US20080063549A1

US20080063549A1 - Linear Compressor

Info

Publication number: US20080063549A1
Application number: US11/587,767
Authority: US
Inventors: Edoardo Biscaldi; Andrea Contarini; Samuele Da Ros; Michele Libralato; Giovanni Strappazzon; Nicola Trivillin
Original assignee: Elettromeccanica SpA
Current assignee: ACC Compressors SpA
Priority date: 2004-04-29
Filing date: 2005-03-23
Publication date: 2008-03-13
Also published as: CN101014770A; JP2007534882A; EP1740831B1; MXPA06011843A; WO2005106249A1; DE602005020763D1; EP1740831A1; ATE465343T1; BRPI0509952A; ITPN20040029A1; CN100540894C

Abstract

The present invention refers to a linear compressor comprising a piston (9) arranged slidably inside a cylinder (15); in said cylinder (15) the 15 piston (9) defines a first low-pressure chamber (21) and a second high-pressure chamber (22), said piston being further provided with communication means (23) between said first and said second chamber (21, 22), which are controlled by valve means (24) to establish such communication therebetween.

Description

The present invention refers to a linear compressor, in particular intended for use in household-type or industrial-type refrigeration equipment.
In the course of these last few years, owing to an increasingly felt need on the part of refrigeration equipment manufacturers to be able to rely upon the availability of compressors characterized by an ever increasing level of efficiency, there has been a gradual change in the type of compressors used, i.e. manufacturers have gradually moved over from the conventional reciprocating compressors driven by a rotary electric motor to reciprocating compressors of the linear kind, i.e. driven by a linear electric motor as generally consisting of a stator relative to which there is arranged slidably, with a linear reciprocating motion, generated by a magnetic field, a moving element on which there is mounted the piston of the compressor.
Although clearly advantageous from a performance, efficiency and reliability point of view, linear compressors generally present some problems in connection with the ability to actually obtain such favourable characteristics thereof, and these problems may make these compressors rather complex to design and construct. In the first place, unlike conventional compressors, linear compressors do not have any mechanically, i.e. positively determined top and bottom dead centres, so that the need arises here for a special system to be provided to control the position of the piston inside the cylinder, in view of preventing the same pistons from detrimentally knocking against the cylinder head. Such control shall operate to a great accuracy extent in order to ensure that the top dead centre of the piston is situated at an extremely reduced distance from the cylinder head, generally in the order of 0.1 mm, so as to minimize the so-called clearance volume. Any greater distance, even by just a few tenths of a millimetre, would bring about a drastic drop in performance characteristics, so that the need arises here for quite expensive and sophisticated control provisions of an electronic type to be implemented to this purpose.
Moreover, these linear compressors require that a return force proportionate to the displacement of the moving element be available in view of ensuring a correct displacement stroke of the piston, and the simplest method available to ensure this force lies in providing a mechanical spring duly connected to both the stationary base of the compressor and the moving element, such that the thereby obtained mechanical system is capable of working in the way of a forced resonant harmonic oscillator at line frequency; however, linear compressors equipped with a system of this kind generally experience a drastic deterioration in the overall performance characteristics thereof when departing from the resonance condition.
It is to be further noticed that the value of the refrigeration capacity, i.e. cooling effect that can be obtained with linear compressors of the currently known kind cannot be varied to a significant extent, i.e. by a factor of approx. 2, without this implying a considerable performance loss; such variation in the refrigeration capacity may for instance be required in view of particular application-related needs, or even in view of saving energy by adapting the power output of the compressor to the actual refrigeration capacity needed by the refrigerator in a given period of operation thereof. In fact, as this has already been noted above, a variation in the piston stroke entrains a performance loss owing to the constraints deriving from the position of the dead centre relative to the cylinder head and the position of the middle oscillation point of the piston, even the latter being practically unmodifiable without a reduction in motor efficiency and, hence, compressor performance resulting therefrom. Neither a variation in the oscillation frequency of the piston is a feasible option, since it cannot be practiced without incurring a loss in compressor performance, owing to the constraint requiring that the resonance condition of the system be kept unaltered.
It therefore is a main object of the present invention to do away with all of the aforementioned drawbacks of prior-art solutions by providing a linear compressor in which performance and efficiency characteristics are in no connection with, i.e. not tied to and anyway substantially unaffected by construction, design and operating requirements and variations thereof.
A major purpose of the present invention within the above-indicated object thereof is to provide a linear compressor in which the possibility is given to obtain wide variations in the refrigeration power output thereof, without these variations determining a deterioration in the overall performance characteristics of the same compressor.
Another major purpose of the present invention is to provide a linear compressor in which the positioning of the top dead centre of the piston relative to the cylinder head may allow for greater tolerances, without this impairing the general performance characteristics of the compressor, thereby allowing the systems used to control the position of the piston to be simplified and, hence, made far less expensive.
A further major purpose of the present invention is to ensure the possibility for the refrigeration power output of the compressor to be modulated by varying the stroke of the piston, or the oscillation amplitude thereof relative to the midpoint, and the clearance or dead volume, while maintaining energy efficiency values unvaried at a high level.
Another major purpose yet of the present invention is to provide a linear compressor that is significantly simplified in its construction, while at the same time ensuring an unvaried, or possibly even improved, efficiency, along with an improved flexibility in operation as compared to prior-art linear compressors.
A last, although not less important purpose of the present invention is to provide a linear compressor which is low and competitive in costs and capable of being manufactured with the use of readily available machinery and techniques.
According to the present invention, the above indicated aims and advantages, along with further ones that will become apparent from the description given below, are reached in a linear compressor incorporating the features and characteristics as recited in the appended claim 1.
Further features and advantages of the linear compressor according to the present invention may be more readily understood from the description that is given below of a particular, although not sole embodiment, which is illustrated by way of non-limiting example with reference to the accompanying drawings, in which:
FIG. 1 is a longitudinal sectional view of a linear compressor according to the present invention;
FIG. 2 is a detail view of the sectional view of FIG. 1;
FIG. 3 is a diagrammatical view of a schematical representation of the operating cycle of a linear compressor according to the present invention;
FIG. 4 is a view, along a section plane at 90° with respect to FIG. 1, of a linear compressor according to the present invention.
With reference to the above cited and listed Figures, the linear compressor, as generally indicated at 1, comprises a stator body 2 that is substantially constituted by an outer yoke 3, around which there is wound a coil (not shown), and an inner yoke 4 facing the outer yoke 3 and spaced from the latter so as to define an air gap 5.
The linear compressor further comprises a moving element 6 comprising a base plate 7, from which a pair of arms (not shown) extend in a manner known as such in the art, these arms being provided each with a magnet and being accommodated inside the air gap 5. The linear compressor also comprises a shaft 8, to the end portion of which there is firmly joined a piston 9, wherein the base plate 7 is firmly joined to a pan 10, which is in turn connected to a resonance spring 11.
As largely known in the art, energizing the compressor with an alternating current supply gives rise to the generation of a magnetic flux that causes the moving element 6 to perform a reciprocating translatory motion relative to the stator body 2; via the shaft 8, this motion is then transmitted to the piston 9.
The shaft 8 is accommodated slidably within a cylindrical guide body 12 that terminates with a flange 13, which is advantageously obtained integrally with said guide body 12, and against there abuts the liner 14 of a cylinder 15, inside which there is arranged slidably said piston 9. The flange 13 is provided with at least a suction port 16, controlled by suction valve means 17 such as a reed valve, in communication with the suction conduit comprising a conduit 18 connected to a reservoir 19 in which there is collected the gas being taken in, which then flows into the conduit 18 to eventually undergo compression. In an advantageous manner, the suction conduit is fitted between the flange 13 and the stator body 2, so that the inflow of the gas in the cylinder 15 takes place from the side at which the wall temperature is at its lowest. This contributes to reducing the extent to which the gas is heated up during the suction phase, thereby boosting compressor efficiency.
The cylinder 15 is closed by a head 20; the stroke of the piston 9 inside the cylinder 15 is therefore limited, on a side, by the flange 13 and, on the opposite side, by the head 20. The latter is provided with at least an exhaust or delivery port 25 provided with exhaust or delivery valve means 30, e.g. a reed valve, for controlling said delivery port 25.
Inside the cylinder 15, the piston 9 defines a first chamber 21, or a low-pressure chamber as this shall be better explained further on, and a second chamber 22, or a high-pressure chamber as this shall again be better explained further on, in which both of these chambers are variable-volume ones depending on the position of the piston 9 in the cylinder 15; the two chambers 21 and 22 are set in communication with each other via at least a through-aperture 23 provided in the piston 9 and controlled by communication valve means 24, such as a reed valve.
Thanks to the thus obtained configuration, the piston 9 is capable to alternately compress the gas in either direction inside the cylinder 15, thereby generating two compression stages, in which the exhaust or delivery phase in the first low-pressure chamber 21 occurs at the same time as the suction phase in the second high-pressure chamber 22, through the communication aperture 23 and the communication valve means 24 provided in the piston 9. In other words, the compression cycle is split into two stages in phase opposition with respect to each other with a phase shift of 180°.
The way in which the above-described linear compressor works is as follows, with reference in particular to FIG. 2 and the diagram of the compression cycle illustrated in FIG. 3, in which in correspondence to the axis X there is represented the position of the piston 9 in the cylinder 15 and in correspondence of the axis Y there is represented the value of the gas pressure: starting from the bottom dead centre X₁, in which the piston 9 is positioned adjacent to the flange 13, as the piston 9 displaces away from the flange 13 during its stroke, the gas contained in the clearance volume of the first chamber 21 (curve A-B of the diagram in FIG. 3) undergoes an expansion from the pressure P₁to the pressure P₂corresponding to the suction pressure (point B); this pressure P₂causes the suction valve means 17 to open and, as a result, the gas to be taken in from the reservoir 19 of the suction conduit into the first low-pressure or pre-compression chamber 21 through the conduit 18 and the suction port 16 (suction phase, line B-C); at the end of the suction phase, the piston 9 is at its top dead centre X₂, which is situated at a minimum distance from the head 20, thereby defining the maximum volume for the first low-pressure chamber 21.
Concurrently with this suction phase A-B-C in the first low-pressure chamber 21, and starting from the position of the piston 9 corresponding to the bottom dead centre X₁, the gas existing in the second high-pressure chamber 22 is first compressed from the pressure P₁up to the point at which the exhaust or delivery pressure P₃is reached (curve A-E), whose value is such as to cause the exhaust or delivery valve means 30 to open and, hence, the exhaust, i.e. delivery of the gas through the delivery port 25 in the head 20 (delivery phase, line E-F). At the end of this exhaust or delivery phase (point F), the piston 9 is in a position corresponding to its top dead centre X₂, at a minimum distance from the head 20 where the delivery port 25 is situated. From this position, the displacement motion of the piston 9 starts then to be reversed, thereby determining the expansion of the gas in the clearance volume in the second high-pressure chamber 22 (curve F-D) and, at the same time, the compression of the suction gas contained in the first low-pressure chamber 21 (curve C-D) up to the point at which the pressure P₁is reached establishing a balance between the first chamber 21 and the second chamber 22 (point D). This pressure value causes the communication valve means 24 provided on the piston 9 to open, thereby enabling the pre-compressed gas to flow from the first chamber 21 into the second chamber 22 via the through-aperture 23 in the piston 9 (line D-A), until the latter eventually reaches the bottom dead centre X₁in its displacement stroke. In this manner, the linear compressor according to the present invention performs a cycle corresponding to two compression stages in phase opposition with respect to each other, as indicated in FIG. 3 by the area comprised between the points A, B, C, D as far as the low-pressure stage taking place in the first chamber 21 is concerned, and the area comprised between the points A, E, F, D as far as the high-pressure stage taking place in the second chamber 22 is concerned.
In FIGS. 1 and 4 there is furthermore illustrated a lubrication and cooling system that may advantageously be used in a linear compressor according to the present invention: with the help of the pump 31, the lubricant is collected into the reservoir 32 and caused to flow through a first channel 33 and one or more first ports 34 provided circumferentially in the cylindrical guide body 12 so as to lubricate the shaft 8. From here, the lubricant then flows, through a second channel 35, a second port 36 provided in the flange 13, and a third channel 37 in the liner 14 of the cylinder 15, into a hollow space, i.e. jacket 38 provided in the liner 14 to cool down the wall of the cylinder 15, in view of keeping it at a lower temperature to the purpose of both favouring the thermal efficiency and boosting the overall energy efficiency of the compressor.
In order to lubricate the piston 9 and reduce frictions during the displacement thereof, a certain amount of lubricant is dragged by the outer surface of the portion of the shaft 8 entering the first low-pressure chamber 21.
From the description given above it can therefore be readily appreciated that the linear compressor according to the present invention is actually capable of reaching all of the afore indicated aims and advantages: in fact, with the linear compressor according to the present invention the possibility is given for high refrigeration outputs to be obtained even in the presence of limits of oscillation of the piston 9—corresponding to the top and bottom dead centres—that may be situated at a considerable distance from the respective cylinder head portions, i.e. the head 20 and the flange 13. No need therefore exists for minimum clearances, in the order of magnitude of 0.1 mm, to be in any case ensured, as this is on the contrary required in prior-art single-stage compressors, thereby allowing for the use of less sophisticated and accurate and, therefore, less expensive systems for controlling the position of the piston 9.
Moreover, fractioning in the above-described way the compression phase into two stages enables far higher volumetric efficiencies to be obtained as compared to single-stage compressors for a same amount of clearance volume. In addition, the refrigeration capacity output of the compressor can be modulated up to a factor of 2 by either varying the displacement stroke of the piston 9 relative to the midpoint between the top dead centre and the bottom dead centre or varying the clearance volume, while maintaining energy efficiency at an unaltered high level and excluding any appreciable deterioration in the overall performance of the compressor.
In this way, the performance and efficiency characteristics of the compressor according to the present invention are in no connection with, i.e. not tied to and anyway substantially unaffected by construction, design and operating requirements and variations thereof.
A further advantage of the linear compressor according to the present invention derives from the fact that it is embodied in such manner as to ensure a maximum extent of simplicity in its construction and, at the same time, high performance capabilities without any need for any component part of the compressor, such as the piston, the cylinder head or the valves, to be duplicated.
It will of course be appreciated that the present invention may be subject to a number of modifications and variants, and may be used in conjunction with a number of different applications, without departing from the scope of the present invention.
It should further be noticed that the materials used to implement the present invention, as well as the shapes and the size of the individual component parts, may each time be selected so as to most appropriately fit any particular need or comply with any application-related requirement, without this implying any departure from the scope of the present invention.

Claims

1-10. (canceled)

11. Linear compressor comprising a piston (9) arranged slidably inside a cylinder (15), said piston (9) defining in said cylinder (15) a first low-pressure chamber (21) and a second high-pressure chamber (22), said piston (9) being further provided with communication means (23) between said first and said second chamber (21, 22), which are controlled by valve means (24) to establish such communication therebetween, characterized in that said first and second chambers (21, 22) have substantially the same diameter, said piston (9) being part of a moving element (6) comprising a base plate (7) from which there extends a shaft (8) supporting said piston (9) at an end portion thereof.

12. Linear compressor according to claim 11, wherein said first and said second chamber (21, 22) have a volume that is variable depending on the position of said piston (9) in said cylinder (15).

13. Linear compressor according to claim 11, wherein said communication means comprise at least a communication aperture (23) provided in said piston (9).

14. Linear compressor according to claim 13, wherein said shaft (8) is accommodated slidably inside a guide body (12) terminated by a flange (13) that is provided with at least a suction port (16) controlled by suction valve means (17).

15. Linear compressor according to claim 14, wherein said flange (13) is provided integral with said guide body (12).

16. Linear compressor according to claim 13, wherein said piston (9) compresses a fluid in either direction inside said cylinder (15), thereby generating two compression stages, in which the exhaust or delivery phase in said first low-pressure chamber (21) occurs concurrently to the suction phase in said second high-pressure chamber (22) through said communication aperture (23) and said communication valve means (24).

17. Linear compressor according to claim 13, wherein the displacement stroke of said piston (9) moving away from said flange (13) generates, in the first low-pressure chamber (21), the expansion phase (AB) and suction phase (BC) as controlled by said suction valve means (17) and, in the second high-pressure chamber (22), the compression phase (AE) and exhaust phase (EF) as controlled by said exhaust valve means (30) said communication aperture (23) being closed by said communication valve means (24) during said suction phase in the first low-pressure chamber (21) and said compression phase in the second high-pressure chamber (22).

18. Linear compressor according to claim 11, characterized in that it further comprises a lubrication and cooling system comprising a first channel (33) establishing a communication between a reservoir (32) and one or more first ports (34) provided circumferentially in the cylindrical guide body (12), and a circulating pump (31) adapted to deliver a lubricant between said shaft (8) and said guide body (12) from said reservoir (32) through said first channel (33) and said first ports (34).

19. Linear compressor according to claim 18, wherein said lubricant further flows from said guide body (12), through a second channel (35), a second port (36) provided in said flange (13) and a third channel (37) provided in said cylinder (15), into a hollow space or jacket (38) obtained in a wall of said cylinder (15).