US20100310394A1

US20100310394A1 - Compressor unit

Info

Publication number: US20100310394A1
Application number: US12/864,250
Authority: US
Inventors: Marco Giacchi; Bernd Gromoll
Original assignee: BSH Bosch und Siemens Hausgeraete GmbH
Current assignee: BSH Hausgeraete GmbH
Priority date: 2008-02-06
Filing date: 2009-01-28
Publication date: 2010-12-09
Also published as: WO2009098157A1; CN101939542A; RU2010135408A; DE102008007661A1; EP2240692A1

Abstract

A compressor unit having a cylinder and a piston, wherein the cylinder and the piston delimit a compressor chamber. The compressor unit also includes a capsule that surrounds at least the cylinder and the piston and a linear drive to drive a relative movement of the cylinder, the piston and the capsule. The compressor chamber has an outlet opening that opens into an interior of the capsule.

Description

The present invention relates to a compressor unit comprising a cylinder and a piston that delimit a compressor chamber, a linear drive for driving a relative movement of the cylinder and piston and a capsule that surrounds at least the cylinder and the piston.
Such a capsule is conventionally used as a reservoir for gas to be compressed, from which the gas is sucked into the compressor chamber, and a pipeline for drawing off the compressed gas is guided out of the compressor chamber through the interior.
One problem with these conventional compressor units is that the gas-filled interior and the capsule surrounding the same not only absorb operating noise from the compressor, but also prevent the release of heat into the environment. Heat which is generated during operation of the compressor by means of adiabatic compression of the gas in the compressor chamber, partly passes over into the cylinder and piston and finally heats the gas in the interior around the cylinder and piston. This heating process reduces the quantity of gas which is sucked in and compressed in each movement cycle of the piston, thereby negatively affecting the degree of efficiency of the compressor.
A further problem is associated with the design of the conventional linear compressors. These generally include a permanent magnet armature which can be linearly moved to and fro in an air gap of an electromagnet. These linear drives are characterized in that the driving force acting on the armature can be directly transferred to the piston without interconnected levers or suchlike and consequently practically without any frictional losses, but, contrary to a conventional rotary drive, the amplitude of the piston movement is not predefined in a constructionally-specific fashion, but can instead be influenced by the intensity of the magnetic field acting on the armature. To achieve a high degree of efficiency of the compressor, the compressor chamber at the upper dead center of the piston movement should be as small as possible, but it should also be prevented that the cylinder and piston strike one another at the dead center in order to keep the operating noise of the compressor and the stresses within a limit.
The object of the present invention is to specify a compressor unit of the type cited in the introduction, which enables an efficient dissipation of the heat which is released in the compressor chamber and/or despite a high degree of efficiency prevents excessive stress on the material or noise development as a result of the cylinder and piston striking one another.
The object is achieved by an outlet opening of the compressor chamber opening into an interior of the capsule instead of an intake opening, in the case of a compressor unit comprising a cylinder and a piston, which delimit a compressor chamber, a linear drive for driving a relative movement of the cylinder and piston and a capsule, which at least surrounds the cylinder and the piston. As a result, if the compressor unit is operated, the interior of the capsule is under high pressure. Although the compressed gas ejected by the compressor chamber is significantly warmer than the low pressure gas previously ingested, so that the cylinder and piston, if they essentially release their heat into the environment only by way of the gas in the interior, cannot be colder than this gas during stationary operation, its heat conductivity as a result of its high density is greater by a multiple than that of the uncompressed gas, so that an overheating of the cylinder and piston can consequently be reliably prevented.
A further effect which results from the presence of the compressed gas in the capsule surrounding the cylinder and piston is that the high pressure of the compressed gas acts on the rear of the piston. The linear drive 1 must therefore essentially only provide driving power in an expansion phase of the compressor chamber, if the piston is withdrawn from the cylinder counter to the pressure of the compressed gas in the interior. The counter movement of the piston requires hardly any external drive force, since the pressure of the gas in the interior is essentially sufficient to repel the piston. If the linear drive were not to provide any driving power during the compression phase, the piston would then come to a standstill shortly before the upper dead center, if the pressures in the compression chamber and in the interior of the capsule become the same. A small amount of operating energy of the linear drive during the compression phase is sufficient to overcome this pressure equalization position and to expel the contents of the compressor chamber. The energy with which the cylinder and piston could strike one another in the case of pressure fluctuations lies in the order of magnitude of this drive energy and can therefore be kept significantly lower than in the case of a compressor, in which the linear drive in the compression phase must work against the pressure developing in the compressor chamber.
The piston in the cylinder is preferably gas-pressure mounted. By comparison with oil lubrication, the gas pressure mounting is advantageous in that it enables a practically frictionless piston movement. The heat dissipation from the piston via a gas thrust bearing is less efficient than via an oil film, but this is not critical in the present case since the piston is able to output sufficient heat via the high pressure gas of the interior.
To feed the gas thrust bearing, bores are preferably provided in a casing of the cylinder, which connect a gap between the casing and a lateral surface of the piston to the interior.
Furthermore, channels for supplying pressurized gas from the compressor chamber into the gap can extend between a face surface and the lateral face of the piston. These channels facilitate maintenance of the gas thrust bearing in the vicinity of the upper dead center, if, as a result of the pressure equalization and/or the overpressure in the compressor chamber, the inflow of gas from the interior via the bores of the casing ceases.
The channels can be embodied as bores or as open grooves.
The outlet opening of the compressor chamber can be formed in the piston. An entire face surface of the cylinder is thus available in order to accommodate there an inlet valve with a large cross-sectional surface and accordingly minimal drop in pressure.
A suction gas line is preferably routed through the interior to the compressor chamber in order to supply the suction gas rapidly and with minimal heating by means of the pressurized gas from the interior of the compressor chamber.

Further features and advantages of the invention result from the subsequent description of exemplary embodiments with reference to the appended Figures, in which;

FIG. 1 shows a schematic section of an inventive compressor unit;

FIG. 2 shows an enlarged detail of the compressor unit according to a first embodiment;

FIG. 3 shows a top view onto the face surface of the piston of the compressor unit according to the first embodiment;

FIG. 4 shows the detail in FIG. 2 in accordance with a second embodiment, and

FIG. 5 shows a top view onto the face surface of the piston in accordance with the second embodiment.

The linear compressor unit shown in FIG. 1 has a linear drive 1 with a permanent magnet armature suspended so as to be oscillatable in a gap 2 between two opposite electromagnets 3. The electromagnets 3 each have an E-shaped brace with windings surrounding a central arm of the brace. The armature 4 is excited by an alternating current applied to the electromagnets 3 by a supply circuit (not shown) to an oscillation movement in the longitudinal direction of the gap 2 which is controlled by the return springs 23.
A compressor includes a cylinder 7 and a piston 6 which can be moved in the cylinder 7. The piston 6 is coupled to the armature 4 by way of a piston rod 5. The cylinder 7 is fixedly connected to the electromagnet 3 by way of a frame part 24, with which the return springs 23 also engage. The design comprising linear drive 1 and compressor is hermetically enclosed in a capsule 8 and suspended so as to be oscillatable by way of springs (not shown) which engage with the frame part 24 and the capsule 8.
An elastic pipeline 9 extends through a wall of the capsule 8 to a prechamber 10 of the cylinder. The prechamber 10 is separated from a compressor chamber 12 delimited by the cylinder 7 and the piston 6 by a non-return valve 11. A further non-return valve 13 is arranged opposite the non-return valve in a face surface 14 of the piston 6. This is embodied in the present case in the manner of a circular truncated cone and forms the valve seat with its lateral surface in a passage of the piston base. The smaller base surface of the circular truncated cone projects opposite the base of the piston 6 and forms a stop surface.
To drive the oscillation movement of the armature 4 and of the piston 6 effectively, the frequency of the alternating current is attuned to the resonance frequency of the oscillating system comprising linear drive 1, compressor and return springs 23. The amplitude of the oscillation movement is dependent on the electrical power fed by the supply circuit into the electromagnets 3. This may be different in the positive and negative half wave of the alternating current, in particular, it may be larger in the half wave driving an expansion movement of the compressor, than in the half wave driving a compression movement.
If the linear drive 1 drives the piston 6 to an oscillation movement, in an expansion phase of the compression chamber 12, low pressure gas is sucked into the compressor chamber 12 via the pipeline 9 and the prechamber 10. If towards the end of a compression phase of the compressor the pressure in the compressor chamber 12 exceeds the pressure in the interior 16 sufficiently in order to overcome a closing force of the non-return valve 13, the non-return valve 13 opens and the compressed gas escapes into the interior 16 of the capsule 8. In variation, provision can also be made, if the non-return valve 13 projects relative to the base of the piston 6, for said non-return valve to be opened by striking the overhang on the front wall of the cylinder. During stationary operation, the pressure in the interior 16 is only marginally lower than the maximum pressure achieved in the compressor chamber 12, so that the non-return valve 13 only opens briefly before the top dead center is reached. The linear drive 1 therefore operates during the overall expansion phase of the compressor chamber 12 against the pressure of the interior 16, whereas in a compression phase, the pressure in the interior 16 is almost sufficient to compress the gas in the compressor chamber 12. The power which the supply circuit feeds into the electromagnets 3 can therefore be significantly less in a half wave driving the compressor than in a half wave driving the expansion. No complicated monitoring of the piston movement is thus necessary in order to ensure that the piston 6 does not strike the face surface of the cylinder 7, or at least does not do so with excessive force.
FIG. 2 shows a significantly enlarged detailed representation of the compressor in FIG. 1. A fragment of a casing 17 of the cylinder 7 and a lateral surface 18 and the front surface 14 of the piston are apparent. Numerous narrow bores 19 extend through the casing 17, by means of which, provided the pressure in the compressor chamber 12 is lower than in the interior 16, compressed gas flows out of the interior 16 into the compressor chamber 12 and/or into a gap 20 between the casing 17 and the lateral surface 18, thereby forming a pressurized gas cushion.
Shortly before the piston 6 reaches its upper dead center during the course of a compression movement of the piston 6, a pressure equalization results between the compressor chamber 12 and the interior 16 and the gas flow through the bores 19 is disrupted. Numerous bores 21 extending diagonally from the face surface 14 toward the lateral surface 18 of the piston 6 convey the flow of pressurized gas out of the compressor chamber 12 into the gap 20, if the piston 6 approaches the top dead center by way of the pressure equalization position. The flow direction of the gas in the gap 20 consequently reverses during the piston movement, while over a large part of the piston path, from the lower dead center to the pressure equalization position, the pressure in the compressor chamber 12 is lower than in the interior 16 and gas flows through the bores 19 and the gap 20 into the compressor chamber 12, the gas flow in the vicinity of the upper dead center proceeds from the compressor chamber 12 to the interior 16. The piston 6 is thus effectively gas pressure-mounted on both reversal points, if its speed is zero and the duration is accordingly high, and it is only if the piston passes through the pressure equalization position that the gas thrust bearing briefly breaks down. As the piston is moved at this point, the time at which the bearing effect is interrupted is short and the risk of the piston 6 sliding into the cylinder 7 before the bearing effect is used again is minimal
FIG. 3 shows a top view of the face surface 14 of the piston 6 with the input openings of the bores 21 formed thereon. Dashed lines illustrate the radial course of the bores 21 inside the piston 6 toward the lateral surface 18.
FIGS. 4 and 5 each show views similar to FIGS. 2 and 3 in accordance with a second embodiment of the invention. The bores 19 of the piston 6 are replaced here by grooves 22 which are cut into the face and lateral surfaces of the piston at an angle from the outside. The function of the grooves 22 is the same as that of the bores 19 in the first embodiment. The grooves differ by comparison with the bores in that they are easier to manufacture. By contrast, the bores have the smaller dead volume, so that a somewhat higher degree of efficiency can be achieved with the first embodiment than with the second.

Claims

1-16. (canceled)

17. A compressor unit, comprising:

a cylinder;

a piston, the cylinder and the piston delimiting a compressor chamber;

a capsule surrounding at least the cylinder and the piston; and

a linear drive to drive a relative movement of the cylinder, the piston and the capsule;

wherein the compressor chamber has an outlet opening that opens into an interior of the capsule.

18. The compressor unit of claim 17, wherein the piston is gas pressure-mounted in the cylinder.

19. The compressor unit of claim 18, wherein the piston has a lateral surface, and wherein the cylinder has a casing with bores that connect a gap between the casing and the lateral surface with the interior of the capsule.

20. The compressor unit of claim 19, wherein channels to supply pressurized gas from the compressor chamber into the gap extend between a front face and the lateral surface of the piston.

21. The compressor unit of claim 20, wherein the channels are embodied as bores.

22. The compressor unit of claim 20, wherein the channels are embodied as grooves.

23. The compressor unit of claim 17, wherein the outlet opening is formed in the piston.

24. The compressor unit of claim 23, wherein the outlet opening is formed on a base of the piston.

25. The compressor unit of claim 17, further comprising a valve to seal the outlet opening.

26. The compressor unit of claim 25, wherein the valve is a non-return valve.

27. The compressor unit of claim 25, wherein the valve is a stop valve, and wherein the stop valve opens towards an interior of the piston.

28. The compressor unit of claim 26, wherein at least one section of the non-return valve protrudes beyond a base of the piston.

29. The compressor unit of claim 28, wherein the at least one section of the non-return valve is formed from a shock-absorbing material.

30. The compressor unit of claim 29, wherein the shock-absorbing material is Viton.

31. The compressor unit of claim 26, wherein the non-return valve is a circular truncated cone having a lateral surface, and wherein the circular truncated cone and the lateral surface form a valve seat in a base of the piston.

32. The compressor unit of claim 31, wherein the circular truncated cone has a smaller base surface, and wherein the circular truncated cone and the smaller base surface protrude beyond the base of the piston.

33. The compressor unit of claim 17, further comprising a suction gas line that is guided through the interior of the capsule to the compressor chamber.