ring piston can bend between the individual drive rollers of the ring body in accordance with the relative position of the ring piston to the separating slide. Furthermore, in the top dead center, in which the ring piston rests in the region of the separating slot, the ring piston can lift off the diametrically opposite drive roller, as a result of which additional stresses act on the further drive rollers and result in bearing losses and a reduction in efficiency.
  • the comparatively high speed of the drive rollers can also limit the drive speed upwards in accordance with the difference in diameter of the annular piston and drive rollers. Due to the limit speeds for the bearings, especially roller bearings, of the drive rollers, a limit must be observed.
  • the annular piston is supported on a bearing ring or the like which has a reduced wall thickness in the rolling region and is preferably designed as a carrier of the same strength.
  • the ring piston is floatingly arranged on two eccentrically mounted drive rollers or the like, the eccentrics being arranged offset in the circumferential direction by a predetermined angle.
  • the annular piston can also be arranged on elastically resilient elements which are rotatably arranged with respect to the eccentric by means of a bearing.
  • the roll-off area is evenly distributed and stress peaks between the annular piston and the cylinder wall are largely reduced and avoided. Impacts and shock-like loads, in particular when rolling over the separating slot, are reduced, so that materials for lower loads for the cylinder can also be used. Inexpensive and, above all, good heat-conducting materials, in particular aluminum alloys, can be used for the cylinder or the cylinder housing. It is very important for heat dissipation both from the interior of the ring piston and from the housing. With suitable ventilation measures, such as ventilation slots or the like. both in that Heat as well as inside the ring piston can be dissipated under optimal conditions. Further advantages and features essential to the invention result from the exemplary embodiments.
  • F ig. 1 schematically shows a cross section of a compressor with a cylinder 2 in which an annular piston 4 is rotatably arranged.
  • the annular piston 4 bears against the inner wall 8 of the cylinder 2, which has cooling fins 9 on the outside, over a predetermined rolling region A.
  • the cylinder 2 has a continuous separating slot 10 which extends in the longitudinal direction and in which a separating slide 12 is arranged.
  • the separating slide 12 is by means of a compression spring 14 the ring piston 4 tracked.
  • the isolating slide 12 is shown in the working position "lower dead center" corresponding to a compression ratio of 1: 2.
  • a pressure slot 16 is present in the separating slide 12, to which a pressure valve 18 is assigned in the cylinder.
  • a suction slot 20 can also be seen in the cylinder 2.
  • the annular piston 4 has a constant wall thickness over its entire circumference and is mounted on five rollers 21 to 25 inside.
  • the rollers 21 and 25 are spaced apart from one another such that the annular piston 4 lies flat against the inner wall 8 of the cylinder 2 via the central rolling region A.
  • the center 26 of the annular piston 4 is arranged at a distance 30 from the center 28 of the circular cylinder wall 8, which corresponds to the natural eccentricity e corresponding to half the diameter difference between the cylinder wall and the piston, plus a deformation d.
  • the last-mentioned deformation or enlargement of the normal eccentricity e by d results in the desired flat contact in the rolling area A.
  • the center point 26 rotates on a circular path K around the center point 28.
  • the wall thickness of the The annular piston 4 and the deformation d are adjusted in accordance with the operational requirements.
  • the rollers 21, 25 closest to the unwinding area A are at a considerably greater distance from one another than the rollers 21 to 25 from one another.
  • the drive rollers can be arranged asymmetrically distributed on the rotating body.
  • these drive rollers can be arranged on different radii or can also have different roller diameters in order to achieve reliable support of the annular piston even with large deformations.
  • the diameter of the annular piston is approximately smaller than that of the cylinder with regard to an economical suction volume.
  • FIG. 2 shows a longitudinal section of the compressor according to FIG. 1.
  • Two flange shafts 31 are connected to a drive shaft 42, which in turn are connected by means of connecting bolts 41.
  • the connecting bolts carry three axially spaced drive rollers 23 on which the ring piston 4 is supported by means of roller bearings.
  • the flange shaft 31 on the right in the drawing has a central bore 37, through which cooling air can be blown in.
  • the cooling air exits through the holes 68 in the housing cover 66. According to the invention, there is internal cooling and heat build-up inside the compressor is avoided. Cylinder and ring piston have approximately the same temperature, so that changes in length are kept within narrow limits.
  • the axial sealing gap between the annular piston 4 and the housing cover 66 can be kept very small according to the invention.
  • the eccentricity and the deformation are predefined in a structurally simple manner by suitably arranging the holes 43 for the connecting bolts 41 of all the drive rollers, which are provided off-center.
  • Fig. 3 shows schematically the compressor acc. 1, but rotated by 180 ° about the longitudinal axis, with a compression ratio of approx. 1: 7.
  • the resulting gas forces 32 and the spring force 34 of the separating slide 12 deform the annular piston 4 between the drive rollers according to the dash-dotted line 36, as a result of which additional radial forces initially act on the drive rollers in the unloaded suction zone.
  • the annular piston 4 is additionally loaded in terms of strength by considerable bending forces.
  • FIG. 4 schematically shows the compressor according to FIG. 1 when the annular piston 4 is rolled over by the pressure slot 16 and suction slot 20. Because of the gas forces and the spring force 34 acting from the separating slide 12, the annular piston 4 tries to detach itself from the diametrically opposite roller 23. The resulting deformation of the annular piston 4 is indicated by the dashed line 38, with a distance 40 from the roller 23 can be determined.
  • FIG. 5 shows an essential embodiment of the invention, components which correspond in their mode of operation to the embodiment explained above have the same reference numerals and will not be explained further.
  • a drive shaft 42 two axially spaced drive eccentrics 44 are arranged as pressure bodies, only one of which can be seen and which are flattened in the angular range B for the purpose of mass compensation.
  • Longitudinal bores 48 in the drive eccentrics 44 together with the flattened portion 50 provide a good supply of cooling air into the interior of the annular piston 4.
  • the annular piston 4 is rotatable on a needle bearing 52 with an inner bearing ring 54 on the associated eccentric.
  • the course of the wall thickness can be calculated exactly and set in conjunction with the selected preload or deformation d such that the ring piston 4 lifts off in the rolling region A in the event of an impermissibly high delivery pressure; Reliable overload protection is provided.
  • the annular piston 4 is supported on the entire circumference via the individual rollers of the needle bearing 52.
  • the slight deformation d of the needle bearing according to the invention is in the range from 0.2 to 0.7% of the bearing diameter and practically does not impair the kinematic behavior of the needle bearing.
  • the proposed machine can be manufactured inexpensively and is equally suitable for use in a vacuum and in a high pressure range.
  • a compressor manufactured in this way has a delivery capacity of 810 1 / min, for example, with a stroke volume of 0.27 l and a speed of 3,000 rpm; the inside diameter of the cylinder is 125 mm and the outside diameter of the ring piston is 113.4 mm.
  • Fig. 7 shows a bearing ring 54 with a uniform wall thickness over the circumference. This corresponds to an annular piston according to the known embodiment from FIG. 1. Due to the deflection in the rolling area, there are much higher stress peaks compared to the special embodiment of FIG. 6. These stress peaks are indicated by the arrows 60 and they cause knocking and material fatigue when the separating slot is rolled over.
  • Fig. 8 shows a further essential embodiment, the annular piston 4 is resiliently supported on resilient elements 93.
  • These elements 93 are designed as spiral spokes of a wheel with an outer ring 92 and an inner ring 94.
  • the inner ring 94 is supported on a roller bearing 64.
  • the roller bearing 64 is in no way deformed in this embodiment and can therefore be completely sealed according to the invention.
  • the outer ring 92 is also comparatively thin-walled, so that the ring body 4 in the rolling area A lies flat against the inner wall, with a sufficiently uniform force distribution also being provided.
  • the wheel is expediently made from a single piece, which results in advantages in manufacture and assembly.
  • FIG. 9 corresponds in principle to that of FIG. 8, but now individual elements 93 designed as curved leaf springs are provided for supporting the annular piston 4.
  • the inner ring 94 is not deformed, so that conventional, sealed roller bearings or the like can also be used here.
  • FIG. 10 shows a longitudinal section through an embodiment with a floating annular piston 4.
  • two drive rollers 62, 63 are arranged axially spaced apart, the rolling movement of which is generated by means of eccentrics 44, 45 fastened on the drive shaft 42.
  • the power transmission to the associated drive roller 62 takes place in each case via a commercially available roller bearing 64.
  • these roller bearings are not deformed and can also be easily sealed at the side, which is particularly advantageous for vacuum applications.
  • the drive shaft 42 is laterally mounted in a housing cover 66, wherein cooling air can be blown through bores 68. Such internal cooling avoids heat accumulation inside the compressor with all the disadvantages associated with it.
  • Conventional reed valves 70 are located above the separating slide 12, via which the compressed medium is pushed out.
  • Fig. 11 shows a cross section of the machine acc. 10.
  • the cylinder housing 2 has a number of longitudinal channels 72 for coolant, for example air or water.
  • An integrated valve 74 is provided in the separating slide 12, which will be explained further below with reference to FIGS. 24, 25. Throttling and deflection losses are avoided by the tongue valve integrated according to the invention.
  • Inside the ring piston 4, the one drive roller 62 can be seen completely in an axial view. In the axial direction behind it, that is, behind the plane of the drawing, is the second drive roller 63, of which only a small, approximately sickle-shaped area can be seen, which is identified by crossed lines for the purpose of highlighting.
  • the annular piston 4 is pressed flat against the wall 8 of the cylinder housing 2 by the laterally offset arrangement of the drive roller 62 in the roll-off angle region 56.
  • the eccentric has 44 longitudinal channels 76 for internal cooling.
  • FIG. 12 schematically shows the annular piston 4 in the unstressed state, the cylindrical inner wall 8 of the cylinder housing 2 being touched linearly in the region of the Y axis, here on the left in the figure.
  • the eccentricity e corresponds to half the difference between the inner diameter of the cylinder 2 and the outer diameter of the ring piston 4.
  • the two drive rollers 62, 63 which have a predetermined smaller diameter than the inner bore of the ring piston 4, are arranged such that in the area of the X -Axis is given to the inner bore of the annular piston 4.
  • the two eccentrics 44 are each pivoted by an angle b with respect to the Y axis.
  • the outer diameter of the drive rollers 62, 63 is at least 0.5% smaller than the inner diameter of the annular piston 4.
  • a spring travel f is thus present between the annular piston 4 and the drive rollers.
  • the outer diameter of the drive rollers 62, 63 are in the range between 5 to 0.5%, be preferably 2%, smaller than the inner diameter of the annular piston 4; sufficient travel is guaranteed. In the context of the invention, at least one such pair of drive rollers 62, 63 is required. According to the required axial length of the annular piston, several such pairs of drive rollers can also be arranged, expediently axially evenly spaced on the drive shaft.
  • FIG. 13 shows the enlargement of the natural eccentricity e by an amount d, namely in the direction of the Y axis to the left.
  • the annular piston 4 resiliently hugs the wall 8 of the cylinder housing 2 in the rolling region A and wraps around the drive rollers 62, 63 in an enlarged angular range C.
  • the annular piston 4 lifts by the amount d + f the drive rollers 62, 63.
  • Fig. 15 corresponds essentially to Fig. 13, wherein in addition to the drive rollers 62, 63, a roller 80 is arranged inside the annular piston 4 on the drive shaft, namely between the two axially spaced pairs of the drive rollers 62, 63.
  • This roller 80 protrudes beyond the drive rollers 62, 63 on the Y axis, diametrically opposite the rolling region, and only the free path s is available.
  • the deflection of the annular piston 4 is limited to the free path. At high pressures, a secure support of the Ring piston 4 reached.
  • the maximum deformation forces or the stress distribution are predetermined by the difference in diameter of drive rollers 62, 63 and inner diameter of the annular piston 4.
  • this small diameter difference is between 0.8 to 3%; slight relative movements between the drive rollers and the ring piston are thus achieved.
  • the correspondingly low strength stress of the annular piston 4 also enables the use of inexpensive materials.
  • An exact prediction of the resilient contact pressure is made possible within the scope of the invention by correspondingly specifying the wall thickness of the annular piston 4 and the diameter difference mentioned.
  • FIG. 17 shows an embodiment of the machine with a floating ring piston 4, partly in a longitudinal section.
  • the drive rollers 62, 63 are now mounted directly on the two drive eccentrics 44 by means of slide bearings.
  • the annular piston 4 is sealed laterally on the housing cover 66 by means of elastic sealing elements 84 which are guided by springs 82.
  • the machine according to the invention with oil-flooded plain bearings is in turn designed correspondingly to the right of the center line 86, from which the drive shaft 42 can also be driven.
  • oil is supplied through an axial bore 88 and guided via radial bores 90 to the radial bearings of the drive rollers 62, 63 mentioned for the purpose of lubrication.
  • the oil enters between the drive rollers 62, 63 into the interior of the annular piston 4 and can be discharged from there through bores 68 of the housing cover 66.
  • FIG. 19 shows the drive roller 62, while in FIG. 19 a section along section line A in accordance with FIG. Fig. 18 is shown.
  • curved guide vanes 96 are arranged between the outer ring 92 and the inner ring 94, via which air is sucked into the interior of the annular piston for cooling.
  • the 20 and 21 show, in a view and in an axial section, the two drive eccentrics 44, 46.
  • the drive eccentric 44 has a feather key groove 98 for attachment to the drive shaft.
  • the drive eccentric 46 contains a longitudinal groove 100, through which a screw 102 is guided, which engages in a thread 104 of the eccentric 44.
  • the two eccentrics 44, 46 can thus be rotated against each other for the purpose of tolerance compensation and for setting the pretension explained above, the mutual bracing and locking being carried out by means of the screw 102.
  • FIG. 22 and 23 show in a cross section or longitudinal section an embodiment of the machine which in its kinematic principle of the gem. 5 corresponds, but now the annular piston 4 is arranged radially on the outside with respect to the now piston-like housing 2.
  • This embodiment is particularly suitable for belt drive or direct flange mounting on an electric drive motor.
  • the housing 2 has cooling bores 106, and the annular piston 4 is supported directly in a drive ring 110 via a needle bearing 108.
  • the drive ring 110 is offset by an amount e + d from the housing 2, so that the annular piston 4 clings to the outer surface of the circular housing 2 in the rolling region A.
  • the drive ring 110 has A recess 112 over an angular range C, and there the outer bearing ring 114 has a reduced wall thickness.
  • the corresponding kinematic and voltage requirements are thus achieved as in the embodiment according to FIG. 5.
  • the isolating slide 12 is guided in the housing 2 and can be moved towards the center.
  • the spring 14 and the pressure valves 18 are arranged in a central bore 116.
  • the pumped medium under pressure is discharged through the bore 118 in a cover plate 120.
  • the suction takes place via a bore 122 adjacent to the slide 12.
  • the drive ring 110 is mounted on both sides with respect to the cover disks 120 by means of roller bearings 124.
  • the drive ring 110 contains an annular groove 126 for a belt drive and further cooling fins 128 which are arranged eccentrically to the bearing ring for the purpose of mass balance. According to the invention, cooling takes place by convection of the rapidly rotating outer ring 110, warm air being sucked out of the outer surface of the annular piston 4 via central bores in the outer ring.
  • the fixed housing 2 can be additionally cooled by air or water by means of the cooling bores 106. When water cooling, the cover plates 120 are covered by means of closed plates 130.
  • the advantage of this design is the compact circular structure, the self-acting cooling of the drive ring 110, the combined cooling option by means of air and water and, moreover, the appropriate mass balance by changing the wall thickness of the drive ring.
  • the separating slide 12 has at least one radially continuous slot 132, a number of such slots 132 being expediently axially spaced apart.
  • the resilient valve plate 134 is guided laterally in said slot 132.
  • the valve plates 134 extend over a predetermined large length of preferably approximately 80% of the isolating slide length, so that a large outlet cross section is provided when the valve is open.
  • a valve plate is pressed against the sealing surface by an end pressure Pe, a correspondingly increased pressure must be present to open the valve.
  • the opening pressure P 1 is equal to Pe multiplied by the square of d 2 divided by d 1.
  • the pressure in the cylinder must rise to 9.7 bar in order to raise the valve. The resulting pressure peak leads to an increase in the temperature of the gas and additional storage and material loads.
  • valve plate 134 which is essential to the invention, there is only a quasi-linear contact in the separating slot 10. Furthermore, this substantial curvature reduces the outflow losses, and at the same time the valve lift can also be limited.
  • the preferred embodiment set forth with reference to FIGS. 21 and 22 is particularly suitable for vacuum and low pressure operation as well as for oil flooding.
  • the curved valve plate 134 which reciprocates on the inner surface of the valve slot 10 prevents backflow in the throttle gap and is insensitive to the conveyance of gases containing steam and liquids and furthermore allows the cooling oil to escape undisturbed in the case of an oil-flooded machine.
  • 26 to 28 show an embodiment of an integrated valve which is particularly suitable for high pressures and dry-running machines.
  • the separating slide 12 consists of two parts 136, 138, between which valve plates 140 are clamped.
  • Friction losses in the separating slot of the housing are not insignificantly reduced.
  • the medium enters the internal valve chambers through extensive slots 142 and flows out at the top 144 of the slide valve. Due to the large-area inlet slots 142, low gas speeds occur and the deflection losses are kept low.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
  • Compressor (AREA)
EP84114423A 1983-12-05 1984-11-29 Machine en particulier pour la compression et le déplacement des fluides Expired - Lifetime EP0147654B1 (fr)

Priority Applications (1)

Application Number Priority Date Filing Date Title
JP26631285A JPS61192884A (ja) 1984-11-29 1985-11-28 流体を圧縮し搬送するための作動機械

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE3343908 1983-12-05
DE19833343908 DE3343908A1 (de) 1983-12-05 1983-12-05 Maschine, insbesondere arbeitsmaschine zum verdichten und foerdern von fluiden aller art

Publications (3)

Publication Number Publication Date
EP0147654A2 true EP0147654A2 (fr) 1985-07-10
EP0147654A3 EP0147654A3 (en) 1985-08-14
EP0147654B1 EP0147654B1 (fr) 1990-05-02

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Application Number Title Priority Date Filing Date
EP84114423A Expired - Lifetime EP0147654B1 (fr) 1983-12-05 1984-11-29 Machine en particulier pour la compression et le déplacement des fluides

Country Status (4)

Country Link
US (1) US4580957A (fr)
EP (1) EP0147654B1 (fr)
JP (1) JPS60173384A (fr)
DE (2) DE3343908A1 (fr)

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE3542776A1 (de) * 1985-12-04 1987-07-23 Kurt G Dipl Ing Fickelscher Roll-ring-maschine zum verdichten und foerdern von fluiden
DE202016101907U1 (de) * 2016-04-11 2017-07-12 Ulrich Gmbh & Co. Kg Schlauchpumpe

Families Citing this family (19)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE3343908A1 (de) * 1983-12-05 1984-06-28 Kurt G. Ing.(grad.) 6710 Frankenthal Fickelscher Maschine, insbesondere arbeitsmaschine zum verdichten und foerdern von fluiden aller art
DE3530436A1 (de) * 1985-08-26 1987-02-26 Kraftwerk Union Ag Rollkolbenverdichter
DE3611326A1 (de) * 1986-04-04 1987-10-15 Siemens Ag Rollkolbenverdichter
DE3727697A1 (de) * 1987-03-23 1989-03-02 Siemens Ag Rollkolbenverdichter
DE3821168C1 (en) * 1988-06-23 1989-11-30 Kurt G. Dipl.-Ing. 6710 Frankenthal De Fickelscher Bearing arrangement
GB2221257A (en) * 1988-07-27 1990-01-31 Liou Yan Ming Compressor with a rotor mounted eccentrically on a shaft by a bearing
US4975031A (en) * 1989-01-09 1990-12-04 General Electric Company Rotary compressor with compliant impact surfaces
EP0426887B1 (fr) * 1989-11-07 1992-11-25 Werner Riester GmbH & Co. KG Armaturen- und Maschinenantriebe Dispositif de roulements
US5061227A (en) * 1989-11-13 1991-10-29 Renk Aktiengesellschaft Bearing system for wave generator drive
US5116208A (en) * 1990-08-20 1992-05-26 Sundstrand Corporation Seal rings for the roller on a rotary compressor
CN1034604C (zh) * 1993-01-06 1997-04-16 三星电子株式会社 旋转压缩机的压缩气体排出装置
US7832257B2 (en) * 2007-10-05 2010-11-16 Halliburton Energy Services Inc. Determining fluid rheological properties
CN102734165A (zh) * 2011-04-11 2012-10-17 广东美芝制冷设备有限公司 容量控制式旋转压缩机
CN102748287A (zh) * 2011-04-19 2012-10-24 广东美芝制冷设备有限公司 旋转式压缩机
CN102767518A (zh) * 2011-05-03 2012-11-07 广东美芝制冷设备有限公司 旋转压缩机
RU2605269C2 (ru) * 2014-10-14 2016-12-20 Григорий Иванович Поздняков Насос гидравлический пластинчатый
CN106122013B (zh) * 2016-07-29 2018-09-28 珠海格力节能环保制冷技术研究中心有限公司 一种滚动转子压缩机
CN107489616A (zh) * 2017-08-24 2017-12-19 广州市德善数控科技有限公司 一种压缩机滑片的控制机构
WO2019180900A1 (fr) * 2018-03-23 2019-09-26 三菱電機株式会社 Compresseur hermétique

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FR981898A (fr) * 1949-03-02 1951-05-30 Pompe volumétrique à piston à mouvement hypocycloïdal
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DE2541835A1 (de) * 1975-09-19 1977-03-24 Standard Oil Co Ohio Drehkolbenmotor
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DE7906528U1 (de) * 1979-03-09 1979-12-13 P.A. Rentrop, Hubbert & Wagner Fahrzeugausstattungen Gmbh Maschine, insbesondere arbeitsmaschine zum verdichten und foerdern von fluiden aller art

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US1692639A (en) * 1926-12-11 1928-11-20 Henry L Elsner Pump
DE669091C (de) * 1934-03-23 1938-12-16 Carl Steinmann Drehkolbenverdichter
FR981898A (fr) * 1949-03-02 1951-05-30 Pompe volumétrique à piston à mouvement hypocycloïdal
US2922378A (en) * 1955-06-20 1960-01-26 Richard E Pabst Rotary pump
US2992769A (en) * 1957-03-20 1961-07-18 Petty Lab Inc Rotary fluid compressors
DE2541835A1 (de) * 1975-09-19 1977-03-24 Standard Oil Co Ohio Drehkolbenmotor
DE2911655A1 (de) * 1979-03-24 1980-10-02 Erich Becker Rollkolbenpumpe
DE3343908A1 (de) * 1983-12-05 1984-06-28 Kurt G. Ing.(grad.) 6710 Frankenthal Fickelscher Maschine, insbesondere arbeitsmaschine zum verdichten und foerdern von fluiden aller art

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DE3542776A1 (de) * 1985-12-04 1987-07-23 Kurt G Dipl Ing Fickelscher Roll-ring-maschine zum verdichten und foerdern von fluiden
DE202016101907U1 (de) * 2016-04-11 2017-07-12 Ulrich Gmbh & Co. Kg Schlauchpumpe

Also Published As

Publication number Publication date
DE3343908A1 (de) 1984-06-28
EP0147654B1 (fr) 1990-05-02
JPS60173384A (ja) 1985-09-06
DE3482128D1 (de) 1990-06-07
EP0147654A3 (en) 1985-08-14
US4580957A (en) 1986-04-08

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