KR20060003912A - Linear compressor - Google Patents

Abstract

A linear compressor comprising a cylinder part (10) including a cylinder bore (71) and a piston (4) disposed in, and slidable within the bore (71). A main spring (15) connects the cylinder part (10) directly of indirectly with the piston (4). A connecting rod (47) connects between the main spring (15) and the piston (4). A stator (5) has an air gap through which the connecting rod (47) passes. At least one armature pole (2) is located along the connecting rod (47).

Description

LINEAR COMPRESSORS {LINEAR COMPRESSOR}

1 is a partial exploded view from above of a linear compressor according to a preferred embodiment of the present invention;

2 is an exploded view of the motor end of the compressor assembly viewed from another direction;

3 is an exploded view of the head end of the compressor assembly seen from one direction;

4 is an exploded view of the head end of the compressor assembly seen from another direction;

The present invention relates to a method for manufacturing a linear compressor, and more particularly to a method for manufacturing a linear compressor that can also be used in refrigerators and other applications.

Compressors, particularly refrigerator compressors, are usually driven by rotary electric motors. However, even in the most effective form of these compressors, the losses associated with crank systems that convert rotational motion into linear reciprocation are significant. Alternatively, a rotary compressor that does not require a crank can be used, but a high centripetal load is generated that causes significant frictional losses. Linear compressors driven by linear motors do not have these losses, and the laterally flexible connecting rods allow low bearing loads to allow the use of aerodynamic gas bearings, as disclosed in US Pat. No. 5,525,845. Can be designed to have a sufficiently low bearing load.

A paper on hydrostatic gas bearings was published in 1970 by The Machinery Publishing Company Limited, London, and the author is J W Powell, entitled "Design of Static Bearings." However, the production of effective gas bearings is difficult with conventional manufacturing tolerances and devices.

Conventional compressors are mounted in sealed housings that, in use, act as a reservoir of refrigerant gas. Refrigerant gas is introduced from this reservoir into the compressor and discharged through an exhaust conduit leading from the compressor through the housing.

The operation of the compressor involves the reciprocating operation of the moving part which causes vibration of the compressor unit in all three axial directions. In order to reduce the external noise effect of this vibration, the compressor is mounted to an isolation spring in a sealed housing.

In a linear compressor, the piston vibrates with respect to the cylinder in only one axial direction, and if there is a fixed part, the resulting reaction force is applied to that part. One proposed solution to this problem is to simultaneously operate a pair of compressors in a balanced and opposing configuration. However, such devices are too complicated and expensive for use in products such as household refrigerators. Another proposed solution is to add a resonance counterweight to reduce vibration. However, this approach limits the operation of the compressor, since the counterweight is a negative feedback device and limited to the fundamental imbalance. Another solutions is the minutes of Montpellier in mice four issued in 1990 the International Christchurch Cooler conference in Massachusetts ^{(Proceedings of the 6 th International Cryocooler} Conference) from a presenter Gully and Hanes titled "Small rotary and linear for IR system Vibration characteristics of the cryogenic cooler. This solution involves independently supporting the piston and cylinder portions of the compressor in the housing such that the stator acts as a counterweight. However, problems have arisen when implementing such a design in a domestic refrigerator when the piston mass is low. In such a compressor, as the discharge pressure increases, the force of the compressed gas acts as a spring (gas spring), which increases the operating speed as the discharge pressure increases. This is the "second" mode where only the "third" vibration mode (the piston and cylinder vibrate in phase with each other but in a different phase than the compressor shell) is preferred (the shell does not vibrate and the piston and cylinder are in a different phase). The problem is because it's a bit above. Thus, the shell begins to vibrate excessively as the "gas spring" begins to operate and consequently raises the "second mode" frequency to the "third mode" frequency.

It is desirable for the compressor to be compact for many applications. This reduces the size of all components including the spring and its resonant system. To reduce the size of the compressor, the compressor must be operated at higher frequencies. The reduced magnitude at higher frequencies is associated with increasing the stress of the spring component. In some linear compressors the main spring is made of pressed steel plate. Edges cut in press operations require careful polishing to restore the original strength of the spring steel plate and are often broken by accidental stress concentrations.

Another problem with compressors is usually heat concentration in the vicinity of the cylinder and cylinder head of the compressor. This heat concentration is generated by the friction between the movable components and by the heat transferred from the compressed refrigerant. Heat concentration causes significant problems such as increased wear and tolerances and operating conditions between components that change with the cycle at which the compressor is operated. This effect is particularly noticeable in linear compressors, which operate for long periods of time when using hydrostatic gas bearing systems, and where tight gaps are important.

Another exothermic effect arises from irreversible heat loss before re-expansion from the refrigerant remaining in the compression space after compression. Up to 15% of the compressed refrigerant is not discharged in the linear compressor and up to 5% of the compressed refrigerant in the crank driven compressor. This heat source, negligible in conventional compressors, is an important source in linear compressors.

One approach to cooling the cylinders and cylinder heads of a compressor involves the use of liquid refrigerant from a condenser in subsequent refrigeration systems. For example, in US Pat. No. 2510887, the liquid refrigerant from the condenser is supplied to the first cooling jacket surrounding the cylinder and from there to the second cooling jacket surrounding the compressor head and then from the venturi device between the cylinder head and the condenser. It is discharged to the connecting discharge line. This is the content of a standard crank driven compressor. In the context of a standard crank driven compressor, US Pat. No. 5694780 discloses a circuit in which liquid refrigerant from a condenser is raised to higher pressure by a pump. This liquid refrigerant is pumped into the cooling jacket surrounding the compressor cylinder. This liquid refrigerant is forced to the discharge manifold from the cooling jacket into the cylinder head of the compressor, where the liquid refrigerant mixes with the compressed refrigerant as the compressed refrigerant exits the compressor. Such a device has the disadvantage of requiring an additional pump to force the liquid refrigerant through the cooling jacket surrounding the cylinder and then against the pressure of the compressed gas into the discharge manifold.

Many linear compressors are manufactured inside conventional compressor seats as intended as a replacement for existing rotary reciprocating compressors that are commercially available. In order to achieve this compact size, a compressor has been manufactured in which the stator, armature, cylinder and piston are all located concentrically. However, conventional compressor dimensions impose constraints on the size of the mechanical compartment of the refrigerator, resulting in wasted space in the compartment surrounding the compressor.

It is an object of the present invention to provide a method of manufacturing a linear compressor that takes several ways to overcome the above disadvantages.

The present invention is a method of manufacturing a linear compressor, in order to connect the piston of the linear compressor to the spring of the linear compressor,

(a) positioning the piston in a predetermined axial position in the bore of the cylinder of the compressor;

(b) positioning a piston connection point of the spring already connected with the cylinder portion at a position based on a predetermined displacement of the spring or at a position at which the spring exerts a predetermined reaction force; And

(c) connecting the piston and the piston connection point of the spring to a rigid axial spacing fixed according to the spacing obtained through steps (a) and (b),

Each of the above steps is a method of manufacturing a linear compressor that can be performed in any order.

Various changes in construction of the invention and a wide variety of embodiments and applications can be proposed by those skilled in the art to which the invention pertains without departing from the scope of the invention as defined in the appended claims. The present disclosure is illustrative only and is not meant to limit the technical spirit of the present invention.

[Overall Configuration]

The practical embodiment of the invention shown in the figures includes a permanent magnet linear motor that drives a resonant reciprocating compressor and operates together in a hermetic casing. The compressor includes pistons 3 and 4 which reciprocate in the cylinder bore 71 and operate on a working fluid which is alternately introduced into and discharged from the compression space at the head end of the cylinder. The cylinder head 27 connected to the cylinder surrounds the open end of the cylinder bore 71 to form a compression space, and includes inlet and outlet valves 118 and 119 and a manifold connected thereto. The compressed working gas exits the compression space through the outlet valve 119 and flows into the discharge manifold. The discharge manifold directs the compressed working fluid into the cooling jacket 29 surrounding the cylinder 71. The discharge tube 18 passes through the hermetic casing starting from the cooling jacket 29.

The cylinder 71 and the jacket 29 are integrally formed as a single body 33 (for example, casting). The jacket 29 includes one or more end opening chambers 32 which are substantially aligned with the reciprocating axis of the cylinder 71 and which surround the cylinder 71. The end opening chamber 32 is substantially enclosed to form a jacket space (by the cylinder head assembly 27).

The linear motor comprises a pair of opposing stator parts 5, 6 that are rigidly connected to the cylinder casting 33.

The pistons 3, 4 reciprocating in the cylinder 71 are connected to the cylinder assembly 27 via a spring system. It operates at or near its natural resonant frequency. The primary spring element of the spring system is the main spring 15. The pistons 3, 4 are connected to the main spring 15 via a piston rod 47. The main spring 15 is connected to a pair of legs 51 extending from the cylinder casting 33. The pair of legs 41, stator parts 5, 6, cylinder molding 33 and cylinder head assembly 27 together constitute what will be referred to as cylinder part 1 in the description of the spring.

The piston rod 47 connects the pistons 3 and 4 to the main spring 15. The piston rod 47 is preferably a rigid piston rod. The piston rod has a plurality of permanent magnets 2 spaced along it, forming the armature of the linear motor.

For low frictional loads between the pistons 3, 4 and the cylinder 71, in particular to reduce any lateral loads, the piston rod 47 is elastic with both the main spring 15 and the pistons 3, 4. It is connected flexibly and flexibly. In particular, an elastic connection is provided between the main spring end 48 of the piston rod 47 in the form of a molten plastic connection between the overmolded button 49 on the main spring and the piston rod 47. The piston rod 47 comprises a pair of spaced apart circular flanges 3, 36 which fit in the piston sleeve 4 at its other end to form a piston. These flanges 3, 36 are continuous with, but alternately located with, the pair of hinge regions 35, 37 of the piston rod 47. The pair of hinge regions 35 and 37 are formed to have a major bending axis at right angles to each other.

At the main spring end 48 the piston rod 47 is effectively radially supported relative to the main spring 15 by its connection. The main spring 15 is provided to provide reciprocation but substantially prevent any lateral movement or movement across the reciprocating direction of the piston in the cylinder. In a preferred embodiment of the present invention, the lateral stiffness is approximately three times the axial stiffness.

The assembly comprising the cylinder portion is not fixedly mounted in the hermetic casing. It is free to move in the reciprocating direction of the piston away from the support connection to the casing: the discharge tube 18, the liquid refrigerant injection line 34 and the rear support spring 39. Each of the discharge tube 18, the liquid refrigerant injection line 34 and the rear support spring 39 is formed to be a spring of known characteristics in the reciprocating direction of the piston in the cylinder. For example, the tubes 18, 34 may be formed of spiral or helical springs adjacent the ends passing through the hermetic casing 30.

The total reciprocation is the sum of the movements of the pistons 3 and 4 and the cylinder part.

[Gas bearing]

The pistons 3, 4 are supported radially in the cylinder by aerodynamic gas bearings. The cylinder portion of the compressor includes a cylinder casting 33 having a bore 71 therethrough and a cylinder liner 10 in the bore 71. The cylinder liner 10 may be made from a suitable material to reduce piston wear. For example it may be made from fiber reinforced plastic composites such as carbon fiber reinforced nylon with 15% PTFE (which is also preferred for piston rods and sleeves) or may be cast iron with self-lubricating graphite of its flaky graphite. The cylinder liner 10 has an opening 31 extending therethrough from its outer cylindrical surface 70 to its inner bore 71. The pistons 3, 4 move in the inner bore 71, and these openings 31 form a gas bearing. The supply of compressed gas is supplied to the opening 31 by a series of gas bearing passages. The gas bearing passageway opens at its other end to the gas bearing supply manifold, and the gas bearing supply manifold is annular chamber around the cylinder liner 10 at its head end between the liner 10 and the cylinder bore 71. It is formed as. The gas bearing supply manifold is then supplied by the compressed gas manifold of the compressor head by the small supply passage 73. The small dimensions of the feed passage 73 control the pressure in the bearing supply manifold to limit the gas consumption of the gas bearing.

The gas bearing passage is formed as a groove 80 in the outer wall 70 of the cylinder liner 10. These grooves 80 combine with the walls of the other cylinder bores 71 to form enclosed passages leading to the opening 31. Although grooves may optionally be provided in the inner wall of the cylinder bore 71, it will be appreciated that the grooves are easier to form in the liner 10 than in the cylinder casting 33 and on the outer surface rather than the inner surface. The ability to machine grooves into the surface of one or the other part is an important fabrication improvement compared to having to drill or boring the aisle.

It has been found that the pressure drop in the gas bearing passage is required to be similar to the pressure drop in the discharge flow between the pistons 3 and 4 and the bore 71 of the cylinder liner 10. Since the gap between the pistons 3 and 4 and the cylinder liner bore 71 is only 10 to 15 microns (for an effective compact compressor), the cross-sectional dimension of the passage is also very small (a depth of about 40 microns x 150 microns wide). Is required. These small dimensions make production difficult.

However, in a preferred embodiment of the present invention, such a configuration is easily achieved by increasing the length of the passageway so that the cross-sectional area can also be increased to, for example, 70 microns x 200 microns. This has the advantage that the groove 80 of any suitable shape can be formed in the surface of the liner portion 10. The groove 80 can be formed with any path, and if the curved path is selected, the length of the groove 80 is greater than the straight path from the gas bearing supply manifold and each gas bearing forming opening 31. Can be significantly longer. This preferred embodiment has a gas bearing groove 80 that follows the helical path. The length of each path is chosen to suit the desired cross-sectional area of the path that can be selected for easy fabrication (by other forming methods such as machining or precision casting).

[Cylinder part]

Each part 5, 6 of the stator has a winding. Each part 5, 6 of the stator is formed with an "E" shaped stack, with the winding supported around the central pole. The windings are insulated from the lamination stack by plastic bobbins. This particular form of each stator portion does not limit the invention, and many possible structures will be apparent to those of ordinary skill in the art.

As already mentioned, the cylinder part 1 comprises a cylinder 71 and a cooling jacket 29, a cylinder head 27 and a linear motor stator part 5, 6 coupled thereto, all of which are rigid with each other. It is connected. The cylinder part 1 also includes mounting points for the main spring 15, the discharge tube 18 and the liquid injection tube 34. It also has a mount for connecting the cylinder part to the main spring 15.

The cylinder and jacket casting 33 has upper and lower mounting legs 41 extending from its ends, away from the cylinder head. The spring 15, the preferred form of which will be described later, comprises a rigid mounting bar 43 at one end for connection with the cylinder casting 33. A pair of laterally extending lugs 42 extend from the mounting bar 43 in the direction towards the cylinder casting 33 and are spaced apart from the mounting bar 43. The lug 42 is arranged to axially coincide with the portions 67, 68 of the spring immediately adjacent to the entry into the mounting bar 43 of the spring end. Each of the upper and lower mounting legs 41 of the cylinder casting 33 includes mounting slots for either of the lugs 42. The mounting slot has the form of a rebate 75 on the inner surface 76 of the mounting leg 41, which extends from the free end 77 of the leg 41 toward the cylinder jacket 29. At least one tapered protrusion 78 is formed on the inward surface 82 of each rebate 75. Each such protrusion 78 has a vertical surface 79 that faces the cylinder casting 33 so that the protrusion 78 forms a barb for snap fit connection during assembly. In particular, the lateral spacing of the lugs 42, which is approximately fitted to the space between the opposing surfaces 82 of the rebate, allows the lugs 42 to be barb 78 only through deformation of the lugs 42, the mounting legs 41 or both. ) To pass through. Once passed through the protrusion or barb 78, lug 42 is caught between vertical plane 79 of barb 78 and vertical plane 83 forming the end face of rebate 75. Additional rebates or depressions 84 are formed on the outer surface 85 of each of the mounting legs 41. This rebate 84 extends axially along the outer surface 85 such that each of the rebates 84 is at least fitted with each rebate 75 on the inner surface 76 of its mounting leg 41. It is spaced apart from the free end 77 of 41 by a fixed distance. The rebates 75 and 84 are of sufficient depth to provide a place therebetween where they fit or overlap at least with the axial opening 86.

Preferably the clamping spring 87 formed from the stamped and folded non-magnetic sheet metal has a central opening 88 through it, which can be fitted over a pair of mounting legs 41. The clamping spring 87 has a rear extending leg 89 which engages with each mounting leg 41. The free end 90 of these legs 89 can slide within the outer rebate 84 of the mounting leg 41 and pass through the axial opening 86 between the outer and inner rebates 84, 75. Small enough to be. By the lugs 2 of the main spring bar 43 located in the inner rebate 75 of the mounting leg 41, these free ends 90 press the lugs 42 and force them to their respective barbs 78. With respect to the vertical plane (79). Retention of the clamping spring 87 in the pressurized state provides a predetermined preload on the lug 42.

It is preferable that the clamping spring carries out the work of mounting the stator parts 5, 6. The central hole 88 penetrating the clamping spring 87 is made close to the mounting leg 41 in the lateral direction at least relative to the mounting leg 41. The clamping spring 87 includes a stator partial clamping surface 91 in each side region 92 spanning between the mounting legs 41 of the cylinder casting 33.

The cylinder casting 33 includes a pair of protruding stator support blocks 55 extending from the spring side surface 58 of the jacket 29 at positions between the positions of the mounting legs 41.

Each " E " shaped stack of stator portions 5, 6 has a vertical step 69 directed vertically on its outer face 56. As shown in FIG. In each case this is an outward step in a direction away from the motor air gap. The portion of each face 56 close to the air gap is suitably supported with respect to the support block 55 of the cylinder casting 33 or the stator engaging surface 91 of the clamping spring 87. When in the proper position the attraction between the parts of the motor pulls the stator parts 5, 6 towards each other. The width of the air gap is maintained by the position of the vertical step 57 with respect to the outer edges 40, 72 of the mounting block 55 and the clamping spring 87, respectively. In order to further place the stator portions 5, 6 in a vertical direction, the notch 57 is fitted with the dimensions of the "E" shaped stack stack in a direction perpendicular to its outer edge. It includes.

This part of the motor is assembled in a series of operations. The piston assembly is introduced into the cylinder casting 33. First, a piston assembly comprising a piston rod 47 with integral flanges 3 and 36, a piston sleeve 4 and an armature magnet 2 is assembled. The piston formed by the piston sleeve 4 and the leading flange 3 forming the piston face is pushed into the bore 71 of the cylinder and the cylinder opening 7 lying between the legs 41 of the cylinder casting 33. Is introduced through. The piston connecting rod 47 therefore lies between the legs 41. The inward face 76 of the leg 41 includes an axial slot 28 extending from the rebate 75 along the centerline. Outwardly extending lugs 130 on the piston connecting rod 47 reciprocate in these slots 28 during operation. The accuracy of the assembly is generally such that the lug 130 does not contact the surface of the slot (in the absence of knocking or external movement). However, at its end adjacent to the cylinder casting 33, the slot 28 comprises a thin and narrow portion 131 which provides a fit with the lug 130 of the connecting rod 47. The piston assembly has a cylinder bore until the lug 130 is engaged in the narrow portion 131 and the face of the piston is in position relative to the machined cylinder head receiving face 133 of the cylinder casting 33. It is pushed into 71.

The clamping spring 87 fits over the mounting leg 41 and the main spring 15 is engaged by engagement of the lug 42 into the inward rebate 75 and through the barb or protrusion 78. 41). The clamping spring 87 is pushed away from the cylinder casting 33 and compressed against the mounting lug 42 so that the stator portion engaging surface 91 of the clamping spring 87 and the mounting block of the cylinder casting 33 ( Sufficient space for the introduction of the stator part between 55 is possible. The stator parts 5, 6 are then introduced into place and the clamping springs 87 are released. The width of the stator parts 5, 6 maintains a predetermined compression in the clamping spring 87.

Thereafter a connection is made between the main spring 15 and the piston rod 47. The piston is at a predetermined position facing the cylinder head more than when the compressor is operating. The connection is made by melting the plastic of the molded button 25 on the main spring 15 and the plastic of the rear end 48 of the piston connecting rod 47. Melting is performed by hot plate welding. In hot plate welding the spring 15 preferably extends to a predetermined position or until the spring 15 exerts a predetermined force. Hot plate welding and melting of the two plastic components with the piston face in a predetermined position and the spring 15 in a predetermined displacement is performed by the piston against the cylinder casting 33 when the spring 15 is released to its neutral position. Provide accurate placement. This is independent of accumulated inaccuracies due to any eccentricity in the form of springs 15 or tolerances of the components of the assembly chain.

[Cylinder head]

The open end of the cylinder casting 33 is surrounded by the compressor head 27. The compressor head thereby surrounds the cylinder 71 and the open end of the cooling jacket chamber 32 surrounding the cylinder 71. In its overall form the cylinder head 27 comprises a stack of four plates 100-103 with an intake muffler / intake manifold 104.

The open end of the cylinder casting 33 includes a head fixing flange 135. The head retaining flange 135 has several threaded holes 136 spaced along its periphery, in which fastening bolts are tightened to pull the stack of plates 100-103 and face the cylinder molding 33. Secure in.

An annular rebate 133 is provided on the face of the flange 135. On the opposite side of the flange 135, the rebate 133 includes outwardly extending lobes 137 and 138 that act as ports for the evacuation tube 18 and return tube 34, respectively.

An opening is provided between the three chambers of the cylinder casting 33.

The first head plate 100 is fitted over the open end of the cylinder molding 33 in the annular rebate 133. This is relatively flexible and acts like a gasket. It surrounds the cylinder jacket opening but has a central opening and does not cover the open end of the cylinder 71. The compressed gas return port 110 extends through the plate 100 adjacent to the lobe 138 associated with the liquid refrigerant return pipe 34. The edge of the opening 110 proximate the outer wall of the cooling jacket 29 is slightly spaced from the wall at least near the lobe 138. The effect of this is that as the gas moves through the opening 110 into the cooling jacket chamber, a small zone is created that is decompressed immediately after the plate 100 and adjacent to the lobe 138 of the rebate 135.

Another hole 115 through the plate 100 is provided at a position closer to the discharge pipe 18.

The second head plate 101 is fitted on the first plate 100. The second plate 101 has a larger diameter than the plate 100 and is rigid. It is made of steel, cast iron, or sintered steel. The plate 101 is wider than the rebate on which the plate 100 is placed. The plate 101 is placed against the face of the flange and compresses the first plate 100 against the rebate. The plate 101 has an opening 139 that is spaced along its periphery and freely penetrates the threaded portion of the bolt.

The second head plate 101 includes a compressed gas discharge opening 111 that is fitted with the opening 110. It also includes another opening 117 that fits with the opening 115 of the first plate 100.

The portion of the plate 101 surrounds the cylinder opening 116 of the plate 100. Intake port 113 and discharge port 114 pass through this portion of plate 101. The spring steel inlet valve 118 is fixed to the side facing the cylinder of the plate 101 with its head covering the intake port 113. The base of the inlet valve 118 is clamped between the plate 100 and the plate 101 and its position is fixed by the dowel 140. The spring steel discharge valve 119 is attached to the face of the plate 101 away from the cylinder. The head part covers the discharge opening 114. The base of the valve 119 is clamped between the second plate 101 and the third plate 102 and is positioned by the dowel 141. Discharge valve 119 fits within and operates within discharge manifold opening 112 of third plate 102 and discharge manifold 142 formed in fourth plate 103. Inlet valve 118 is placed in and actuated within the cylinder compression space (spaced from the base).

The spring steel inlet and outlet valves 118 and 119 operate under the influence of the overall effective pressure. When the piston retracts from the cylinder 71, there is a lower pressure on the cylinder side of the inlet valve 118 than on the manifold side. Thus the inlet valve 118 is opened to allow refrigerant to enter the compression chamber. When the piston is advanced in the cylinder 71, there is a higher pressure in the compression space than the inlet manifold, and this pressure difference keeps the inlet valve 118 in the closed position. The inlet valve 118 is pressed towards this closed position by its restoring force.

The discharge valve is likewise pressurized normally to the closed position, which is maintained by the effective pressure at the retraction of the piston in the cylinder 71. This is pushed to the open position by higher pressure in the compression space than the discharge manifold upon advancing the piston in the cylinder 71.

The third head plate 102 fits into the circular rebate 143 of the face 144 facing the cylinder of the fourth plate 103. The plate 102 is relatively flexible and acts like a gasket and is compressed between the fourth plate 103 and the second plate 101. The third plate 102 includes a large opening 112 that fits with a wide rebate 142 at the face 144 of the third plate 103. Opening 112 and rebate 142 form an outlet manifold through which refrigerant compressed from outlet port 114 flows.

The other opening 121 through the third plate 102 is fitted with the rebate 145 at the face 144 of the fourth plate 103. The opening 121 is also fitted with the openings 117, 115 of the second plate 101 and the first plate 100. The gas filter 120 receives the compressed refrigerant from the rebate 145 and guides it to the gas delivery supply passage 73 through the through holes 146 and 147 of the first and second plates.

The intake opening 95 through the third plate 102 is fitted with the intake port 113 of the second plate 101. The opening 95 is also fitted with the intake port 96 passing through the fourth plate 103. At the face 98 of the fourth plate 103, a truncated cone or tapered inlet 97 is connected to the inlet port 96. The intake port 96 is surrounded by the intake muffler 104. The intake muffler 104 comprises, for example, one piece molding having a space substantially open to the side facing the cylinder casting. When connected, this space is surrounded by the fourth plate 103. The suction muffler 104 is plated by some possible means, for example by fitting the peripheral lip of the muffler in the channel at the face of the fourth plate 104 or bolting the head plate to the cylinder casting through a flange on the muffler. And a muffler to the fourth plate 103 which is part of the stack. The suction muffler 104 includes a passage 93 extending from the enclosed suction manifold space to open in a direction away from the cylinder molding 33. This passage is a refrigerant intake passage. The inner projection 109 of the intake tube 12 extending through the sealed housing with the compressor placed in the sealed housing extends into the intake passage 93 with a margin of tolerance.

[Cooling jacket]

As previously described, an interconnected cooling jacket chamber 32 is provided around the compression cylinder 71. In refrigerant gas compression, the gas reaches a state having a temperature much higher than the temperature at which the gas enters the compression space. This heats the cylinder head 27 and cylinder wall liner 10.

In the present invention a supply of liquid refrigerant is used to cool the cylinder wall / liner, cylinder head and compressed refrigerant.

Liquid refrigerant is supplied from the condenser outlet in the refrigeration system. It is supplied directly to the cooling jacket chamber 32 around the cylinder. The freshly compressed refrigerant released passes through the chamber before leaving the compressor via the discharge tube 18. In the chamber 32 the liquid refrigerant absorbs a large amount of heat from the compressed gas, the peripheral wall of the cylinder casting 33, and the cylinder head 27 and evaporates.

The liquid refrigerant leaving the condenser is at a slightly lower pressure than the refrigerant immediately after compression, which interferes with the liquid flow to the discharge side of the pump. However, passive devices are preferably used to transfer the liquid refrigerant to the cooling jacket. In the present invention this is achieved by several structural aspects of the compressor. Firstly a low pressure subzone is provided immediately adjacent the outlet from the liquid return line 34 to the jacket space. The cause of this low pressure zone has already been described. This is shown through the flow of compressed gas into the jacket through the compressed gas opening 110 in the head plate 100. The second aspect is that a slight inertial pumping effect is produced by the reciprocating motion of the liquid refrigerant return pipe 34 in the longitudinal direction. This reciprocation is the result of the operation of the compressor, and is the result of the associated reciprocation of the entire cylinder portion and the vibrations of the cylinder create a variable acceleration of the axially aligned length 148 of the return line 34 and thus the inlet ( Creating a varying pressure at the end of the return line 34 of the liquid phase that meets 33). Since the liquid immediately evaporates as it enters the cooling jacket during the high pressure fluctuation, any backflow during the low pressure fluctuation is a gas flow. Since the gas is of low density compared to the liquid, a very small amount of mass flow returns to the cooling jacket through a small notch that meets the liquid return line 34. In conclusion there is a net flow of refrigerant into the cooling jacket.

A third aspect is that the compressor can be located below the condenser in the domestic refrigerator or refrigerator / freezer as intended. It can be seen that when the liquid refrigerant in the condenser does not enter the cooling jacket at a sufficient rate, the head of the liquid refrigerant is formed in the line extending to the compressor, leading to the pressure of the liquid refrigerant to the pressure of the compressed refrigerant. It can be seen that this effect sometimes provides a measure of automatic correction when the relative pressure in the system does not otherwise indicate to the liquid refrigerant being moved to the cooling jacket space.

[Refrigerant]

Another way to reduce the heat generated by the compressor relates to the refrigerant. Using isobutane (R600a) as the refrigerant in the linear compressor exhibits a synergistic effect than the use of other refrigerants.

One basic difference between a linear compressor and a conventional compressor is the volume on the piston at " top dead center. &Quot; In conventional compressors, this volume is fixed at a small rate of displacement by the shape of the compressor. In a free-piston linear compressor, this "dead" volume varies with input power and discharge pressure. At moderate power, the dead volume can be a significant proportion of operating displacement of up to 25%.

The dead volume acts as a gas spring that absorbs and releases energy upon compression and re-expansion. However, this is an unstable spring due to irreversible heat exchange with the cylinder wall. This heat exchange is approximately proportional to the temperature rise of the refrigerant upon compression (or to the temperature drop upon expansion). This temperature rise depends on the specific heat ratio of the gas. For a commonly used refrigerant R134a, this ratio is 1.127. This is similar to other refrigerants commonly used, for example CFC 12. However, for isobutane (R600a), the specific heat ratio is 1.108. The table below shows the adiabatic temperature of the gas being compressed to a saturation pressure corresponding to the condensation temperature of the refrigerator (0 ° C. evaporator temperature) and the freezer or refrigerator / freezer (evaporation temperature −18 ° C.). Both have a temperature of 32 ° C. superheated gas entering the compressor and a temperature of 42 ° C. condenser (these are typical values).

 Refrigerant Evaporator temperature -18 0 Isobutane 87 ℃ 69 ℃ HFC134a 99 ℃ 77 ℃ CFC12 104 ℃ 79 ℃

Therefore, it can be seen that the properties of isobutane and other refrigerants having a low specific heat ratio correspond specifically to the physical properties of the linear compressor. Experiments have demonstrated this improvement.

A series of tests were conducted in a 450 liter refrigerator. The mass flow meter was placed in the suction line immediately before the compressor. This refrigerator is tested in an environmental chamber in accordance with standard AS / NZS 4474.1-1997. This test was first done using refrigerant R134a. This test was then made in two cases using refrigerant R600a (isobutane). The test results are as follows:

R134a R600a (Test 1) R600a (Test 2) T _condenser 34.3 33.3 33.5 T _evaporator -3.5 -6.1 -4.5 Input power 22.3 18.1 17.8 Freezing force 75.25 75.01 75.07 Displacement 7.96 8.7 8.48 Coefficient of performance 3.38 4.14 4.23

The test results show that the coefficient of performance of linear compressors using isobutane is significantly improved compared with the use of R134a.

[Spring System]

The cylinder 9 is supported by the liquid refrigerant injection tube 34 and the discharge tube 18 having a combined stiffness k _cylinder in the axial direction. The support spring 39 has considerable low rigidity in the axial direction and has a negligible effect. This piston sleeve 4 is radially supported by a gas bearing formed by a cavity and a passage at the boundary of the cylinder bore and the cylinder liner 10. Since the main spring has a rigidity (k _main ) in order to obtain the resonance vibration of the piston and the cylinder, the resonance frequency f _n may be calculated as follows.

Where m _piston , m _cylinder are the elastic masses of the piston and the cylinder spring, and the natural frequency f _n is a frequency of 10 to 20 Hz, which is typically less than the desired operating frequency, thus increasing the frequency with increasing stiffness resulting from the compressed gas. Make it possible.

As already explained, the compressor motor comprises two stator parts 5, 6 and a magnet of armature 22. The magnetic interaction of these stators 5, 6 and armature 22 creates a reciprocating force on the piston.

Alternating current (not necessarily sinusoidal) in the stator coils causes real movement of the piston relative to the cylinder assembly casting when the natural frequency is close to the natural resonant frequency of the mechanical system. This vibration force creates a reaction force on the stator part. The stator parts 5, 6 are thus fixedly attached to the cylinder assembly by clamp springs 87.

In the present invention, the main spring 15 has a very large rigidity than that of the effective cylinder spring. The main spring raises the "second" mode frequency above the "third" mode frequency so that the "gas spring" further separates these mode frequencies.

The actual operating frequency (“second” mode frequency) is determined by the complex relationship between the mass of the piston and the cylinder and the rigidity of the cylinder spring and main spring 15. Also, at high discharge pressures, the stiffness of the compressed gas must be added to the stiffness of the main spring. However, for cylinder springs that are fairly soft (for example with one tenth of the stiffness of the main spring), the operating frequency can be determined quite accurately by

In addition to the basics of piston / cylinder motion, external vibrations from other sources can be almost eliminated by making the cylinder springs relatively soft and by reducing the vibration mass. The stiffness of the cylinder spring does not have a specific cylinder spring, and the inherent stiffness of the discharge tube 18 (typically approximately 1000 N / m) (or the stiffness of both the discharge tube and the cooling tube, if a refrigeration tube is used) , By using 2000 N / m).

In order for the compressor to resonate at approximately 75 Hz at a piston mass of approximately 100 g and a cylinder-to-piston mass ratio of 10 to 1, the main spring k _main needs approximately 40000 N / m. Typically the value of the gas spring is less than the value of the main spring but is not significantly smaller. In this case the operating frequency can be expected to be 99 Hz if the gas spring (k _gas ) is operating at approximately 15000 N / m.

[Main Spring]

Higher operating frequencies reduce motor size but require higher spring stiffness, which in turn requires higher stresses on the springs. Thus, the use of the highest quality spring material is important with regard to compressor life. In the past, main springs made of spring steel sheets which have been pressed are mainly used. However, in press operations the cut edges require careful polishing to recover the fatigue strength inherent in the spring steel sheet.

In a preferred embodiment of the present invention, the main spring is formed of a piano wire of circular cross section, which has a very high fatigue strength and does not require subsequent polishing.

Preferred embodiments for the main spring are shown in FIGS. 1 and 2. The spring is a continuous loop twisted by double helix. The length of the wire forming the spring 15 has a free end fixed inside the mounting bar 43 with a lug 42 for mounting to one of the compressor parts. As can be seen in FIG. 3 and the above description, these lugs 42 are mounted in the cylinder part 1, more specifically in the slot of the leg 41. The spring 15 has an additional mounting point 62 for mounting to another compressor part, as can be seen in FIG. 3 and the above description. In a preferred form, it is connected to the piston rod 47 via a molded plastic button 25. The spring 15 comprises a pair of bends 63, 64 of generally the same radius of curvature, each of which passes around the cylinder mounting bar. Each of these curved portions extends in length over approximately 360 degrees. Each bend is gently curved at both ends. In the mounting bar transitions 65 and 66, the curves curve so that the lengths 67, 68 of the curves are aligned radially in the cylinder mounting bar. The sharper curves 65 and 66 are selected to maintain a generally uniform stress distribution along the transition. The alignment of the cylinder mounting ends 67, 68 is then aligned with the cylinder mounting bar. The curved portions 63 and 64 of the constant curvature of the spring 15 can have any length thereof. In the disclosed embodiments, each of them is approximately 360 ° in length.

In the manner disclosed in FIGS. 1 and 2, the mounting bar 43 of the spring 15 is located at the far side of the spring. The central mounting point 62 is located near the spring. Each of the constant curves 63, 64 is gently curved at its lower end so as to be radially aligned and continuous with each other across the diameter of the plain circle of the spring at the mounting point 62. This diameter alignment is generally perpendicular to the alignment of the ends 67, 68 in the cylinder portion mounting bar 43.

The curved portions 63, 64 of constant radius are in a torsional state by the displacement of the piston mounting point 62 relative to the mounting bar 43. Because of the constant radius the torsional stress along each of the bends 63, 64 is also generally constant. By orienting radially or generally radially at the cylinder mounts 67, 68 and the piston mount point 62, the torsional stress when the end of the spring is fitted into the mounting bar 43 and at the piston mount point 62. And improves the mountability of the spring 15 to both the cylinder portion and the piston portion.

[Application of Linear Compressor]

The linear compressor according to a preferred embodiment of the present invention is mainly used for the adoption in domestic refrigeration systems such as a refrigerator, a refrigerator or a refrigerator / freezer combined apparatus. This compressor can directly replace other types of compressors, such as rotary crank compressors. The linear compressor receives the vaporized refrigerant in the low pressure state through the suction tube 12 and discharges the compressed refrigerant in the high pressure state through the discharge stub 13. In refrigeration systems, the discharge stub 13 is generally connected to a condenser. Suction tube 12 is connected to receive the vaporized refrigerant from one or more evaporators. The liquid refrigerant delivery stub 14, as already described, receives the condensation refrigerant from the condenser (or from the accumulator or from the refrigerant line after the condenser) for cooling the compressor. A process tube 16 extending through the hermetic casing is also included for use in emptying the refrigeration system and charging optional refrigerant.

The configuration of the refrigeration system including the compressor does not form part of the present invention. Many useful refrigeration system configurations are well known and available. The compressor according to the invention is also useful in any such system.

The usefulness of the compressor according to a preferred embodiment of the invention is not limited to refrigeration systems for domestic refrigerators and freezers, but can also be used for refrigeration systems for air conditioners. It can also be used as a gas compressor for non-refrigerant applications, in which case the liquid refrigerant delivery for cooling can be eliminated.

Operation and control of the linear compressor is carried out by proper excitation of the stator parts 5, 6. The power supply connector 17 is provided to fit in the opening 19 via the hermetic casing 30. Suitable control systems for excitation of the windings do not form part of the present invention, and many alternative drive systems for brushless DC motors are well known and available. Suitable drive systems are further described in PCT / NZ00 / 00105.

Many advantages of a linear compressor according to a preferred embodiment of the present invention have been outlined in the detailed description above. Additional benefits not yet open are:

(a) Elastically connecting the piston connecting rod to the piston and to the mounting leg 41 of the cylinder casting 33 via the main spring, while lightening the piston and appropriately selecting materials for the piston sleeve and cylinder liner. Connecting eliminates the need for additional lubricant in the compressor. Thus, in a refrigeration system, the harmful presence of lubricating oil in the refrigerant or the need for a separator or accumulator for lubricating oil is eliminated.

(b) Lightening the piston and connecting rod assembly reduces the magnitude of the vibratory momentum of the two main compressor structures and consequently reduces the vibration transmitted to the hermetic casing.

(c) The main spring is formed of a material that has a high fatigue life and does not require further polishing after formation, and the main spring shape and configuration are characterized by the use of materials, the weight of the spring and the spring for the required spring strength. Minimize the size.

(d) Positioning the armature magnets on the piston connecting rods with direct operation of the piston connecting rods within the stator air gaps provides a lateral compact linear motor for accurate and effective alignment of the armature operation between the two parts of the stator. To provide.

(e) Flexible connection between the piston and the piston connecting rod and from the connecting rod to the cylinder portion ensures that the piston sleeve does not apply local pressure inside the cylinder bore due to inaccurate and non-axial orientation.

(f) The use of isobutane refrigerants in linear compressors provides a synergistic advantage compared to the use of refrigerants having a higher specific heat ratio.

(g) Delivering liquid into the cooling jacket surrounding the cylinder provides evaporative cooling to the combined compressed gas stream leaving the cylinder and cylinder head assembly and the compressor. Discharge to high pressure gas in the jacket is achieved without the need for an active pump device.

(h) The whole compressor configuration is well suited for use in household products. The dimensions of the machine space for the refrigerator or freezer can be reduced due to the low overall height of the hermetic casing. In view of the internal space, the machine compartment generally extends completely across the width of the installation, regardless of the components that need to be accommodated. In general, however, its height and depth are determined by the maximum height and depth requirements in fitting the compressor. The low profile of the linear compressor according to the invention contributes by the components of the compressor to be arranged end to end rather than concentrically.

(i) The motor part and the main spring of the compressor are subjected to elastic preloads to the extent determined by the geometry, size and shape of the part. Due to incomplete assembly, due to the usual manufacturing tolerances, or due to the influence of motor operation, the operation of the compressor reduces or eliminates the possibility of loosening of parts and thereby knocking, rattling or failure.

(j) By integrating the piston rod into the main spring in a hot plate welding process in the above manner, the piston face is accurately positioned with respect to the neutral position of the main spring. Thus, the compressor can be operated to surely have a low cylinder head clearance that does not strike the cylinder head at the forward end of its run.

Claims (3)

As a method of manufacturing a linear compressor, in order to connect the piston of the linear compressor to the spring of the linear compressor, (a) placing the piston in a predetermined axial position in the bore of the cylinder of the compressor; (b) placing a piston connection point of the spring already connected with the cylinder portion at a position based on a predetermined displacement of the spring or at a position at which the spring exerts a predetermined reaction force; And (c) connecting the piston and the piston connection point of the spring at fixed axial spacing fixed according to the spacing obtained through steps (a) and (b), Wherein each step may be performed in any order. 2. The method of claim 1, wherein step (c) includes incorporating an extension from the piston to a mounting button on the spring, wherein the step of incorporating the extension in the axial dimension of the button is a preliminary assembly dimension. Method for manufacturing a linear compressor, characterized in that is modified from. 3. The method of claim 1 or 2, wherein step (b) places the piston at a position in accordance with the displacement beyond any normal operating position of the piston such that an element connected to the piston is engaged with the element of the cylinder portion. Placing the element of the piston portion at a predetermined position in a plane perpendicular to the reciprocating axis of the piston in the cylinder.

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