Differential pressure lubrication system for rolling piston compressor

The lubricating oil pool 11 in the bottom of the shell 1 of a rolling piston refrigerant gas compressor is communicated directly with the space 16 inside the cylindrical piston 4 via a supply passage 10c in a side plate 10. Sufficient clearance is provided between the ends of the piston and the compressor side plates 9, 10 to enable limited communication between the space 16 and the compression and suction chambers 6, 18. The compressor discharge is supplied to the space 7 within the shell, and the resultant differential pressure applied to opposite ends of the supply passage causes a steady flow of oil into the piston interior to properly lubricate the moving parts of the compressor.

BACKGROUND OF THE INVENTION 
This invention relates to an improved differential pressure lubrication 
system for an eccentric rolling piston, sliding vane type of fluid 
compressor, as particularly used to compress refrigerant gases in 
refrigerators and air conditioners. 
In conventional units of this type an electric motor and a rolling piston 
compressor driven thereby are mounted within a sealed pressure shell or 
casing. Refrigerant gases drawn in from an external accumulator or the 
like are compressed and discharged into the space within the shell, from 
which they flow to a condenser, evaporator or the like. A pool of 
lubricating oil is maintained within the shell and its surface is in 
direct contact with the high pressure discharge from the compressor. An 
oil flow path is established to properly lubricate the rolling and sliding 
friction members of the compressor such that the high pressure or supply 
end of the path is simply immersed in the pool of oil while the low 
pressure or return end is communicated with a suction passage of the 
compressor. The resultant differential pressure between the supply and 
return ends of the path establishes a steady flow of lubricating oil 
through the frictional members of the compressor. Such a relatively high 
differential pressure often produces an attendantly excessive flow of 
lubricating oil, however, which unduly loads the compressor, generates 
vibrations, results in an excessive amount of lubricating oil being 
entrained in the refrigerant fluid, etc. 
In an effort to solve this "over-lubrication" problem, as disclosed in 
laid-open Japanese patent application No. 131393/83 and as shown in FIG. 
1, the low pressure or return side of the oil supply passage is 
communicated with the compression chamber of the compressor in order to 
reduce the overall differential pressure to which the lubrication system 
is subjected. More specifically, by the action of an electric motor 2 
mounted in a sealed shell 1, a crankshaft 3 is rotated to reduce the 
volume in a compression chamber 6 defined between a rolling piston 4 and a 
cylinder 5 to thereby compress refrigerant gases drawn in from an 
accumulator or the like, not shown. The compressed gases are released into 
the space 7 within the shell from which they are supplied to a condenser 
or the like via a discharge outlet 8. The lubricating oil 11 enters the 
compressor through a passage 9c formed in a side plate 9 and lubricates, 
in succession, bearing 9a adjacent end seal 12, eccentric 3a and bearing 
10a in side plate 10. The oil then flows into the compression chamber 6 
through a return passage 13 in the side plate 10, from which it is 
discharged together with the compressed gas into the space 7 within the 
shell and falls back into the supply pool. The bearings 9a, 10a have a 
relatively large clearance as exaggeratedly shown in FIG. 1 to establish a 
sufficient flow path for the oil, while the tolerance or clearance between 
the ends of the piston 4 and the side plates 9, 10 is relatively close to 
thereby effectively isolate the space 16 within the piston from the 
compression chamber 6. The necessary lubricating oil is supplied to the 
latter through the return passage 13. 
Since the mean or average pressure in the compression chamber 6 lies 
between the suction pressure and the discharge pressure, with the latter 
being applied directly to the surface 11a of the oil pool, the 
differential pressure applied to the opposite ends of the oil flow path is 
thus considerably lower than in the more conventional arrangement 
described above, and this attendantly reduces the oil flow rate to thereby 
avoid such problems as undue loading, vibration, etc. 
A disadvantage with the FIG. 1 approach is that the pressure at the bearing 
end 10b of the side plate 10 must be isolated from the discharge pressure 
within the space 7 in the shell. This requires a mechanical seal 14 which 
not only adds to the production cost, but also increases the mechanical 
loss due to friction and represents a further source of wear and 
deterioration. A further disadvantage is that the oil flow path includes 
successive restrictions represented by the bearing 9a, the clearance 
between the eccentric and the inner surface of the piston 4, and the 
bearing 10a, and even a partial blockage at any one of these points can 
result in overheating, seizure, and the destruction of the entire 
compressor unit. 
SUMMARY OF THE INVENTION 
The present invention seeks to effectively avoid the drawbacks and 
disadvantages of the prior art as discussed above by providing a 
simplified and cost effective differential pressure lubrication system for 
a rolling piston compressor wherein the exit or return end of the oil 
supply passage is communicated directly with the circumferential space 
within the piston flanking the eccentric. The crankshaft bearings within 
the side plates are provided with closer tolerances than in the prior art 
to prevent any excessive outward flow of lubricating oil therethrough, and 
the clearances between the ends of the piston and the side plates are 
established at a sufficient value to enable an adequate flow of oil into 
the suction and compression chambers while still ensuring a sufficient 
compression seal. Such an arrangement eliminates the need for any bearing 
and shaft seals, thereby reducing the cost and complexity of the 
compressor.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
Referring now to a first embodiment of the invention as illustrated 
schematically in FIGS. 2 and 3, wherein like reference numerals are used 
to designate the same structural elements as shown in FIG. 1, a sliding 
vane 17 separates the compression chamber 6 and a suction chamber 18 
within the cylinder 5, the former communicating with a discharge orifice 
22 and the latter communicating with a suction inlet 19. The vane is 
reciprocated by the outer surface 4a of the eccentrically driven rolling 
piston 4. An oil supply passage 10c defined in the side plate 10 has its 
lower end in direct communication with the oil pool 11 and its upper end 
in direct communication with the inner circumferential space 16 within the 
piston. The side plate bearings 9a, 10a are machined to closer tolerances 
than those of FIG. 1 to limit the outward flow of oil therethrough, and 
the clearances between the ends 4b, 4c of the piston and the side plates 
9, 10 are established at a sufficient level or value, on the order of 
several tens of microns, to enable a sufficient passage of lubricating oil 
between the space 16 and the compression and suction chambers 6, 18 while 
still maintaining an adequate compression seal. 
With such a construction the discharge pressure in the space 7 within the 
shell forces the oil up through supply passage 10c and into the space 16, 
whose pressure takes a level between the suction and discharge pressures 
owing to the limited communication with the compression and suction 
chambers 6, 18 via the clearances at the ends 4b, 4c of the piston. The 
oil thus drawn into the space 16 effectively lubricates the side plate 
bearings 9a, 10a as well as the contact surfaces between the eccentric 3a 
and the inside of piston 4, and small but sufficient amounts of such oil 
are also "pumped" into and out of the compression and suction chambers to 
coat them with a thin film and thereby lubricate their surfaces. Some of 
the lubricating oil will pass from the space 16 into the shell space 7 
through the side plate bearings 9a, 10a, while greater quantities of oil 
will exit the compression chamber 6 through the discharge orifice 22 in a 
fine mist. These minute oil particles or droplets condense into larger 
particles due to the high pressure level in the space 7 and fall back into 
the pool 11. Some small quantities of the oil mist will unavoidably be 
entrained in the compressed refrigerant gas exiting through the discharge 
outlet 8, but this is common and does not appreciably detract from the 
system performance. If necessary or desired a downstream separator can be 
provided in the system to filter out and return such oil particles. 
FIG. 4 shows in greater detail a side plate 10 and crankshaft 3 journaled 
therein for use in the schematic embodiment of FIGS. 2 and 3, although the 
presentation of FIG. 4 is reversed or as viewed from the back side of FIG. 
2. The upper end of the oil supply passage 10c terminates in a recess or 
pit 10d in the side plate 10, the eccentric 3a is provided with an oblique 
or helical groove 3b, and a portion of the crankshaft disposed within the 
side plate bearing 10a is provided with a similar oblique or helical 
groove 3c. The pit 10d and groove 3c facilitate the lateral dispersion of 
lubricating oil throughout the bearing 10a since one end of the groove 3c 
comes into direct communication with the pit during each rotation of the 
crankshaft. Although not clearly visible in FIG. 4, the pit 10d also opens 
directly into the space 16 within the piston 4 on the right side of the 
eccentric as viewed in FIG. 4; the groove 3b facilitates the distribution 
of the lubricating oil to the space 16 on the left side of the eccentric 
and thence to the opposite side plate bearing 9a. 
The embodiment of FIGS. 5 and 6 is characterized by the stator 2a of the 
electric drive motor being axially displaced from the rotor 2b a distance 
1, by the crankshaft groove 3c extending to a distance m from the bearing 
end 10b of the side plate, and by a thrust bearing or pedestal 3d being 
formed on the end of the eccentric adjacent the side plate 9. With such a 
construction the axial offset between the rotor and stator of the drive 
motor generates a thrust force in the direction indicated by the arrow in 
FIG. 5, and such force is borne by the thrust bearing 3d. This arrangement 
ensures that the crankshaft is constantly urged against the side plate 9, 
which effectively suppresses any vibrations and attendant noise which 
might be generated by the axial freedom and movement of the crankshaft. 
The groove 3b in the eccentric is extended into the thrust bearing 3d to 
ensure the proper lubrication of the face thereof and to implement the 
lateral distribution of the oil to the side plate bearing 9a. Moreover, 
the extension of the crankshaft groove 3c to the distance m from the 
bearing end 10b ensures the full and effective lubrication of the side 
plate bearing 10a. 
FIG. 7 shows a construction of the side plate 9 wherein a helical groove 9d 
is formed in the bearing portion 9a and extends to a distance n from the 
bearing end 9b to ensure the proper lateral distribution of the 
lubricating oil. As an obvious alternative, a groove corresponding to 9d 
could instead be provided on the left end of the crankshaft as viewed in 
FIG. 6, similar to the groove 3c. 
FIG. 8 shows a modification wherein the interior or bearing surface of the 
rolling piston 4 is provided with a plurality of helical grooves 4d to 
replace the groove 3b in the eccentric. 
In the embodiment of FIGS. 9-11 the crankshaft is provided with a central 
coaxial bore 3e extending from the compressor end thereof to a point just 
beyond the bearing end 10b whereat radial outlet ports 3f are provided, 
and a cap 20 is fitted over the side plate 9 to enclose both the bearing 
boss of the latter and a discharge valve 21 communicating with the 
compression chamber 6 via the discharge orifice 22. This establishes a 
high speed flow of the compressed refrigerant gas through the crankshaft 
bore 3e and out the radial ports 3f along the path shown by the arrows. 
With the cap 20 disposed in close proximity to the bearing end 9b of the 
side plate a high velocity flow is established into the bore 3e as seen in 
FIG. 10, and in a similar manner with the ports 3f having a sufficiently 
small diameter a corresponding high velocity gas flow is also established 
across the bearing end 10b of the opposite side plate. If the bearing ends 
of the respective side plates are now provided with chamfers 9e and 10e as 
shown in FIGS. 10 and 11 surrounding the crankshaft, the high velocity gas 
flows induce low pressure regions in the chamfer recesses and this assists 
in drawing out lubricant from the ends of the grooves 9d and 3c to ensure 
a steady supply of oil to the ends of bearings 9a and 10a. 
As will be obvious to those skilled in the art, the principles of this 
invention are equally applicable to both horizontally and vertically 
oriented compressors although only the former have been shown in the 
drawings by way of example. In the case of a vertically oriented 
compressor the side plate 10 would be disposed above the surface 11a of 
the oil pool, and the supply passage 10c would simply be extended by a 
tube leading downwardly and terminating in the pool. As is also obvious, 
the oil supply passage could just as well be provided in the side plate 9, 
or for that matter a passage could be provided in both side plates. Such 
passage could also be provided by a separate length of tubing extending 
from the oil pool through one of the side plates and into the space 16. 
By way of representative example, the clearance between the crankshaft and 
the side plate bearings 9a, 10a may be on the order of 10.about.20 
microns, and that between the ends of the piston 4 and the side plates may 
be on the order of 3.about.30 microns.