Scroll device with eccentricity adjusting bearing

A scroll-type pumping apparatus comprising a stationary scroll 1, an orbiting scroll 2 attached to an orbiting scroll shaft 4, and a crankshaft 14. An eccentric hole is formed in the end of the crankshaft and an eccentric bearing 26 rotates within this eccentric hole. The eccentric bearing has a central aperture eccentric to the bearing's outer circumference in which the orbiting scroll shaft is placed and freely rotates.

BACKGROUND OF THE INVENTION 
This invention relates to a scroll type positive fluid displacement 
apparatus for compressing, expanding or pumping fluids. 
The principles of operation of a scroll apparatus will be explained with 
reference to FIG. 1, which shows a stationary involute or spiral-shaped 
scroll 1 and an orbiting scroll 2 of like shape but displaced and rotated 
180.degree.. The orbiting scroll 2 performs orbital motion about a point 
without rotation. Thus, a side of the orbiting scroll moves so as to 
always remain in a parallel position. Compression pockets 3 and 5 are 
formed in the space between the stationary and orbiting scroll members 
whose volumes are decreased (assuming compressing operation) during 
orbiting about a central point O as shown in the sequence of FIGS. 
1(a)-1(d) until they merge into a single, similarly, shrinking central 
outlet pocket 8'. At the same time, new inlet pockets are formed as shown 
in FIG. 1(c) at 3, 5, which progressively shrink or are compressed. 
The U.S. Pat. Nos. 3,884,599 and 4,065,279 disclose scroll type positive 
fluid displacement apparatus to ensure axial and radial sealing of the 
scroll assembly. 
The U.S. Pat. No. 1,906,142 shows an exteriorly cylindrical boss for 
permitting the scrolls to move a little in a radial direction on a crank 
pin. 
SUMMARY OF THE INVENTION 
An object of the present invention is thus to provide a new scroll type 
positive fluid displacement apparatus which is able to seal a compression 
pocket in a radial direction and to restrain a driving shaft from 
producing a moment which would incline the driving shaft. 
The object is accomplished by providing a new scroll type apparatus 
comprising a stationary scroll member having a spiral-shaped wrap, an 
orbiting scroll member having a spiral-shaped wrap of the same shape as 
that of the spiral wrap of the stationary scroll member but having a 
rotated orientation, a compression pocket formed by a space between the 
stationary scroll member and the orbiting scroll member, a driving shaft 
having a first eccentric hole formed at its end with a predetermined 
eccentricity, causing the orbiting scroll member to orbit, an eccentric 
bushing rotatably mounted in the first eccentric hole, having a second 
eccentric hole formed in the eccentric bushing with a predetermined 
eccentricity, a pin formed on a surface of the orbiting scroll member on 
an opposite side to the spiral-shaped wrap of the orbiting scroll member 
having a common axis with that of the orbiting scroll member and mounted 
in the second eccentric hole, a bearing mounted within the region radially 
enclosing the bushing, for radially supporting the driving shaft, 
supporting member supporting the bearing, and a rotation preventing means 
for allowing orbital motion without rotation of the orbiting scroll member 
and the supporting member.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
FIG. 2 shows a scroll compressor, which has been developed in the 
corporation to which this application is assigned, wherein a thrust 
bearing 9 supports the back of a base plate 3 of the orbiting scroll 
member 2. A space 12 is formed for an Oldham coupling between the support 
member 10 for the bearing and the base plate 3, such support member being 
bolted or the like to the stationary scroll member 1. The Oldham coupling 
is a well-known mechanism for inducing orbital motion while preventing 
rotation. An oil passage 13 communicates the space 12 with the interior 
space 25 formed between the said supporting member and the motor, and an 
oil hole 15 is formed eccentrically in a driving shaft 14 supported at its 
upper and lower ends by bearings 17, 18. When a stator 19 of a motor is 
energized, the driving shaft 14 is rotated. The orbiting scroll member 2 
which is guided by an Oldham coupling member 11 moves in a orbiting motion 
in accordance with the rotation of the driving shaft 14. 
Reserved oil 24 within a bottom of a sealed or airtight shell 23 is sucked 
upwardly into the oil hole 15 by a centrifugal force due to the rotation 
of the driving shaft 14, and then the sucked oil is supplied to the 
bearings 17, 26, thrust bearing 9, and to the Oldhams coupling 11 and the 
space 12. The supplied oil in the space 12 flows downwardly through the 
oil passage 13 into the space 25, and then falls down into the bottom of 
the shell 23. 
The base plate 3 has a pin 4 at the center of the base plate 3. The pin 4 
is supported by an inner bearing 16 mounted in a hole positioned 
eccentrically with respect to the axis of the driving shaft 14. The 
orbiting scroll 2 as a result accomplishes the compressing process shown 
in FIGS. 1a, b, c and d. 
After the gaseous fluid is taken from an intake tube 7 into the compression 
pockets 3 and 5 through inlet pockets 6 located at circumferential edges 
of the orbiting scroll member 2, the gaseous fluid is exhausted through 
the outlet 8, through an outlet pocket 8', after the pockets 3 and 5 have 
been shifted to the interior of the scroll members 1 and 2. 
In the construction shown in FIG. 2, the entire compressor may vibrate 
because of the unbalanced forces from the orbiting movement of the 
orbiting scroll member 2 as it follows the rotation of the driving shaft 
14. To remedy this problem, a first counter-weight 21 is placed 
eccentrically on the driving shaft 14, and a second counter-weight 22 
placed eccentrically on a rotor 20 of the motor which statically and 
dynamically balance the driving shaft 14 so that the compressor can 
operate without vibration. 
In FIG. 3(a) is shown an enlarged side cross-section of a portion of FIG. 2 
including the stationary scroll 1, the orbiting scroll 2, and the driving 
shaft 14. In FIG. 3(b) is shown a top cross-section of a portion of FIG. 
3(a) along the twisted cross-sectional line IIIb--IIIb of FIG. 3(a), under 
the condition that the orbiting scroll pin 4 is pressed to the bearing 16 
by only the centrifugal force F.sub.c of the orbiting scroll 2 and its 
base plate 3, without accounting for the effect of the force resulting 
from the compressed gas between the elements. A point O.sub.1 represents 
the center of the main bearing 17; a point O.sub.2, the center of the 
driving shaft 14; a point O.sub.3, the center of the inner bearing 16; and 
a point O.sub.4, the center of the orbiting scroll shaft 4. 
The inner bearing 16 and the driving shaft 14 are separated by an 
eccentricity r. The inner diameter of the inner bearing 16 is greater than 
the outer diameter of the orbiting scroll pin 4 by a bearing clearance 
d.sub.1. Likewise, the inner diameter of the main bearing 17 is greater 
than the outer diameter of the crankshaft 14 by a bearing clearance 
d.sub.2. The wrap of the stationary scroll 1 is formed of the vane-like 
continuous protrusion in a spiral shape and the spiral has a radial pitch 
of a distance B. A wrap of the orbiting scroll 2 orbits through a 
horizontal space, defined by actual orbiting diameter D. The wrap of the 
orbiting scroll member 2 is of thickness t. The wrap of the stationary 
scroll 1 does not contact the wrap of the orbiting scroll 2 but is 
separated by gaps C and C.sub.1 in the radial direction. In practice, 
however, C is equal to C.sub.1. 
In the conventional scroll type compressor described above, the actual 
orbiting diameter D may be expressed as 
##EQU1## 
The gap between the wraps of the stationary scroll 1 and the orbiting 
scroll 2, which is C=(B-D)/2, may be alternately expressed by using 
equation (1) as 
##EQU2## 
Generally, since the first term (B-2r-t) in equation (2) representing the 
radial clearance is larger than the second term (d.sub.1 +d.sub.2), the 
sum of the bearing clearances, the gap C always exists. 
In practical operation of the scroll compressor, there is a load force 
F.sub.g on the orbiting scroll pin 4 resulting from the compression of the 
gas. The load force F.sub.g occurs in the direction at a right angle to 
the centrifugal force F.sub.c. The centrifugal force F.sub.c and load 
force F.sub.g combine to form a resultant force F on the orbiting scroll 
shaft 4 to press it in the direction shown in FIG. 4. 
The radial gap C' between the wraps of the stationary scroll 1 and the 
orbiting scroll 2 in the presence of the load force F.sub.g is larger than 
the gap C in the situation in which only the centrifugal force is present. 
Because of the radial gaps C and C' existing between the wraps, the side 
walls of the wraps are not worn by frictional contact between the wraps of 
the stationary and orbiting scrolls 1 and 2. However, since it is 
difficult to seal the radial gap at the end of the compression pockets 3 
and 5, the gas in these compression pockets 5 can leak back toward lower 
pressure such as the inlet side or the lower pressure pocket. Such leakage 
reduces the quantity of gas discharged from the outlet tube 8 and lowers 
the pumping capacity. This leakage also increases the load on the motor 
and decreases the compression efficiency because the leakage gas needs to 
be recompressed. Examples of scroll pumps which attempt to overcome these 
problems are given in U.S. Pat. Nos. 3,884,599 and 4,065,279. 
One possibility of overcoming the above problems is to reduce the 
difference between the sum of the bearing clearances (d.sub.1 +d.sub.2) 
and the radial clearance (B-2r-t). However, the radial clearance term 
includes the manufacturing tolerances for each of the dimensions B, r and 
t, which cannot be precisely controlled, and the bearing clearances 
d.sub.1 and d.sub.2 need to be designed large enough in order to always 
keep (d.sub.1 +d.sub.2) larger than (B-2r-t) regardless of the rotational 
position of the driving shaft. 
But, the bearing clearances d.sub.1 and d.sub.2 should be suitably sized to 
maintain satisfactory lubrication of the bearings and the bearing 
clearances should be no larger than the most suitable value. Therefore 
accurate manufacture is required for the length B between the wraps, the 
eccentricity r and the thickness t of the wraps in order to satisfy the 
condition that (d.sub.1 +d.sub.2) is always larger than (B-2r-t). 
Furthermore, when the center of the stationary scroll 1 deviates for some 
reason from the center O.sub.1 of the main bearing 17, the gaps C and 
C.sub.1 shown in FIG. 3(a) are no longer equal to each other and in the 
extreme case only one of them has an expected large value. It is thus 
impossible to make the radial gaps C.sub.1 and C approach zero, even if 
the value of (d.sub.1 +d.sub.2) is kept larger than (B-2r-t). To eliminate 
the above disadvantage, it is therefore additionally required to 
accurately align the stationary scroll with the center O.sub.1 of the main 
bearing 17. 
The present inventors intended to further develop the scroll compressor 
shown in FIG. 2 into a new scroll compressor that is able to automatically 
seal the compression pockets in the radial direction regardless of error 
of machining or assembly. 
The inventors learned of an eccentric bushing being described in the U.S. 
Pat. No. 1,906,142. That is, a crank pin is formed on the end of a driving 
shaft, an exteriorly cylindrical boss is mounted on the pin, and a hub 
formed at the center of driven member is rotatably mounted on the boss 
which is radially displaceable on the crank pin, so that the piston goes 
in contact with a cylinder wall. 
In FIG. 5, showing a scroll compressor designed according to teachings of 
the U.S. Pat. No. 1,906,142, a pin 51 is eccentrically formed on the end 
of the driving shaft 14 supported by the main bearing 17. 
An eccentric bushing 52 is rotatably mounted on a pin 51 and in a hole 53 
formed at the center of the base plate 3 of the orbiting scroll member 2 
in order to permit the orbiting scroll member 2 to move a little in the 
radial direction around the pin 51. 
Referring to FIG. 5 and FIG. 6 showing an enlarged view of the eccentric 
bushing 52 and the pin 51, during the rotation of the driving shaft 14, 
initially the orbiting scroll member 2 tends to rotate about the pin 51 
and the eccentric bushing 52, though it orbits about a point O.sub.5 of 
the center of the hole 53 without any spin because the Oldham coupling 
member 11 prevents the orbiting scroll member 2 from rotating. 
A radius of the revolution of the orbiting scroll member 2 is represented 
as the distance R between the center O.sub.1 of the driving shaft 14 and 
the center O.sub.5 of the hole 53. 
The radius R is variable, being able to increase toward the direction of 
the centrifugal force F and the radius R is defined as the distance 
##EQU3## 
at which the wall of the orbiting scroll 2 goes in contact with the wall 
of the stationary scroll 1, where P is a pitch of the scroll, t the width 
of the scroll. 
In this type of operation, it was recognized as result of study that 
reactive force F.sub.2 and F.sub.1, resulting at about the center of the 
wall of the main bearing 17 and at about the center of the inner surface 
of the eccentric bush 52, caused the moment of the driving shaft 14 to 
incline the shaft 14. 
This inclination cannot be avoided in the structure shown in FIG. 5, and 
the fact that the distance l.sub.1 between F.sub.1 and F.sub.2 is 
relatively large correspondingly increases the magnitude of the moment. 
The moment makes continuation of normal operation of on a compressor 
difficult owing to the resultant inequality of contact of scrolls with 
each other. 
It was also recognized that the eccentric bushing 52 spins on the inner 
surface of the hole 53 in the orbiting scroll 2, and the pin 51 orbits on 
the inner surface of the eccentric hole in the eccentric bushing 52. 
As a result, the outer surface of the bushing 52 tends to wear because of 
the rotating movement of the outer surface of the hole of the bushing 52 
rotation at the same rotation rate. 
In FIGS. 7, 8(a), 8(b), 9(b) showing a scroll compressor according to this 
invention, a first eccentric hole 16' is formed in the top of the driving 
shaft 14. The eccentricity of the hole 16' is defined by the displacement 
of its center O.sub.5 from the center O.sub.1 of the driving shaft 14. An 
eccentric bushing 26 has a shape such as a hollow cylinder and is placed 
within the eccentric hole 16'. Its eccentricity is defined by a 
displacement e between the center O.sub.5 of its outer circumference and 
the center O.sub.4 of the inner cylindrical aperture or the second 
eccentric hole 16". 
The hole 16" is formed as the inner wall of the eccentric bushing 26 which 
is entirely made of a suitable bearing material, such as a bearing metal. 
An orbiting scroll pin 4 is inserted into the eccentric hole 16" of the 
bushing 26 so that its center also lies at O.sub.4. The eccentricity R of 
the orbiting scroll pin 4 is the distance between the center O.sub.1 of 
the driving shaft 14 and the center O.sub.4 of the orbit scroll pin 4. 
In the structure shown in FIG. 7, it is easy to restrain the moment caused 
by F.sub.1 and F.sub.2, even to be able to make the moment zero by putting 
lines of action that cancel both F.sub.1 and F.sub.2, since the main 
bushing 17 is mounted within the region radially enclosing the bushing 26. 
The structure with regard to the eccentric bushing 26 in FIG. 7 is 
excellent in wear-resistance in comparison with that in FIG. 5, because 
the pin 4 mounted in the hole 16" in the bushing 26 rotates in the hole 
16" of which the speed of the surface is lower than that of the outer 
surface of the bushing 53 when the bushing 53 rotates at the same rotation 
rate. 
Since the other structure and operation of the eccentric bushing 26 of the 
scroll compressor shown in FIG. 7 are the same as or similar to that of 
the scroll compressor shown in FIGS. 2 to 5, further description about 
them will be omitted. 
FIGS. 8(a) and (b) and 9(a) and (b) illustrate the simplification of the 
gaps existing between the eccentric hole and the eccentric bushing 26, or 
between the hole 16" and the orbiting scroll pin 4. 
In this embodiment of the invention for a scroll type compressor, since the 
eccentric bushing 26 can freely rotate around its center O.sub.5, the 
center O.sub.4 of the hole 16" of the bushing 26 will also rotate around 
this center O.sub.5 when any rotational force operates on the eccentric 
bushing 26. Thus the eccentricity R is changed as the result of the 
rotation of the eccentric bushing 26, as shown in FIGS. 9(a) and (b). 
FIG. 9(a) represents a part of the scroll compressor in which a wrap 101 of 
the stationary scroll 1 is located after assembling further to the left 
than a line S which designates the proper designed position of the wrap 
101. This deviation could result from inaccurate machining or assembly of 
the apparatus. The same condition, of course, occurs if a wrap 201 of the 
orbiting scroll 2 is located too far to the right. Notwithstanding the 
incorrect alignment of the scrolls 1 and 2 during fabrication, the scrolls 
are brought into contact by the action of the eccentric bushing 26. 
If F denotes the resultant force of the centrifugal force F.sub.c and load 
F.sub.g resultant from compressing the gas, the eccentric bushing 26 is 
substantially torqued by the resultant total force F and is made to rotate 
around its center O.sub.5 by the component F' of the force F at a right 
angle to the line O.sub.4 -O.sub.5. As a result, the eccentricity R tends 
to increase but is bounded by the contact of the wraps of the scrolls 1 
and 2. The wrap 201 of the orbiting scroll 2 contacts the wrap 101 of the 
stationary scroll 1 to counterbalance the torquing force F'. As shown in 
FIG. 9(a), the two wraps 101 and 201 are thus maintained in contact. 
FIG. 9(b) represents the opposite condition that the wrap 101 of the 
stationary scroll 1 is located to the right of its proper design position. 
There still exists a force component F' that tends to rotate the eccentric 
bush 26 around its center O.sub.5 of shaft but it is considerably less 
than in the prior condition. The wrap 201 of the orbiting scroll 2 
contacts and presses against the wrap 101 of the stationary scroll 1. 
The above description demonstrates that in this invention, the wrap 201 of 
the orbiting scroll 2 is always pressed on the wrap 101 of the stationary 
scroll 2 to sufficiently seal it in a radial direction even though the 
scrolls may be incorrectly aligned due to inaccurate machining or 
assembly. 
The pumping capacity of the scroll compressor based on this invention is 
increased because of the decrease of gas leakage from the compression 
pockets 3 and 5. Also, the pumping efficiency increases because of a 
reduction of load associated with recompressing whatever gas has leaked 
out. 
While the value of eccentricity or orbiting radius R cannot be increased 
without restriction, the permissible orbiting radius R has sufficient 
range to cover the errors occurring in fabrication or assembly.