US20100021323A1

US20100021323A1 - Compressor comprising a compressed gas-assisted piston

Info

Publication number: US20100021323A1
Application number: US12/513,702
Authority: US
Inventors: Jan-Grigor Schubert
Original assignee: BSH Bosch und Siemens Hausgeraete GmbH
Current assignee: BSH Hausgeraete GmbH
Priority date: 2006-11-07
Filing date: 2007-10-31
Publication date: 2010-01-28
Also published as: DE102006052430A1; RU2009119391A; CN101535644A; WO2008055826A1; EP2092195A1; RU2432497C2

Abstract

A compressor includes a cylinder and a piston moveable in an oscillating manner in the cylinder and having transverse play in relation to a movement direction. An end face of the piston delimits a compression chamber in the cylinder. A diameter of the piston reduces toward the end face.

Description

The present invention relates to a compressor having a cylinder in which a piston is movably held by a gas bearing without contact with the cylinder wall.
Such a compressor is known, for example, from U.S. Pat. No. 6,575,716 A1. In this conventional compressor, there is formed in the inner wall of the cylinder a circumferential groove which is supplied with compressed gas via a bore intersecting the cylinder wall. The compressed gas is distributed in the circumferential groove around the entire circumference of the piston and propagates from the groove in the axial direction through a narrow gap between piston and cylinder wall, thereby retaining the piston over its entire circumference without contact with the cylinder wall. When a radial force is exerted on the piston and deflects the latter from its equilibrium position, the gas is not only compressed but also partially expelled on one side of the piston circumference, said expulsion possibly being due, among other things, to the gas escaping back into the groove. While the compressed gas exerts on the piston a restoring force in the direction of the equilibrium position, the ejected gas cannot do so. Because of the escape possibility, the stiffness of the bearing against radial deflection is not overly great.
In order to improve the stiffness of the bearing, it has been proposed to introduce the compressed gas via radial bores with very narrow cross-section into the gap between cylinder wall and piston. Because of the narrow cross-section of the bores, it is only possible for a small amount of gas to flow back when the piston is deflected. A higher radial stiffness of the bearing can therefore be achieved for the same gas throughput.
In order to limit the reverse flow of the compressed gas, the supply bores must have a very small diameter of a similar order of magnitude to the gap width between cylinder wall and piston. In practice this means that the diameter of the supply bores must be no more than a few 10 μm. The production of such narrow bores requires complex processing techniques such as laser ablation, spark erosion or the like. The supply bores can only be produced singly using these techniques, which makes production time-consuming and costly. In addition, the material thickness in which such narrow bores can be produced is limited to a few hundred μm. A workpiece with such a thin wall is easily deformable, so that it is difficult to ensure that the cylinder wall has the dimensional accuracy and stability required for an effective gas bearing.
The object of the present invention is to specify a compressor with a gas bearing supported piston, said compressor being implementable with low cost/complexity and enabling the piston to be supported with good radial stiffness at low compressed gas throughput.
This object is achieved by a compressor having a cylinder and a piston which can be displaced in an oscillating manner in said cylinder and which has transverse play in relation to the direction of movement, an end face of the piston delimiting a compression chamber in the cylinder, characterized in that the piston has a diameter which reduces toward the end face. As the piston's shape tapers toward the end face, some of the gas compressed in the compression chamber by the piston movement is forced into the gap between piston and cylinder wall, and it is the flow of gas escaping from the compression chamber between piston and cylinder wall that provides the gas bearing effect.
In order to ensure precise, rolling-motion-free guiding of the piston, the latter preferably has a guide section of constant diameter in addition to a section having the diameter that reduces toward the end face adjacent to the compression chamber.
In order to drive the compressed gas out of the compression chamber into the gap in a low turbulence manner, the increase in the diameter is preferably constant. It is particularly preferable that the rate of change of the diameter in the direction of the axis is at its maximum directly at the end face and reduces with increasing distance from the end face.
In the simplest case, the inner wall of the cylinder can be completely devoid of supply bores for feeding compressed gas into the gap between inner wall and piston.
In this case, the flow of gas through the gap stops at least once during each oscillation of the piston, so that contact may occur between the piston and the cylinder wall at this time. In order to limit rubbing wear between the piston and the cylinder wall particularly but not exclusively in this case, the piston and/or the inner wall of the cylinder can be provided with a hard coating. The coating can consist of a carbide, e.g. tungsten carbide, DLC (diamond-like carbon) or the like.
In order also to maintain the gas bearing's effectiveness even when the compression chamber is at its maximum extent, supply bores for feeding in the compressed gas can be disposed in an inner wall of the cylinder such that they pressurize a section of the piston facing the compression chamber at the piston movement reversal point at which the compression chamber expansion is at its maximum. This design permits at least a considerable reduction in the number of supply bores compared to a conventional compressor in which the bearing effect is maintained exclusively via compressed gas fed in externally through supply bores.
In order to minimize radial forces on the piston which could force it against the inner wall of the cylinder, the compressor expediently has a drive unit which executes a purely linear motion. Such a drive unit can in particular comprise a magnetic armature which is coupled to the piston and can be driven in a magnetic alternating field, parallel to the direction of movement of the piston.

Further features and advantages of the invention will emerge from the following description of exemplary embodiments and with reference to the accompanying drawings in which:

FIG. 1 shows a schematic section through the piston and cylinder of a compressor according to a first embodiment of the invention;

FIG. 2 shows a schematic section through the drive unit of the compressor;

FIG. 3 shows a section analogous to FIG. 1 according to a second embodiment of the invention;

FIG. 4 shows a corresponding section according to a third embodiment of the invention; and

FIG. 5 shows a head-on view of a bushing used in the third embodiment.

The compressor shown in FIG. 1 comprises a cylinder 21 which is essentially composed of a tubular section 22, a head plate 23 covering one end of the tubular section 22, and a cap 24 mounted to a side of the head plate 23 facing away from the tubular section 22. The tubular section 22, head plate 23 and a piston 25 engaging in the tubular section 22 delimit a compression chamber 26. The compression chamber 26 communicates with two chambers 29, 30 formed in the cap 24 via valves 27, 28 shown schematically in the figure which are preferably formed in one piece from the spring steel head plate 23. The valves 27, 28 are check valves which allow gas to flow only from the upper, suction-side chamber 29 into the compression chamber 26 during an outward movement of the piston 25 or from the compression chamber 26 into the lower, pressure-side chamber 30 during an inward movement of the piston 25.
The piston has, facing the head plate 23, a flat end face 31 whose diameter is much less than that of the compression chamber 26. At its edges, the end face 31 makes a continuously curved transition to a circumferential surface 32 facing the inside of the tubular section 22. The circumferential surface 32 can be divided into three sections in the movement direction of the piston 25: a cylindrical central section 33 whose diameter is no more than a few 10 μm smaller than that of the compression chamber 26 so that its movement is guided in the tubular section 22 with little play and in an essentially non-rolling manner, and, on either side of the central section 33, an inner and an outer section 34 and 35 respectively, whose diameter continuously decreases with increasing distance from the central section 33.
The width of the gap 36 between the circumferential surface 32 and the inner surface of the tubular section 22 increases faster than linearly with increasing distance from the central section 33.
The gap 36 funneling out toward the compression chamber 26 in this way favors the ingress of compressed gas from the compression chamber 26 so that, near the central section 33, the flow of gas through the gap 36 which is narrow at this point is much stronger then in the case of a piston of precisely cylindrical form. This gas flow makes it possible to implement a bearing effect corresponding to that of a conventional gas bearing with gas fed into the gap via supply bores.
The gas bearing effect is only briefly interrupted when no pressure difference exists between the compression chamber 26 and the rear of the piston 25. For a compressor in which the entire cylinder 21 is hermetically sealed in the conventional manner and the rear of the piston 25 communicates with the suction-side chamber 29, this may be the case at the piston movement reversal point facing away from the head plate 23.
If there are flow obstructions on the path of the gas from the rear of the piston via the suction-side chamber 29 and its valve 27 into the compression chamber 26, causing the pressure in the compression chamber 26 to fall below the pressure at the rear of the piston during an outward movement of the piston 25—such a flow obstruction can be, in particular, the valve 27 itself—the flow of gas in the gap 36 is interrupted twice in each movement cycle of the piston—shortly before it reaches the reversal point facing away from the head plate 23 and thereafter—resulting in a temporary reversal of the direction of flow of the gas in the gap 36 between piston 25 and tubular section 22. In order also to amplify this gas flow directed into the compression chamber 26 such that it has a bearing effect, the diameter of the piston continuously decreases from the central section 33 toward a rear end face 37 in the outer section 35 also.
In order to minimize rubbing wear due to the contact between piston 25 and tubular section 22 at the time of stoppage of the gas flow in the gap 36, the circumferential surface 32 of the piston 25 at least in its central section 33 and/or the inner surface of the tubular section 22 is provided with a hard, abrasion resistant coating, e.g. of tungsten carbide, DLC or the like.
FIG. 2 schematically illustrates a drive unit which can be used to drive the oscillating movement of the piston 25 via a piston rod 38. The drive unit comprises two E-shaped yokes 1 with three pairwise opposite arms 3, 4, 5. The mutually facing ends of the arms 3, 4, 5 each constitute pole shoes 7 delimiting an air gap 2. Mounted around each of the inner arms 4 is an excitation winding 8. Current is applied to the two excitation windings 8 by a control circuit, the flow direction in the two excitation windings 8 being determined such that the mutually opposite pole shoes 7 of the central arms 4 form unlike magnetic poles. The pole shoes of the outer arms 3 and 5 each constitute unlike poles to the adjacent central arm 4.
In the air gap 2, an armature 10 is suspended in a reversibly movable manner from two springs 11 between an upper and a lower reversal point (or rather a right and left reversal point in the diagram in FIG. 2). The position of the armature 10 at the upper and lower reversal point is shown with solid and broken lines respectively. The springs 11 are in each case leaf springs punched out of a sheet metal piece and having a plurality of zigzag arms 12. The arms 12 of a spring 11 extend as mirror images of one another from a central point of action on the armature 10 to suspension points 13 on a rigid frame (not shown) on which the yokes 1 and the compressor are anchored. This design means that the springs 11 are difficult to deform in the longitudinal direction of the armature 10 and in each direction orthogonal thereto, so that they reversibly guide the armature 10 in its longitudinal direction.
The essentially rod-shaped armature 10 incorporates a four-pole permanent magnet 14 in its central region. Whereas in a relaxed position of the springs 11 in which the arms 12 of each spring 11 lie essentially in the same plane the magnet 14 is placed centrally in the air gap 2 and a boundary line 15 between its left and right poles in FIG. 1 runs centrally through the central arms 4, when a current is applied to the windings 8 the armature 10 is deflected to the left or to the right depending on the direction of the current.
FIG. 3 shows a variant of the inventive compressor which is likewise combinable with the drive unit shown in FIG. 2. The compressor has a head plate 23 with valves 27, 28 and a cap 24 with chambers 29, 30, as described above with reference to FIG. 1. The piston 25 likewise has a structure with a cylindrical central section 33 and inner and outer sections 34, 35 tapering toward the end faces 31 and 37 respectively. In the tubular section 22 a bushing 39 is incorporated which together with the piston 25 and the head plate 23 delimits the compression chamber 26. Between the bushing 39 and the tubular section 22 is an annular cavity 40 which is sealed at its end facing away from the head plate 23 by an O-ring 41 or the like and communicates with the pressure-side chamber 30 via a bore 42 running obliquely through the tubular section 22 and the head plate 23.
Supply bores 43 with a diameter of a few 10 μm intersect the bushing 39. The axial position of the supply bores 43 is selected such that, at the piston movement reversal point facing away from the head plate 23, shown in the Figure by a dashed outline of the piston 25, the supply bores 42 are near the central section 33 of the piston, whereas, at the piston movement reversal point facing the head plate 23, there does not necessarily have to be axial overlapping of the positions of the supply bores 43 and the piston 25. When the piston 25 is close to said reversal point facing the head plate 23, the overpressure in the compression chamber 26 is sufficient to maintain a sufficient gas flow through the gap 36 for the desired bearing effect. When the piston 25 is close to the reversal point facing away from the head plate 23 at which there is no overpressure in the compression chamber 26 to drive a gas stream through the gap 36, the gas bearing effect is maintained by the supply bores 43, so that no contact with the bushing 39 occurs at any phase of the oscillatory movement of the piston 25.
Due to the effect of the valve 28, a continuous overpressure is maintained in the chamber 30 even while the piston 25 is moving away from the head plate 23. This continuous overpressure allows the supply bores 43 to be continuously fed compressed gas. However, it is also conceivable for the transmission characteristics of the bore 42 and the cavity 40 to be optimized such that a pressure surge, which occurs in the chamber 30 whenever the valve 28 opens and fresh compressed gas from the compression chamber 26 flows into the chamber 30, propagates through the bore 42 and the cavity 40 and reaches the supply bores 43 when the piston 25 is in front of said supply bores 43. This enables the amount of compressed gas required for supporting the piston 25 to be reduced still further.
As only a relatively small number of supply bores 43 are required, here too a reduction in manufacturing complexity can be achieved compared to a conventional gas bearing compressor with axially distributed supply bores.
In this embodiment a hard coating as described above can also be provided on the piston 25 and/or the tubular section 22 in order to avoid rubbing wear each time the compressor is started up, when the pressure in the chamber 30 is not yet sufficient to produce the bearing effect at the supply bores 43.
Analogously to FIGS. 1 and 3, FIG. 4 shows a section through a third embodiment of the compressor according to the invention. Once again, the cap 24, the head plate 23 and the piston 25 are identical to those shown in FIG. 1. The interior of the tubular section 22 is merely enlarged at its end facing away from the head plate 23, and into the enlargement a bushing 44 is inserted which abuts a shoulder 48 of the enlargement and whose inner surface is flush with the inner surface of the un-enlarged portion of the tubular section 22. The tubular section 22 and bushing 44 delimit an annular duct 45 which communicates with the pressure-side chamber 30 via a bore 42.
FIG. 5 shows head-on view of the bushing 44. It can be seen that grooves 47 evenly distributed around the circumference are impressed into an end face 46 of the bushing which, in the assembled state, abuts a shoulder of the tubular section 22 delimiting the enlargement. Unlike bore holes, the grooves 47 with a width and a depth of a few 10 μm and virtually any length can be implemented simultaneously with little cost/complexity. Together with the shoulder 48 of the tubular section 22 they delimit supply passages 43 via which compressed gas can flow from the annular duct 45 to the inside of the tubular section 22 and keep the piston 25 supported in the vicinity of its reversal point facing away from the head plate 23.

Claims

1-10. (canceled)

11. A compressor comprising:

a cylinder; and

a piston moveable in an oscillating manner in the cylinder and having transverse play in relation to a movement direction,

wherein an end face of the piston delimits a compression chamber in the cylinder, and

wherein a diameter of the piston reduces toward the end face.

12. The compressor as claimed in claim 11, wherein the piston comprises:

a first section that is adjacent to the compression chamber, wherein the first section includes the reducing diameter toward the end face; and

a guide section having a constant diameter.

13. The compressor as claimed in claim 12, wherein the piston comprises:

a second section that faces away from the compression chamber,

wherein a diameter of the second section reduces toward a rear end face of the piston.

14. The compressor as claimed in claim 11, wherein the diameter of the piston continuously reduces toward the end face.

15. The compressor as claimed in claim 13, wherein one of the diameter of the first section continuously reduces toward the end face of the piston and the diameter of the second section continuously reduces toward the rear end face of the piston.

16. The compressor as claimed in claim 14, wherein a rate of change of the reducing diameter increases from a center of the piston to the end face.

17. The compressor as claimed in claim 15, wherein a rate of change of the reducing diameter increases from a center of the piston to the rear end face.

18. The compressor as claimed in claim 11, wherein the cylinder comprises:

an inner wall which is devoid of supply bores for feeding compressed gas into a gap between the inner wall and the piston.

19. The compressor as claimed in claim 11, wherein the cylinder comprises:

an inner wall having supply bores for feeding compressed gas such that a section of the piston facing the compression chamber is pressurized at a piston movement reversal point at which the compression chamber is at a maximum extent.

20. The compressor as claimed in claim 11, wherein the cylinder comprises: an inner wall, and

wherein one of the piston and the inner wall of the cylinder includes a hard coating.

21. The compressor as claimed in claim 20, wherein the hard coating includes a carbide.

22. The compressor as claimed in claim 11, wherein the piston is coupled to a magnetic armature that is driven in a magnetic alternating field parallel to the movement direction of the piston.

23. A compressor having a gas bearing, the compressor comprising:

a cylinder having an inner wall; and

a piston moveable in an oscillating manner in the cylinder in a movement direction that is parallel to the inner wall of the cylinder, the piston having transverse play in relation to the movement direction and being configured to be supported in the cylinder by a gas bearing separating the piston from the inner wall of the cylinder,

wherein a shape of the piston tapers toward the end face such that gas compressed in the compression chamber by movement of the piston is forced into a gap between the piston and the inner wall of the cylinder.