Patent Abstract:
A linear abstract includes a cylinder part and a piston part. A spring connects the cylinder part and the piston part and operates in the direction of reciprocation of the piston part relative to the cylinder part. The piston part includes a radially compliant but axially stiff linkage and a piston. The cylinder part includes a cylinder and a cylinder liner therewithin. A bore runs through the cylinder liner and the piston reciprocates in the bore. Gas bearing passages are formed between the cylinder and the cylinder liner leading to openings through the wall of the cylinder liner to the bore. A gas bearing manifold receives compressed gases and supplies the compressed gases to the gas bearing passages. The gas bearing passages follow a tortuous path to the openings.

Full Description:
BACKGROUND TO THE INVENTION 
   Field of the Invention 
   The present invention relates to a linear compressor, particularly but not solely for use in refrigerators. 
   SUMMARY OF THE PRIOR ART 
   Compressors, in particular refrigerator compressors, are conventionally driven by rotary electric motors. However, even in their most efficient form, there are significant losses associated with the crank system that converts rotary motion to linear reciprocating motion. Alternatively a rotary compressor which does not require a crank can be used but again there are high centripetal loads, leading to significant frictional losses. A linear compressor driven by a linear motor would not have these losses, and can be designed with a bearing load low enough to allow the use of aerostatic gas bearings as disclosed in U.S. Pat. No. 5,525,845, where a connecting rod that is compliant to lateral movement allows for the low bearing load. 
   A discussion of aerostatic gas bearings is included in “Design of Aerostatic Bearings”, J W Powell, The Machinery Publishing Company Limited, London 1970. However with normal manufacturing tolerances and equipment production of effective gas bearings is difficult. 
   Conventional compressors are mounted within a hermetically sealed housing which in use acts as a reservoir of refrigerant gas. Refrigerant gas is drawn into the compressor from this reservoir and is exhausted through an exhaust conduit leading from the compressor, through the housing. 
   Operation of the compressor involves the reciprocation of moving parts leading to vibration of the compressor unit, in all three axis. To reduce the external noise effect of this vibration the compressor is mounted on isolation springs within the sealed housing. 
   With a linear compressor the piston vibrates relative to the cylinder in only one axis, with consequent reaction forces on whichever part, if either, is fixed. One solution proposed to this problem is to operate a pair of compressors synchronously in a balanced and opposed configuration. However this arrangement would be too complex and costly for use in a commodity item such as a domestic refrigerator. Another proposed solution is the addition of a resonant counterweight to reduce the vibration. However this approach limits the operation of the compressor because the counterweight is a negative feedback device and is limited to the fundamental unbalance force. A further solution is proposed in “Vibration characteristics of small rotary and linear cryogenic coolers for IR systems”, Gully and Hanes, Proceedings of the 6 th  International Cryocooler Conference, Plymouth, Massachusetts, 1990. This solution involves independently supporting the piston part and the cylinder part of the compressor within the housing so that the “stator acts as a counterweight”. However in implementing this design in a domestic refrigerator there is a problem 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 force (the “gas spring”) which increases the running speed as the discharge pressure increases. This is a problem because the frequency of the “third” vibration mode (where the piston and the cylinder vibrate in phase with each other but out of phase with the compressor shell) is only slightly above the frequency of the desirable “second” mode (where the shell does not vibrate and the piston and cylinder are out of phase). Thus the shell starts to vibrate intolerably as the “gas spring” starts to operate and effectively raises the “second” mode frequency to, and eventually above, the “third” mode frequency. 
   SUMMARY OF THE INVENTION 
   It is an object of the present invention to provide a compact linear compressor which goes some way to overcoming the abovementioned disadvantages. 
   In one aspect the invention consists in a linear compressor including;
         a cylinder part including a head, a cylinder and a cylinder liner within said cylinder, said cylinder liner having a bore therethrough and openings through said cylinder liner into from an outside surface of said cylinder liner to said bore,   a piston part,   a linear motor configured to operate between said piston part and said cylinder part,   a main spring connecting between said cylinder part and said piston part and operating in the direction of reciprocation of said piston part relative to said cylinder part,   said piston part including a radially compliant but axially stiff linkage and a piston, said linkage connecting between said piston and said main spring, said piston moving within said bore of said cylinder liner,   a gas bearing manifold adapted to receive a supply of gases compressed by said compressor,   wherein an outer surface of said cylinder liner mates against an inner surface of said cylinder,   one said surface having one or more grooves extending in a tortuous path from said gas bearing manifold to each of said openings through said cylinder liner, said grooves enclosed by the other said surface to constitute passages.       

   In a further aspect the invention consists in a free piston compressor having:
         a cylinder outer part,   a cylinder inner part within the cylinder outer part and having a bore therethrough,   a piston reciprocable within said bore, and   openings through said cylinder inner part from an outer surface to said bore, said openings distributed to, with a flow of gases therethrough in use, provide gas bearing support to said piston,   the improvement comprising:
           a gas bearing supply manifold at, or adjacent, the interface between said cylinder inner part and said cylinder outer part,   an inner surface of said cylinder outer part mating against said outer surface of said cylinder inner part, and   grooves on at least one of said inner surface and said outer surface extending in a tortuous path from said manifold to said openings said grooves enclosed by the other said surface to constitute enclosed passages.   
               

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is a cross-section of a linear compressor according to the present invention, 
       FIG. 2  is a plan view of a first embodiment of a main spring for the linear compressor of  FIG. 1 , 
       FIG. 3  is a plan view of an alternative second embodiment of the main spring for the linear compressor of  FIG. 1 , 
       FIG. 4  is a perspective view of a preferred third embodiment of the main spring of the compressor for  FIG. 1 , 
       FIG. 5  is a perspective view of the main spring of  FIG. 4  from an alternative direction, 
       FIG. 6  is a perspective view of a cylinder liner according to one preferred form of the present invention, and 
       FIG. 7  is a perspective view of a cylinder liner according to a second preferred form of the present invention. 
   

   DETAILED DESCRIPTION 
   A practical embodiment of the invention, shown in  FIG. 1 , involves a permanent magnet linear motor connected to a reciprocating free piston compressor. The cylinder  9  is supported by a cylinder spring  14  and by a discharge tube  18  within the compressor shell  30 . The piston  11  is supported radially by the bearing formed by the cylinder bore plus its spring  13  via the spring mount  25 . A main spring  15  connects between the piston part  11  and the cylinder part  9 . The total reciprocating movement is the sum of the movement of the piston  11  and the cylinder  9 . 
   This reciprocating movement draws gas in through a suction tube  12  through a suction port  26  through a suction muffler  20  and through a suction valve port  24  in a valve plate  21  into a compression space  28 . The compressed gas then leaves through a discharge valve port  23 , is silenced in a discharge muffler  19 , and exits through a discharge tube  18 . 
   The cylinder  9  is supported by the discharge tube  18  and the cylinder spring  14  which have a combined stiffness, k cylinder , in the axial direction. The piston  11  is supported radially by gas bearings which will be described later. During resonant oscillation of the piston and cylinder the main spring has a stiffness, k main , such that the second mode resonant frequency, f natural , can be estimated from the relation, 
         f   natural     =       1     2   ·   π       ·         k   main     ·         m   piston     +     m   cylinder           m   piston     ·     m   cylinder                   
 
   Where m pistons , m cylinder , are the sprung masses of the piston and cylinder springs, f n ,f natural  is usually 10 to 20 Hz less than the desired running frequency to allow for the increase in frequency due to the stiffness of the compressed gas, the effective cylinder spring (a combination of spring  14  and  18 ), and piston spring  13 . The stiffness of the piston spring k piston  is selected according to the relationship 
         k   piston     =       k   cylinder     ×       m   piston       m   cylinder             
 
   The spring forces are transferred to the piston via the rod end  25  and the radially compliant piston rod  124 . The electromagnetic forces are transferred to the piston via the piston flange  7 , from the bi-polar magnets  22 . The bi-polar magnets  22  are bonded to each other and to the piston flange  7 . 
   The compressor motor comprises a two part stator and an armature. The stator includes an inner stator  6  and a back iron  5 . The inner stator carries coils  1  and  2 . The armature includes bi-polar magnets  22 . The magnetic interaction of the stator  5 ,  6  and armature magnets  22  generates reciprocating force on the piston  11  (attached to the armature by flange  7 ). 
   An oscillating current in coils  1  and  2 , not necessarily sinusoidal, will give rise to substantial movement of the piston  11  relative to the cylinder  9  provided the oscillation frequency of the current is close to the natural resonant frequency of the mechanical system. This oscillating force creates a reaction force on the stator parts. Thus the inner stator  6  must be rigidly attached to the cylinder  9  by adhesive, shrink fit or clamp etc. The back iron  5  is clamped or bonded to the stator mount  17 . 
   The stator mount  17  also clamps the outer ends of the main spring  15  and also keeps the relatively weak back iron  5  round and concentric with the inner stator  6 . The entire compressor assembly is hermetically sealed inside the compressor shell  30 . 
   In the present invention it is proposed that the main spring  15  has a stiffness much greater than the stiffness of the effective cylinder spring, and of the piston spring. This “main spring” raises the “second” mode frequency above the “third” so that the “gas spring” then only separates the modal frequencies further. 
   The actual running frequency (the “second” mode frequency) is determined by a complicated relation of using the mass of piston and cylinder and by the stiffness of the piston spring, cylinder spring, and main spring  15 . Also when the discharge pressure is high the equivalent spring stiffness of the compressed gas must be added to that of the main spring. However, with the cylinder spring quite soft (say with a stiffness 1/100 of the main spring) the running frequency is found reasonably accurately by: 
         f   running     =       1     2   ·   π       ·         (       k   main     +     k   gas       )     ·         m   piston     +     m   cylinder           m   piston     ·     m   cylinder                   
 
   External vibration due to sources, other than from the desirable second mode due to piston/cylinder movement, can be almost eliminated by reducing the oscillating mass and by ensuring that the piston and cylinder springs are relatively soft. The effective cylinder spring stiffness can be reduced to a minimum by having no cylinder spring at all, leaving only the inherent stiffness (from around 1000 N/m) of the discharge tube  18  (or where a cooling tube is used the stiffness of both discharge and cooling tube are combined ie 2000 N/m). With the effective cylinder spring stiffness only including the stiffness of the discharge tube (say  1000 N/m) the stiffness of the piston spring should be: 
         k   piston     =         m   piston       m   cylinder       ×   1000         
 
   For a ten to one cylinder to piston mass ratio this suggests a very soft piston-spring (100 N/m). 
   For the compressor with a main spring to resonate at roughly 75 Hz with a piston mass of around 100 g and a ten to one cylinder to piston mass ratio, the main spring stiffness (K main ) needs to be about 20,000N/m. Typically the value of the gas spring will be lower-than that of the main spring but not substantially lower. In the above case the running frequency is expected to be 99 Hz with the gas spring (k gas ) of approximately 15,000N/m. 
   The piston  11  is supported radially within the cylinder by aerostatic gas bearings. 
   The cylinder part of the compressor includes the cylinder  9 , having a bore therethrough, and a cylinder liner  10  within the bore. The cylinder liner  10  may be made from a suitable material to reduce piston wear. For example it may be formed from a fibre reinforced plastic composite such as carbon fibre reinforced nylon with 15% PTFE (preferred), or may be cast iron with the self lubricating effect of its graphite flakes. Referring additionally to  FIGS. 6 and 7 , the cylinder liner  10  has openings  31  therethrough, extending from the outside cylindrical surface  70  thereof to the internal bore  71  thereof. The piston  11  travels in the internal bore  71 , and these openings  31  form the gas bearings. A supply of compressed gas is supplied to the openings  31  by a series of gas bearing passages  8 . The gas bearing passages  8  open at their other ends to a gas bearing supply manifold  16 , which is formed as an annular chamber around the cylinder liner  10  at the head end thereof between the liner  10  and the cylinder  9 . The gas bearing supply manifold  16  is in turn supplied by the compressed gas-manifold  20  of the compressor head by a small supply passage  73 . The small size of the supply passage  73  controls the pressure in bearing supply manifold  16 , thus limiting the gas consumption of the gas bearings. 
   The gas bearing passages  8  are formed as grooves  80  or  81  in either the bore  74  of the cylinder or in the outer wall  70  of the cylinder liner. These grooves  80  or  81  combine with the wall of the other cylinder or the cylinder liner to form enclosed passages  8  leading to the openings  31 . It will be appreciated that while the grooves could be provided in either part they are more readily formed in the liner part than in the cylinder part, being on an outer surface rather than an inner surface. Being able to machine the grooves into a surface of one or other part rather than having to drill or bore passages is a significant manufacturing improvement. 
   It has been found that the pressure drop occurring in the gas bearing passages needs to be similar to the pressure drop occurring in the exit flow between the piston and the bore of the cylinder liner. Since the gap between the piston  11  and the cylinder liner bore  71  (for an effective compact compressor) is only 10 to 15 microns, the sectional dimensions of the passages  8  need to be very small, for example, 40 microns deep by 120 microns wide. These small dimensions make manufacturing the bearing passages difficult. 
   However, with reference to  FIGS. 6 and 7 , in the preferred embodiment of the present invention matching the pressure drops is made easier by increasing the length of the passages  8  so that the cross-sectional area of the passages can also be increased. The longer but larger cross-section passages have a flow resistance similar to narrower shorter passages. Taking the earlier examples, the dimensions might become 70 microns deep by 200 microns wide. This takes advantage of the ability to form grooves  80  or  81  of any appropriate shape in the surface of the liner part  10  or of the cylinder part  9  which then forms the passages  8  in conjunction with the other part. The grooves can be formed having any path, and if a tortuous path is chosen the length of the grooves can be significantly greater than the direct path between the gas bearing supply manifold and the respective gas bearing forming openings. Two possible options are depicted in  FIGS. 6 and 7 , being helical paths  80  and serpentine paths  81  respectively. The lengths of the respective paths are chosen in accordance with the preferred cross-sectional area of the passage, which can be chosen for easy manufacture (either machining or possibly by some other form such as precision moulding). 
   Higher running frequencies reduce motor size but require more spring stiffness, and consequently higher stresses in the springs. Thus it is important for compressor longevity that the highest quality spring material be used. In the conventional linear compressors main springs made from pressed spring steel sheet are often used. However, the edges cut in the pressing operation require careful polishing to regain the original strength of the spring steel sheet. 
   In the preferred embodiment of the present invention the main spring is formed from circular section music wire. As depicted in first embodiment  FIG. 2  the main spring can be wound to form a spiral spring  15 . The spiral spring  15  has a pair of spiral arms  50 , 51  which are 180 degrees out of alignment so that the path of each arm is between adjacent turns of the other arm. The piston mounting point  52  is at the centre of connecting bridge  53  at the centre and the cylinder mounting point  54  for each arm of the spring at the outer end of the arm. 
   The very high fatigue strength of music wire is utilised effectively and there is no need for a subsequent polishing operation. If increased lateral stiffness is required the music wire could be deformed by 10% to give an elliptical section. To simplify the attachment of the main spring, square section wire could be used, or the connection ends of the spring may be stamped to a flattened shape, as depicted. 
   However, an alternative and second embodiment of the main spring is depicted in FIG.  3 . This spring may also be formed from music wire and take advantage of its high fatigue strength. 
   In  FIG. 3  the spring  59  includes a pair of mounting points  60 , 61  for mounting to one of the compressor parts (the cylinder part) and a central mounting point  62  for mounting to the other compressor part (the piston part). The spring  59  includes a pair of curved sections  63 ,  64  of substantially constant radius of curvature which are each centered on a respective cylinder mounting point  60 , 61 . These sections meet in material continuity at the piston mounting point  62 . Each section curves smoothly at its other end  65 , 66  to be radially aligned with the cylinder mounting point  67 , 68 . The sharper transition curve at  65 , 66  is preferably selected to maintain a substantially even stress distribution along the transition. The cylinder mounting ends  67 , 68  are preferably aligned with the line between the cylinder mounting points  60 , 61 . To get the best performance for the overall space occupied by the spring, the constant curvature sections  63 , 64  of the spring  59  are as long as possible. Consequently they extend for approximately 325 degrees from the piston mounting point  62 , before curving more sharply to the cylinder mounting point  60  or  61  respectively. This configuration allows the spring sections to narrowly avoid interfering with one another. The total spring assumes an approximate figure-eight shape. 
   The constant radius curves  63 , 64  are placed in torsion by the displacement (out of plane) of the piston mounting point  62  relative to the cylinder mounting points  60 , 61 . Being constant radius, the torsion stresses along each of the sections  63 , 64  are also substantially constant. Due to the radial, or substantially radial direction at of the cylinder mounting sections  67 , 68  any torsion stresses in the portion of the spring at the cylinder mounting are at a minimum and mounting of the spring  59  to the cylinder part is improved. The central mounting point  62  of the spring ing has high torsion stresses, however this does not significantly complicate that mounting because that the mounting can be made to encircle the spring arm with a resilient (eg: rubber) boot to allow for movement of the spring arm within the mounting. Movement of the spring arm within the mounting will be cyclical and, due to the symmetry of the spring (the spring is rotationally symmetric through 180 degrees), the cyclic forces should not cause the mounting to creep or walk along the spring arm. It should be noted that this spring configuration has been particularly developed for incorporating the wire formed approach rather than the stamped plate approach. However (subject to limitations in some more complex embodiments referred to below) springs of this geometric form could also be manufactured using the stamped plate method, but some of the advantages (for example, uniform stresses are particularly suitable with wire of constant cross section) would not be realised. 
   It should be appreciated that variations on the spring of  FIG. 3  are also possible without departing from the scope of the invention. In particular, if the spring is formed so that a spring arm is perpendicular to the line between the compressor mountings at the compressor mounting then the arm can continue to form an equivalent (although mirrored) loop, below or above a first loop, back to the other compressor mounting. That loop would of course have a second piston connection point below or above the first loop. At the other compressor connection point the ends can meet, or alternatively this second loop may be continued through the connection point to form a third loop, below (or above as necessary) the second loop, back to the first compressor mounting point (or at least to a mounting point immediately above or below). This chaining of loops can proceed to include as many loops as necessary to achieve a required spring constant. Clearly this is a planar spring configuration that cannot be constructed by stamped plate methods. 
   However the preferred third embodiment for the main spring is depicted in  FIGS. 4 and 5 . 
   In the third embodiment the main spring takes a form other than that of a planar spring. It retains many of the conceptual features of the second embodiment and therefore where similar features are apparent the same reference numerals have been used. 
   The spring  15  has a pair of free ends for mounting to one of the compressor parts, for example the cylinder part. The spring  15  has a further mounting point for mounting to the piston part. 
   The spring  15  includes a pair of curved sections  63 ,  64  of substantially constant radius of curvature which each pass around their respective mounting end. Each of these curved sections extends over a length of approximately 360°. Each section curves smoothly at both of its ends. At the ends  65 ,  66  they curve such that the lengths  67 ,  68  of them at the cylinder mounting ends are radially aligned. The sharper transition curve at  65 ,  66  is selected to maintain a substantially even stress distribution along the transition. The spring  15  of  FIGS. 4 and 5  improves on the spring  59  of  FIG. 3  in that the constant curvature sections  63 ,  64  of the spring  15  may be rendered of any degree length including beyond 360. In the example depicted they are each of approximately 360° in length. 
   In the manner depicted in  FIGS. 4 and 5  the mounting points  60 ,  61  of spring  15  are at an upper side thereof. The central mounting point  62  is at a lower side thereof. The constant curved sections  63 ,  64  each curve smoothly at their lower ends to be radially aligned and continuous with one across a diameter of the general circle of the spring at the mounting point  62 . The alignment of this diameter is substantially perpendicular to the alignment of the ends  67 , 68  at the cylinder part mounting points  60 ,  61 . 
   The constant radius curve  63 ,  64  are placed in torsion by the displacement of the piston mounting point  62  relative to the mounting points  60 ,  61 . The torsion along each of the sections  63 ,  64  is also substantially constant. Due to the radial or substantially radial direction of the cylinder mounting sections  67 ,  68  and the piston mounting point  62 , any torsion stresses at the cylinder mounting ends and at the piston mounting point are at a minimum and mounting of the spring  15  to both the cylinder parts and the piston part is improved.

Technology Classification (CPC): 5