KR20020034883A - Plural cylinder rotary compressor - Google Patents

Abstract

PURPOSE: A double cylinder rotary type compressor is provided to reduce bending and deformation in an intermediate shaft part between a first crank pin and a second crank pin of a crankshaft, to reduce friction loss of a bearing and the crankshaft and friction loss of a rolling piston and a sealed plate, and to prevent a leakage of compressed gas. CONSTITUTION: A cross section of an intermediate shaft of a double cylinder rotary type compressor is larger than a part where cross sections of a first crank pin(13) and a second crank pin(14) are overlapped, and a difference in steps is provided between a first crank pin side and a second crank pin side of the intermediate shaft. Consequently, friction loss and loss of machine efficiency due to partial contact are reduced. Since an extra clearance due to partial contact is reduced, the leakage of compressed gas is prevented and the reduction of volumetric efficiency is suppressed. The reduction of a performance due to the deformation of a crankshaft(3) is suppressed.

Description

Multi Cylinder Rotary Compressor {PLURAL CYLINDER ROTARY COMPRESSOR}

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a multi-cylinder rotary compressor for use in a refrigeration cycle such as a refrigerator or an air conditioner. The present invention relates to a multi-cylinder rotary compressor which reduces bending deformation of a crankshaft and improves efficiency.

As a compressor for an air conditioner or a refrigerator, reciprocating type, rotary type, scroll type, screw type, etc. are used. HCFC type refrigerant | coolant was conventionally used for these compressors. In recent years, however, HCFC refrigerants are also decomposed by ultraviolet rays in sunlight, and it is clear that the generated chlorine destroys the ozone layer in the stratosphere, and in 1987, the Montreal Protocol on Substances That Destroy the Ozone Layer, Since 2004, HCFC refrigerants have been regulated in stages since 2004. For these reasons, each company is participating in the development of a compressor using an HFC-based refrigerant or a natural refrigerant that does not contain a prone in its constituent molecules as an alternative to the HCFC-based refrigerant.

When a natural refrigerant such as HFC refrigerant or carbon dioxide is used for the compressor, the pressure of the working gas may increase, so that leakage of the refrigerant gas from the gap between the vibrating portion of the compression chamber is increased than that of the conventional HCFC system. Therefore, the rotary compressor which can set the leakage from a clearance small may be used for a replacement refrigerant.

Another problem caused by the higher pressure of the working gas than the related art is that the refrigerant gas remaining in the dead volume present in the discharge port of the compression chamber which cannot be stopped mechanically is expanded without being discharged, thereby reducing the volume effect. . In this case, in a compressor having an output of 2 to 3 horsepower or more, the above-described problems of leakage and dead volume become remarkable. Therefore, the rotary type which can set the dead volume of a compression chamber small and the leakage from a clearance is small in many cases is used for the compressor for replacement refrigerants.

By the way, when the replacement refrigerant is used in the one-cylinder rotary compressor, since the difference between the suction pressure and the discharged pressure of the working gas is large, the torque fluctuation range due to compression increases, and the vibration rotation of the eccentric portion of the crank also increases. In particular, in the case of a compressor of about 2 to 3 horsepower, there was a problem that the vibration increased. In order to solve the problem, a multiple cylinder rotary compressor may be employed.

In general, when designing a rotary compressor (for example, using a conventional R 22 refrigerant), mechanical loss of rotational power due to friction between the crankshaft (rotation shaft) and the bearing occurring in the bearing portion, i.e. In order to reduce frictional losses, it is preferable to provide the crankshaft diameter as small as possible in order to reduce the contact area between the crankshaft and the bearing and to reduce the moment of inertia. In addition, by reducing the diameter of the crankshaft, the weight of the piston rotating member configured at the same time as the crankshaft becomes light, and the motor power consumption for driving can be reduced. In addition, there is an advantage that the portion where the crankshaft diameter becomes small and the outer diameter of the entire compressor can also be reduced, which is effective for saving space.

On the other hand, when the compressor has a high condensation pressure as an alternative refrigerant and uses an HFC refrigerant, the latent heat of evaporation is larger and the evaporation gas density is larger than that of the conventional HCFC refrigerant, so that the capacity per exclusion volume (unit exclusion volume) of the compressor is increased. For example, by changing the refrigerant from R 22 to R 410 A, the compression pressure is about 1.5 times.

As a two-cylinder rotary compressor that proposes a dimension of a compression chamber applied to such an alternative refrigerant, Japanese Patent Laid-Open Publication No. Hei 8-144976 excludes the cylinder height as the crank eccentricity in the range of 0.07 to 0.13. This is described.

As described above, there are several points of concern in the case of using an alternative refrigerant such as an HFC refrigerant or a natural refrigerant in a multi-cylinder rotary compressor manufactured using a conventional component.

I) There is a refrigerant in which the pressure difference between the suction pressure and the discharge pressure of the working gas increases.

Ii) By a plurality of cylinders, the distance between the bearings provided on both sides of the cylinder becomes larger than in the case of one cylinder. That is, the support point which receives gas load becomes long.

I) It is aimed to provide a small crankshaft diameter.

Examining these points, the bending deformation of the crankshaft between the bearings arranged on both sides of the plurality of cylinders tends to be large. As the bending deformation of the crankshaft becomes large, the inclination of the crankshaft with respect to the bearing becomes large, causing bias. As a result of this bias, the crankshaft receives a pressing reaction from the bearing. Based on the reaction force, energy loss due to friction becomes loss of driving force.

In addition, the piston tilts because bending deformation occurs between the two bearings on the crankshaft. Therefore, a bias arises between the end surface (cylinder opposing surface part of the sealing member) of the member which seals each cylinder, and the end surface (cylinder opposing surface) of the partition plate, and each piston end surface. Each piston end face receives a reaction force from an opposing end face, and friction loss occurs, and the loss of the driving force is further increased. In addition, one-sided contact may occur between the outer surface of the piston and the inner surface of the cylinder, and a frictional loss caused by the piston's reaction force from the inner surface of the cylinder may also occur.

As described above, the energy loss (friction loss) due to friction cannot be ignored for the performance of the compressor when the amount of bending deformation of the crankshaft becomes large. All of these friction losses are caused by bias. Although the crankshaft is bent, reaction force is generated in the crankshaft by being constrained along the upper and lower end surfaces (cylinder opposing surface portions) of the sealing member constituting the two bearings, the partition plate, and the cylinder. The contact area at each contact is small and the contact surface pressure is high, resulting in mechanical energy loss of the compressor due to the rotation of the crankshaft.

On the other hand, since the frictional loss that occurs in the bearing portion moves on the crankshaft and the bearing on average, all of them are smaller than the frictional loss caused by the above-described bending deformation. Therefore, reducing the crankshaft diameter reduces the frictional loss between the crankshaft and the bearing, and has an effect on performance. However, the larger the bending deformation, the greater the influence on the compressor performance of the frictional loss, which is increased, and therefore, the reduction in the crankshaft diameter is limited.

In addition, in consideration of the occurrence of the inclination of the piston generated by the bending of the crankshaft, it is necessary to set a large gap between the piston end face, the partition plate and the end plate (sealing member) in accordance with the inclination of the piston. In addition, consideration should be given to the clearance between the piston and the vane that is in contact with the piston and divides the compression chamber and the suction chamber. As these gaps increase, the amount of leakage of compressed gas increases, and the volumetric efficiency of the compression chamber decreases.

On the other hand, even in the case of using HFC-based mixed refrigerants or carbon-based natural refrigerants having low operating pressures as alternative refrigerants, when the distance between the bearings increases, the above-described crankshaft bends, and accordingly, the piston is inclined so that the bearing or end Skew on the plate. Since the replacement refrigerant does not have chlorine molecules for suppressing extreme pressure as in the conventional HCFC refrigerants, friction loss is increased at the contact portion, and energy loss and wear are increased.

Solving the above problems is a problem in designing a multi-cylinder rotary compressor using an alternative refrigerant.

The technique described in JP-A-8-144976 discloses design criteria for optimizing the removal efficiency of refrigerant gas. However, it does not describe the friction loss and the decrease of the volume efficiency due to the bending deformation of the crankshaft. Solving them is an important task.

An object of the present invention is to provide a multi-cylinder rotary compressor which reduces bending deformation of a crankshaft and has a good compression efficiency with a long life.

Further, another object of the present invention is to provide a multi-cylinder rotary compressor which reduces bending deformation of the crankshaft while maintaining assembly.

In order to achieve the above object, the multi-cylinder rotary compressor of the present invention includes a motor unit and a compressor unit in a sealed container, the motor unit and the compressor unit are connected along the crankshaft, and the crankshaft has a first crank pin eccentric with respect to the rotating shaft. A partition plate having a second crank pin, wherein the compressor section is provided between the main and sub bearings for supporting the crankshaft, and a through hole of an inner diameter larger than the outer diameter of the first or second crank pin; Between the first and second cylinders partitioned by the first and second cylinders eccentrically interlocked with the rotation of the crankshaft into the first and second cylinders, and between the two compression chambers formed in the first cylinder and the second cylinder. The diameter of the crankshaft was larger than the diameter of the crankshaft fitted to the rotor of the motor unit.

Further, in order to achieve the above object, the multi-cylinder rotary compressor of the present invention includes an electric motor part and a compressor part in a sealed container, the electric motor part and the compressor part are connected along the crank shaft, and the crank shaft is eccentric with respect to the rotating shaft. It has a pin and a second crank pin, has an intermediate shaft between the first crank pin and the second crank pin, the compressor unit is provided between the main and secondary bearings supporting the crankshaft, and the main and secondary bearings, Or an eccentricity of each crankpin, as an intermediate axis of the crankshaft in the first and second cylinders and the first and second cylinders partitioned by a partition plate having a through hole of an inner diameter larger than the outer diameter of the second crankpin. The extension part extended in the direction was provided. Each extension may be integral with each crank pin.

It is preferable that the rotational trajectory of the outermost diameter of the extension part provided in the middle part of a crankshaft is smaller than the internal diameter of the through hole of a partition plate.

Further, in order to achieve the above object, the multi-cylinder rotary compressor of the present invention includes a motor unit and a compressor unit in a sealed container, the motor unit and the compressor unit are connected along the crank shaft, and the crank shaft is eccentric with respect to the rotating shaft. And a second crank pin, and having an intermediate shaft between the first crank pin and the second crank pin, wherein the compressor unit is provided between the main and sub bearings supporting the crank shaft, and the main and sub bearings. First and second eccentrically interlocked with the rotation of the crankshaft in the first and second cylinders and the first and second cylinders partitioned by a partition plate having a through hole of an inner diameter larger than the outer diameter of the second crank. A piston having a piston, the radial end surface of the intermediate shaft of the crankshaft being larger than the overlapping portion of the radial end surface of the first crankpin and the second crankpin, and the first crankpin side It was set as the structure which has a level | step difference between it and the 2nd crank pin side.

That is, the area of the intermediate shaft was formed so as to be different in the axial direction in two stages, and the area was enlarged in the eccentric direction of the cylinder provided with each of the intermediate axes eccentric.

Accordingly, since the intermediate shaft diameter between the crank pin and the crank pin can be set to the maximum possible value, the bending deformation of the intermediate shaft can be reduced, and the upper end of the crank shaft and the piston in the bearing or the first cylinder and the upper sealing plate, Since the bias between the piston lower end portion and the lower sealing plate in the second cylinder, the piston outer surface and the cylinder inner surface is small, the friction loss is reduced, and the loss of mechanical efficiency is reduced. In addition, the extra clearance between the crankshaft, the bearing piston and the vane, and the piston and the sealing plate also becomes small, so that leakage is reduced and the reduction in volume efficiency is suppressed. Therefore, the fall of the performance by the deformation of a crankshaft is suppressed.

In order to achieve another object, the multi-cylinder rotary compressor of the present invention includes a motor unit and a compressor unit in a sealed container, the motor unit and the compressor unit are connected along the crank shaft, and the crank shaft is eccentric with respect to the rotating shaft. And a second crank pin, wherein the compressor unit is provided between the main and sub bearings for supporting the crankshaft, and a partition plate having a through hole of an inner diameter larger than the outer diameter of the first or second crank. And first and second pistons eccentrically interlocked in accordance with rotation of the crankshaft in the first and second cylinders partitioned by the first and second cylinders, the crankshaft being disposed between the first and second crankpins. The extension part extended in the eccentric direction of each crank pin is provided in the intermediate part to connect, and the thickness of a partition plate is with this crank pin of the extension part formed in the one crank pin side. It was smaller than the distance to the face opposing the partition plate of the crank pin of the facing, and the other reference line.

BRIEF DESCRIPTION OF THE DRAWINGS Fig. 1 is a longitudinal sectional view showing a compressor portion and an electric motor portion of a two cylinder rotary compressor showing one embodiment of the present invention.

FIG. 2 is a view excluding a part of the crankshaft 3 of the two cylinder rotary compressor in FIG. 1; FIG.

FIG. 3 is a cross-sectional view of the connecting portion 31 of the intermediate shaft 30 of the crankshaft 3 of FIG. 2 projected in a plane perpendicular to the rotational axis.

4 is a cross-sectional view of a part of the crankshaft 3 of the two-cylinder rotary compressor showing a modification of the example shown in FIG.

FIG. 5 is a cross-sectional view of the connecting portion 310 of the intermediate shaft 300 of the crankshaft 3 of FIG. 4 projected in a plane perpendicular to the rotational axis. FIG.

FIG. 6 is a view excluding a part of the crankshaft 3 of the two cylinder rotary compressor showing a modification of the example shown in FIG.

FIG. 7 is a cross-sectional view of the connecting portion of the first connecting portion 311 and the second connecting portion 312 of the intermediate shaft 300 of the crankshaft 3 shown in FIG. 6 in a plane perpendicular to the rotation axis.

Fig. 8 is a view showing the rotor of the compressor section and the motor section of a two cylinder rotary compressor showing one embodiment of the present invention.

Figure 9 is a view of a part of the crankshaft of the two cylinder rotary compressor showing one embodiment of the present invention.

Fig. 10 is a top view of the crankshaft showing one embodiment of the present invention.

FIG. 11 is a view excluding a portion of the crankshaft of a two cylinder rotary compressor showing another embodiment of the present invention. FIG.

12 is a view of a part of the crankshaft of a two cylinder rotary compressor showing still another embodiment of the present invention;

Fig. 13 is a top view of a crankshaft showing still another embodiment of the present invention.

Figure 14 shows a part of a crankshaft showing yet another embodiment of the present invention.

Fig. 15 shows a part of a compressor showing yet another embodiment.

FIG. 16 shows a part of a compressor showing still another embodiment; FIG.

FIG. 17 shows a portion of a compressor showing still another embodiment; FIG.

18 shows a part of a compressor showing another embodiment.

19 is a diagram showing an example of the results of computer simulation.

Fig. 20 is a diagram showing the relationship between the displacement B of the main bearing upper end portion of the crankshaft and the minimum value C of the set gap in the state where the crankshaft has caused bending deformation.

Fig. 21 summarizes the results of the computer simulation.

22 is a view showing an assembly process of a compressor to which an embodiment of the present invention is applied.

<Explanation of symbols for the main parts of the drawings>

1: sealed container

2: motor part

3: crankshaft

4: main bearing

5: compression mechanism

6: cylinder

8: negative bearing

10: piston

11: cover

13, 14: crank pin

30: middle axis

31, 32: connection part

40, 41: balance weight

50: partition plate

51: through hole

EMBODIMENT OF THE INVENTION Hereinafter, embodiment of this invention is described by drawing. 1 is a longitudinal sectional view showing one embodiment of a multi-cylinder rotary compressor according to the present invention, FIG. 2 is a side view showing the shape of the crankshaft of the multi-cylinder rotary compressor, and FIG. 3 shows the cross-sectional shape of the crankshaft of FIG. It is a cross section.

1 to 3 show a two-cylinder rotary compressor and its components having two cylinders with an electric motor part 2 and a compressor part 5 inside the sealed container 1.

The main bearing 4 which supports the crankshaft 3 is fixed to the inner wall of the airtight container 1 by welding etc. The electric motor 2 is accommodated in one space of the main bearing 4, and the compressor part 5 is accommodated in the other space. The electric motor 2 has the rotor 2a with which the crankshaft 3 is fitted, and has the coaxial stator 2b opposite to this. The stator 2b is fixed to the airtight container 1.

The crankshaft 3 is supported by the subordinate bearing 8 at the front-end | tip part through the main bearing 4. Between these main bearing 4 and the sub bearing 8, it has two cylinders 6a and 6b and the partition plate 50. As shown in FIG. Two sets of compression chambers are formed by partitioning two cylinders 6a and 6b by partition plate 50. That is, the partition plate 5 is in the position which fits between two cylinders 6a and 6b.

The crankshaft 3 has the crank pin parts 13 and 14 in the position in these cylinders 6a and 6b. Moreover, the piston 10a, 10b is accommodated in these cylinder 6a, 6b, respectively. These pistons 10a and 10b are fitted into the crank pin portions 13 and 14, respectively. When the crankshaft 3 is driven to rotate by the electric motor part 2, piston 10a, 10b rotates by phase difference of 108 degrees mutually according to the eccentric rotation of the crank pin parts 13 and 14, respectively.

The vanes 12a and 12b are always pressed against these pistons 10a and 10b by a screw member. In the cylinder 6a, the compression chamber is formed by the piston 10a and the vane 12a, and in the cylinder 6b by the piston 10b and the vane 12b. By the eccentric rotation of the pistons 10a and 10b based on the rotation of the crankshaft 3, the compression chamber in the cylinders 6a and 6b repeats reduction and expansion. When the compression chambers of the cylinders 6a and 6b are enlarged, the refrigerant gas supplied from the suction pipes 19a and 19b is sucked into these compression chambers.

As the crank shaft 3 rotates and the compression chamber shrinks, the refrigerant gas is compressed. When the pressure of the refrigerant gas reaches a certain size (discharge pressure), the compressed refrigerant gas in the cylinder 6a is discharged to the discharge chamber 17 formed by the main bearing 4 and its cover 11. Similarly, the compressed refrigerant gas in the cylinder 6b is discharged to the discharge chamber 15 formed by the sub bearing 8 and its cover 9. Although not shown, the refrigerant gas returning to the compressor through the refrigeration cycle is alternately compressed by the cylinders 6a and 6b and discharged from the sealed container 1 through the discharge pipe 18 to the refrigeration cycle.

Figure 2 is a view excluding part of the crankshaft 3 of the two cylinder rotary compressor showing one embodiment of the present invention. The intermediate shaft 30 connects the first crank pin 13 and the second crank pin 14 which are alternately eccentric with respect to the rotation shaft. It can also be seen as part of the crankshaft 3. The through hole 51 is an opening provided in the partition plate 50. The 1st and 2nd connection parts 31 and 32 are a part of intermediate shaft 30, and are provided in the position accommodated in the through hole 51 of the partition plate 50, respectively. The crankshaft 30 including the first crankpin 13, the second crankpin 14, the intermediate shaft 30, the first connecting portion 32 and the second connecting portion 31 is integrally formed by casting. do.

When the refrigerant is compressed by the rotational interlocking operation of the first piston 10a and the second piston 10b, this compression load (not shown) is applied to the first crank pin 13 and the second crank pin 14. . The direction of this compressive load is the eccentric direction of each crank pin. That is, in FIG. 2, the inclination rotation moment which rotates the crankshaft 3 to the left direction (counterclockwise rotation) moves. Its size is relieved by both crank pins and the working compression pressure, and there is no change in size depending on the rotational position of the crankshaft (3). However, if it mentions in detail, the pressure of the operation chamber formed by the outer wall surface of one piston 10 and the inner wall surface of the cylinder 6 becomes the largest value when it consists of discharge pressure.

Since there is a compressive load by this crank pin, in the direction which compresses the intermediate shaft 30 on the eccentric part side of the intermediate shaft 30, and in the direction which removes the intermediate shaft 30 on the half eccentric part side of the intermediate shaft 30. A working stress is applied. For this reason, in the conventional intermediate shaft of the same diameter as the crankshaft 3, there existed a problem that an intermediate shaft part deform | transforms according to this stress, and the main bearing 4 and the sub bearing 8 are biased.

In order to solve such a problem, in the embodiment shown in FIG. 2, the connection part (intermediate centering on the eccentric direction) which extended in the eccentric direction of each crank pin to the intermediate shaft 30 which is about the same diameter as the crank shaft 3, Extension part which increased the diameter of the shaft 30] was provided. That is, the intermediate shaft 30 in contact with the first crank pin 13 has the first connecting portion 32 extending in the eccentric direction of the crank pin 13 in the intermediate shaft 30 in contact with the second crank pin 14. The 2nd connection part 31 extended in the eccentric direction of the crank pin 14 was provided.

Thus, the second connecting portion 31 provided on the intermediate shaft 30 extends in the eccentric direction of the crank pin 14. That is, the cross section perpendicular to the crankshaft 3 of the connecting portion 31 formed on the intermediate shaft 30 (for example, AA cross section) is included in the cross section of the second crank pin 14, but the first crank pin It extends to the part which is not included in the cross section of (13). Similarly, the first connecting portion 32 provided on the intermediate shaft 30 extends in the eccentric direction of the crank pin 13. That is, the cross section perpendicular to the crankshaft 3 of the connecting portion 32 formed on the intermediate shaft 30 is included in the cross section of the first crank pin 13, but included in the cross section of the second crank pin 14. It is extended to the part that is not.

Thus, by providing the connection parts 31 and 32 extended from the diameter of the crankshaft 3 to the intermediate shaft 30, it becomes possible to support the load which the crank pin gives to the intermediate shaft 30 in FIG. The deformation of the intermediate shaft 30 is alleviated, and the bias of each bearing is suppressed.

Next, the assembly method is briefly described. In FIG. 1, the electric motor part 2 and the main bearing 4 are conventionally assembled in the airtight container 1. The compression mechanism part 5 is attached to the crankshaft 30 including the first crank pin 13, the second crank pin 14, the intermediate shaft 30, the first connecting portion 332, and the second connecting portion 31. Assembly of each component to be formed is demonstrated.

The 1st piston 10a is assembled to the 1st crank pin 13, and the 1st cylinder 6a which covers the periphery is arrange | positioned. Next, the partition plate 50 is arranged. The through hole 51 of the partition plate 50 is inserted into the crankshaft 3 and passes through the second crank pin 14. The inner diameter of the through hole 51 is slightly larger than the outer diameter of the second crank pin 14. And when the partition plate 50 passes the 2nd crank pin 14, the partition plate 50 is moved to the half eccentric direction of the 2nd crank pin 14. As shown in FIG. That is, the partition plate 50 is moved in the direction in which the center of the through hole 51 of the partition plate 50 coincides with the center of the crankshaft 3. Subsequently, the assembly of the compression mechanism part 5 is completed by inserting the 2nd piston 10b into the 2nd crank pin 14, arrange | positioning a 2nd cylinder around it, attaching the sub bearing 8, and screwing it. .

Points to keep in mind regarding this assembly are shown below. First, the diameter of the through hole 51 of the partition plate 50 can pass through the 2nd crank pin 14 in the 1st. Secondly, each connecting portion provided on the intermediate shaft 30 does not come into contact with the inner wall of the through hole 51 even when the crank shaft 3 is rotated. Since the intermediate shaft 30 is disposed within the inner diameter of the through hole 51, the load is caused when the connecting portions provided in the intermediate shaft 30 and the inner wall of the through hole 51 interfere with each other. Third, the relationship between the axial dimension on the intermediate shaft 30 of the first connecting portion 32 and the second connecting portion 31 and the thickness of the partition plate 50 is attached to the predetermined position. Possible relationship. That is, after the partition plate 50 passes through the second crank pin 14 through which the partition plate 50 passes too much at the time of assembly (near the intermediate shaft 30), the center of the through hole 51 is the crankshaft. The partition plate 50 is movable so that it may coincide with the center of (3).

The first matter is solved by making the inner diameter of the through hole 51 larger than the outer diameter of the second crank pin 14. In the multi-cylinder rotary compressor, the piston end face (axial donut shaped plane) and the partition plate end face consisted of a sealing element for partitioning the compression chamber. The larger the distance between the trajectory of the outer diameter of the eccentrically interlocking piston and the through hole inner diameter of the partition plate is, the higher the sealing property is. Therefore, the smaller the through hole inner diameter is preferable. On the other hand, in order to assemble the partition plate between the crank pins, the inner diameter of the through hole of the partition plate must be larger than the outer diameter of at least one crank pin, and the difference should preferably be the minimum that the crank pin can pass through. .

The second matter will be described with reference to FIG. 3. This figure is AA sectional drawing which shows the cross-sectional shape of the connection part 31 of the crankshaft 3 of FIG. In FIG. 3, RP 2 is an outer radius centering on the center axis of the 2nd crank pin 14 (solid line), and RH is an inner radius of the through-hole 51 (dotted line) of the partition plate 50. In FIG. RJ is RH in the outer radius of a circle (circle with two dashed-dotted lines) showing the trajectory of the second connecting portion 31 furthest from the center of the crankshaft 3 when the crankshaft 3 is rotated. Is less than In addition, the circle | round | yen with two 2-point chain part is a crank pin 13.

That is, if the rotational trajectory of the part which is separated from the center of the most intermediate shaft 30 of two connection parts is inside than the inside diameter of the through-hole 51, they will not connect. In this embodiment, as shown in FIG. 3, since the rotation trajectory of the outermost diameter of a connection part is set smaller than the internal diameter of the through-hole 51, the crankshaft 3 can be rotated without a trouble.

The third item will be described. In this case, the thickness of the first connecting portion 32 and the thickness of the second connecting portion 31 (the size in the axial direction of the intermediate axis) are the same, and the thickness of each connecting portion is half the height of the intermediate axis, and the partition plate 50 Is assumed to be thicker than the thickness of each of the first connecting portion 32 and the second connecting portion 31. Then, the partition plate 50 which passed the crank pin 14 at the time of assembly contacts the 2nd connection part 32. As shown in FIG. At this time, since the through hole 51 did not miss the second crank pin 14, the horizontal movement is restricted by the second crank pin 14. For this reason, the through-hole 51 cannot be arrange | positioned in the position of the intermediate shaft 30 in the said assumed condition.

In this embodiment, the shape of the connection part provided between two crank pins is prescribed | regulated as follows from the relationship between the thickness of the partition plate 50, and the thickness of a connection part. The surface opposite to the 1st crank pin 13 of the 1st connection part 32 by the side of the 1st crank pin 13 which does not let the through-hole 51 of the partition plate 50 at the time of assembly | assembly [in this example, Surface on the second crank pin 14 side of the first connecting portion 32 and the surface (reference line) facing the partition plate 50 of the second crank pin 14 passing through the through hole 51. The distance was made larger than the thickness of the partition plate 50. The thickness of the 2nd connection part 31 on the side of the 2nd crank pin 14 which let the through-hole 51 pass in the range which this condition is satisfied can be set freely. This configuration enables the partition plate 50 to move to a predetermined position after the through hole 51 passes through the second crank pin 14.

According to this embodiment, the cross-sectional shape of the 2nd connection part 31 exists in the part which overlaps with the 2nd crank pin 14 and the through-hole 51 of the partition plate 50 in a fixed position, and The part becomes a cross-sectional shape provided with the part which does not overlap with the cross section of the 1st crank pin 13. As shown in FIG. Therefore, the through hole 51 can pass through the second crank pin 14 and the second connecting portion 31 during assembly, and can rotate without being connected during operation.

In this embodiment, the cross-sectional shape of the second connecting portion 31 that occupies most of the intermediate shaft 30 overlaps with the cross section of the first crank pin 13 in which the through hole 51 does not pass during assembly in the axial direction. Since it expands to the part which is not, the cross-sectional area of the intermediate shaft 30 increases, and the bending deformation of the intermediate shaft 30 becomes small.

4 is a diagram showing another modified example. FIG. 5 is a BB cross-sectional view showing the cross-sectional shape of the first connecting portion 310 of the intermediate shaft 300 in the state accommodated in the through hole 51 of the partition plate 50 in FIG. In FIG. 5, the first crank pin 13 whose shape in the vertical section with respect to the rotation axis of the first connecting portion 310 provided on the intermediate shaft 300 of the crank shaft 3 is closer to the motor 2 of the two crank pins. It is expanded so that there may exist a part included in the cross section of (), and not included in the cross section of the 2nd crank pin 14. As shown in FIG. In Fig. 5, RP 1 is the outer radius around the central axis of the first crank pin 13 (solid line), and RH is the inner radius of the through hole 51 (dotted line) of the partition plate 50. RJ is less than RH in the outer radius of the first connection 310.

As in the first embodiment, by increasing the cross-sectional area of the first connecting portion 310 to the maximum within the range in which the partition plate 50 can be assembled, the rigidity of the intermediate shaft 300 can be increased and the bending deformation of the shaft is small. You can do

In addition, the first connecting portion 310 has an eccentric unbalance weight with respect to the fixed shaft. In a structure in which the unbalanced weights of two crank pins and pistons are balanced by balance weights 40 and 41 provided above and below the rotor 2a of the motor 2, an eccentric weight is provided near the motor 2. The balance weight 40 provided in the lower end part of the rotor 2a can be made small. As a result, it is possible to reduce the balance weight of the material, reduce the required space, or reduce the axial bending deformation due to the centrifugal force of the balance weight, thereby reducing the bearing friction loss and reducing the vibration.

In this case, in assembly, the first crank pin 13 is a crank pin through which the through hole 51 passes.

FIG. 6 shows another modified example, and FIG. 7 shows a cross-sectional shape of the first connecting portion 311 connected to the first crank pin 13 in the intermediate shaft 300 of FIG. 6 and the second crank pin ( It is DD sectional drawing which shows the cross-sectional shape of the 2nd connection part 312 connected to 14).

The minimum distance LP 1 between the first crank pin 13 and the second connecting portion 312 is larger than the thickness BS of the partition plate 50. The vertical cross section with respect to the rotation axis of the 1st, 2nd connection part 311, 312 is contained in the vertical cross section with respect to the rotation axis of the crank pin connected, respectively, and is not included in the vertical cross section with respect to the rotation axis of the other crank pin. It is extended to the part.

Moreover, when the cross-sectional shape of the said two connection parts is made into the circle of concentric same outer diameter, the cross-sectional shape of both connection parts containing this part can be formed to the maximum and easily processable.

According to this embodiment, the thickness of the 2nd connection part 312 of the side through which the through hole 51 does not pass can be made larger than the thing shown in FIG. The cross-sectional shape of the connecting portions 311 and 312 of the intermediate shaft 300 is a range in which one crank pin and the through hole 51 of the partition plate 50 overlap, and overlap with the cross section of the other crank pin. It is an unused part. Therefore, the through hole 51 passes through the crank pin to which the other connection part and the other connection part are connected with respect to one connection part whose distance from a crank pin is larger than the thickness of the partition plate 50 at the time of assembly. In addition, both connecting portions can rotate without being connected to the through hole 51 during operation. Therefore, since the cross-sectional shape of a connection part expands in the range which can be assembled of the partition plate 50, the cross-sectional area of this part increases and it is a structure which made the bending deformation of the intermediate shaft 300 small.

Figure 8 shows another embodiment of a two cylinder rotary compressor 1 of the present invention. The same code | symbol as FIG. 1 is attached | subjected about the main structure.

In FIG. 8, the description about the compression chamber was described as follows. The compression mechanism part 5 consists of the 1st compression chamber 31 and the 2nd compression chamber 41. As shown in FIG. The 1st compression chamber 31 is the upper sealing board 20 integrated into the main bearing 4 (radial bearing) which supports the 1st cylinder 6a and the crankshaft 3, and the 1st compression chamber It is comprised by the partition plate 50 which partitions the 31 and the 2nd compression chamber 41. As shown in FIG. The 2nd compression chamber 41 is comprised from the lower sealing board 60 integrally processed by the sub bearing 8 which supports the 2nd cylinder 6b and the crankshaft 3, and the partition plate 50. have.

In addition, although it is not described in detail in this figure in FIG. 8, one part of the upper sealing board 20 consists of a structure which contacts a part of crankshaft in a drust direction. Moreover, it consists of the structure which supports the load of a thrust direction in the part 100 of the lower sealing board 60 crankshaft.

9 is a view excluding a part of the crankshaft 3 of the two cylinder rotary compressor 1 showing another embodiment of the present invention. FIG. 10 is a top view seen from the A side in the axial direction of FIG.

The inside of the crankshaft is provided with the hollow part 33 as shown to FIG. 10 (a) and FIG. 10 (b). The difference between FIG. 10 (a) and FIG. 10 (b) is that the intermediate shafts 112a and 112b have a substantially elliptical shape in FIG. 10 (a), and FIG. It is.

As shown in Fig. 10A, the area (cross section) in the direction perpendicular to the axial direction of the intermediate shaft 112a on the side of the first crank pin 13 is enlarged in the eccentric direction of the first crank pin 13; I have a structure. In other words, the center P _u1 of the intermediate shaft 112a is shifted from the rotation center O ₁ of the crank shaft 3 in the longitudinal direction of the first crank pin.

Moreover, the intermediate shaft 112b by the side of the 2nd crank pin 14 is the structure extended in the eccentric direction of the 2nd crank pin 14. As shown in FIG. In other words, the center P _s1 of the intermediate shaft 112b is shifted in the longitudinal direction of the second crankshaft from the rotation center O ₁ of the crankshaft 3. Therefore, there is a stepped structure between the intermediate shaft 112a and the intermediate shaft 112b.

In Fig. 10B, the cross-sectional shapes of the intermediate shafts 112a and 112b are circular. Therefore, it is possible to reduce (a) the amount of deviation 10 from the center P _u2, P _s2 of the center of rotation O ₁ of the crank shaft (3). Further, by making the shape circular, processing is slightly easier than the mold release of Fig. 10A. Also in any of Figs. 10A and 10B, since the diameter of the intermediate shaft can be increased, the bending deformation of the intermediate shaft becomes small.

Figure 11 is a view excluding a part of the crankshaft of another embodiment. 9, the intermediate shafts 112a and 112b have substantially the same axial length in FIG. 9, but in this embodiment, the stepped portion is shifted to the first crank pin 13 side. In other words, the height h1 of the intermediate shaft 112a is formed to be smaller than the height h2 of the intermediate shaft 112b.

In the two-cylinder rotary compressor of the present embodiment, as shown in Fig. 8, two thrust bearings are provided. The thrust bearing is an upper side of the first crank pin 13 and a lower side of the second crank pin 14. In the installation of the compressor of the present embodiment, the electric motor unit 2 is installed on the upper side, that is, the first crank pin 13 is installed on the upper side. Therefore, the thrust load is larger in the lower thrust bearing. Therefore, when the crank pin is tilted due to the deformation of the intermediate shaft, the unbalanced contact with the thrust bearing is larger at the lower side. Therefore, as shown in Fig. 11, by lowering the intermediate shaft 112b on the lower side, the fall deformation of the lower crank pin, that is, the second crank pin, can be reduced, thereby reducing the unbalanced contact with the thrust bearing. .

Next, the assembling process of the compressor of the present embodiment will be described with reference to FIG. In Fig. 22, the upper sealing plate is assembled downward, but this may be the opposite and the crankshaft may be in the transverse direction.

Fig. 22 (1): The first cylinder 6a and the piston 10a are temporarily assembled by bolts from the lower side of the upper sealing plate 20 as an integral part of the upper sealing plate 20 and the main bearing 4. Assemble

Fig. 22 (2): The crankshaft 3 is inserted so that the hole of the piston 10a and the hole of the main bearing 4 coincide with each other. Next, the crankshaft 3 is set. To set the crankshaft 3 is to position the crankshaft, determine the position while measuring the gap between the piston 10a and the first cylinder 6a with a gap gauge, and from the lower side of the upper closing plate 20, Fasten with bolts.

22 (3): The partition plate 50 is inserted. The partition plate 50 is inserted until the step of the intermediate shafts 112a and 112b is attached to the intermediate shaft 112a.

Fig. 22 (4): The partition plate 50 is shifted horizontally to the place where the intermediate shaft 112a enters.

Fig. 22 (5): A partition plate 50 is set.

Fig. 22 (6): The second cylinder 6b is inserted and set with a buried bolt not shown.

Fig. 22 (7): Insert and set the lower sealing plate 60 serving as the secondary bearing.

The structural relationship between the intermediate shaft 112 and the partition plate 50 in the case of assembling is considered in order to change the partition plate 50 in the process shown in Fig. 22 (4). It is necessary to make it thinner than 112b. When the partition plate 50 is thicker than the intermediate shaft 112b, the partition plate 50 abuts against the first crank pin 13 so that the partition plate 50 cannot be shifted laterally.

In the structure of the crankshaft 3 in FIG. 11, since the intermediate shaft 112b is sufficiently thick, it becomes easy to assemble.

12 is a crankshaft 3 in still another embodiment. FIG. 13 is a top view of the crankshaft 3 in FIG. 12 as seen from the upper direction B. FIG. 9 differs from that in FIG. 9 in which the outer periphery of the intermediate shaft coincides with the outer periphery of the crank pin closer to the intermediate axis. In contrast, in the present embodiment, the intermediate axis 112a (shown in dashed lines) on the side of the first crank pin 13 is configured to be concentric with the centers X ₁ and X ₂ of the first crank pin 13. In addition, the intermediate shaft 112b on the side of the second crank pin 14 was configured to be concentric with the centers Y ₁ and Y ₂ of the second crank pin 14. Each intermediate shaft is configured to be inside the outer periphery of each crank pin. As a result, the diameter of the intermediate shaft can be increased to reduce bending deformation, and the machining can be easily performed for concentricity with the crank pin.

FIG. 14 is a configuration in which angles are removed by smoothing the stepped portions between the intermediate shafts 112a and 112b of FIGS. Thereby, the stress concentration of each part can be reduced.

Fig. 15 is a sectional view showing a part of a compressor showing still another embodiment. The difference between the configuration of this embodiment and FIG. 9 is a configuration in which a portion coincident with the outer circumference at the predetermined height in the middle of the two intermediate shafts is widened in both deviation directions of the respective intermediate shafts. In other words, it is a structure changed on the long side and short side of the crank pin of the step part eccentric direction. The stepped portion 113a of the intermediate shaft 113 is raised to the limit of the second cylinder, and the stepped 113b is raised to the limit of the first cylinder. By this structure, the intermediate shaft 113 can utilize the space formed by the 1st crank pin, the 2nd crank pin, and a partition plate to the maximum, and can also reduce the bending deformation of an intermediate shaft further.

16 is a sectional view showing a part of a compressor showing still another embodiment. The difference between the embodiment of Fig. 9 and the present embodiment is that the intermediate axes 112a and 112b are provided in the embodiment of Fig. 9, but the present embodiment eliminates the intermediate axis. That is, the first crank pin 13 and the second crank pin 14 extends to the inner direction of the partition plate 50 to eliminate the intermediate shaft. In this configuration, the crankshaft is less likely to be bent than in the embodiment shown in FIGS. 9 to 15 in which the intermediate shaft is provided. The hole provided in the partition plate 50 is larger than the maximum circle diameter of the rotational trajectory of the 1st crank pin 13 and the 2nd crank pin 14, and needs to prevent contact. In this case, the required seal area at the contact surface between the partition plate 50 and the first piston 10a and the second piston 10b becomes smaller by making the hole a larger diameter.

In order to ensure the seal area, the body thickness may be thickened in a direction in which the outer diameters of the first and second pistons 10a and 80 are increased. At this time, since the volume of a compression chamber becomes small, the internal diameters of the 1st, 2nd 1st cylinders 6a and 40 also become large, and it is good to maintain a volume. That is, the shape of the cylinder is flat. In this configuration, the diameter of the crankshaft 3 between the two compression chambers formed by the cylinders 6a and 6b is larger than the diameter of the crankshaft 3 fitted to the rotor 2a of the electric motor unit 2. I can make it big.

As shown in Fig. 16, when there is no intermediate axis, the rigidity of the axis is most high and it is difficult to bend, but when there is an intermediate axis, the shorter the length, the better. Therefore, what is necessary is just to have the structure which made the partition plate 50 thin when there exists an intermediate axis. In the case where the partition plate 50 is thinned up to the limit of the strain strength, it is most preferable that there is no intermediate axis.

17 is a view showing a part of a compressor showing still another embodiment. In the compressor of Fig. 16, the second crank pin 14 is configured such that the thickness of the partition plate 50 is extended. As a result, the rigidity of the crankshaft is heightened and bending deformation becomes difficult. The crank pin may be extended by the first crank pin. Also in this case, similarly to Fig. 14, the shorter the length is, the shorter the length is when there is an intermediate axis. Moreover, in the compressor provided with the thrust bearing mechanism in the crankshaft 3, this embodiment raises an effect. In addition, the seal area of the 1st and 2nd pistons 10a and 10b and the partition plate 50 is also the same as that of the method demonstrated by FIG.

FIG. 18 is an embodiment showing the configuration of reducing the bending deformation of the crankshaft in the embodiment shown in FIG. The first crank pin 13 extends in the direction of the upper sealing plate, and the second crank pin 14 extends in the direction of the lower sealing plate 60.

It is quite difficult to actually measure the amount of bending deformation of the crankshaft of the compressor described above. Therefore, the present inventors analyzed the bending deformation of the crankshaft by computer simulation using the finite element method (FEM). The crankshaft is subjected to a gas load, a force for the vane to press the roller by a screw (not shown), and a force required from the crankshaft and the centrifugal force of the roller. As a result, bending deformation occurs in the intermediate bearing between the first crank pin and the second crank pin.

Fig. 19 shows a computer simulation result 510 in the case where an ASHRAE / T condition when R 410 A is used as a refrigerant, i.e., a condition of a suction gas pressure of 0.996 MPa and a discharge gas pressure of 3.374 MPa is set as an input condition. Illustrated. In addition, the analysis model 610 input as a calculation initial value is also shown overlapping. The crankshaft causes bending deformation in the portion of the intermediate shaft. As an evaluation method of the analysis result of the crankshaft, how the displacement of the crankshaft of the part corresponding to the position of the upper end part of a main bearing was made with respect to the minimum value of a setting clearance as each ratio.

Fig. 20 is a diagram showing the relationship between the displacement B of the main bearing upper end portion of the crankshaft and the minimum value C of the set gap in the state where the crankshaft has caused bending deformation. FIG. 20 shows the crankshaft 710 of the bending deformation and the crankshaft 3 in the initial state.

Fig. 21 is an integrated view of the results of computer simulation. The abscissa is represented by X (kg) as the product of the elastic modulus of the crank material and the intermediate axis cross-sectional area (cross-sectional area perpendicular to the rotation axis of the intermediate axis, including the area of the hollow hole penetrating through the entire crankshaft). The longitudinal axis is shown as Y as the bearing clearance ratio by dividing the displacement B of the upper part of the main bearing of the crankshaft by the minimum value C of the set clearance.

In this figure, the main bearings contact when the bearing clearance ratio Y is 1 or more, and the larger the value, the greater the contact reaction force, thereby increasing the friction loss. If the bearing clearance ratio Y is less than 1, the shaft and the main bearing are not in contact.

The solid line shown in Fig. 21 is the external insertion line of the average value approximating the data point as an exponential function, and shows this equation at the same time. Moreover, the broken lines shown above and below the external insertion line of the average value are the external insertion lines when the approximate exponential functions become the maximum and the minimum, respectively. As a case where the bearing clearance ratio Y is smaller than 1, referring to the minimum external insertion line, it can be seen that the lamination X between the modulus of elasticity and the intermediate cross-sectional area is preferably 4.066 × 10 ⁶ kg or more as indicated by the F arrow in FIG. 21. . In addition, referring to the average external insertion line in consideration of the average design, it can be seen that X is preferably 5.163 × 10 ⁶ kg or more shown by E in FIG. In addition, referring to the maximum external insertion line in consideration of the safe case, it can be seen that X is better than 7.454 × 10 ⁶ kg shown by G in FIG. In addition, since the result data is calculated so that the gap is made the minimum value and the deformation in the analysis is maximized, it is usually sufficient if the condition indicated by F in FIG. 21 is satisfied. By setting the elastic modulus and the intermediate shaft cross-sectional area that satisfies these, the bending deformation of the shaft is reduced, the friction loss can be reduced, or the increase in the gap of the sealing plate, the roller, the partition plate, etc. can be suppressed, and the leakage of the compressed gas can be reduced. Can be.

The higher the elastic modulus of the crankshaft material is, the better it is difficult to bend, but considering the material cost and workability, it is a FC-based material that is cast steel. Is preferred. In addition, even if the material is within the range, the material having the lower elastic modulus is cheaper, so it is selected in relation to the intermediate shaft diameter. As a result, when a material having a high modulus of elasticity is used, it is made of special specifications and the like, and the cost is high.

According to the present invention, in the rotary compressor having a plurality of cylinders, the bending deformation between the crank pins can be reduced while the crankshaft diameter is designed to be small, and the crankshaft and the bearings, the crank pins and the piston inner surface or each Since the bias between the end plate and partition plate end face and the piston end face for sealing the cylinder is small, the friction loss is reduced and the loss of mechanical efficiency is reduced. In addition, since the extra clearance between the crankshaft, the bearing, the piston, the end plate and the partition plate end face can be made small, the leakage can be reduced and the reduction in volume efficiency can be suppressed. These effects can suppress the deterioration of performance due to the deformation of the crankshaft.

Claims (10)

The motor unit and the compressor unit are provided in a sealed container, the motor unit and the compressor unit are connected by a crankshaft, the crankshaft includes a first crank pin and a second crankpin eccentric with respect to the rotating shaft, and the compressor unit supports the crankshaft. And first and second cylinders provided between the secondary bearing and the primary and secondary bearings and partitioned by partition plates having through holes having an inner diameter larger than the outer diameter of the first or second crank pins. 1, the first cylinder and the second piston eccentrically interlocked in accordance with the rotation of the crankshaft in the second cylinder, the diameter of the crankshaft between the two compression chambers formed in the first cylinder and the second cylinder is fitted to the rotor of the motor unit A multiple cylinder rotary compressor, characterized in that it is larger than the diameter of the attached crankshaft. A main and secondary bearing in which the motor unit and the compressor unit are connected by a crankshaft in the sealed container, the crankshaft having a first crankpin and a second crankpin eccentric with respect to the rotating shaft, and wherein the compressor unit supports the crankshaft; First and second cylinders provided between the main and secondary bearings and partitioned by partition plates having a through hole having an inner diameter larger than an outer diameter of the first or second crank pin; As a multi-cylinder rotary compressor provided with the 1st, 2nd piston which eccentrically cooperates with the rotation of the said crankshaft in 2 cylinders, and the said crankshaft, A multi-cylinder rotary compressor provided on an intermediate shaft for connecting the first and second crank pins, extending in the eccentric direction of these crank pins and integrally formed on each crank pin. An electric motor part and a compressor part are connected by a crankshaft in an airtight container, the crankshaft having a first crankpin and a second crankpin eccentric with respect to the rotational axis, and the compressor part between the main and secondary bearings supporting the crankshaft. A first cylinder and a second cylinder, each of which is provided by a partition plate having a through hole of an inner diameter larger than an outer diameter of the first or second crank pin, and in the first and second cylinders of the crankshaft. A multi-cylinder rotary compressor having first and second pistons eccentrically interlocked with rotation and the crankshaft, It is provided in the intermediate shaft which connects between the said 1st and 2nd crank pin, extends in the eccentric direction of these crank pins, and is provided with the connection part integrally formed in each crank pin, The rotational trajectory of the outermost diameter of this connection part is provided. Is smaller than the inner diameter of the through hole of the partition plate. In the sealed container, the motor part and the compressor part are connected by a crankshaft, the crankshaft having a first crank and a second crankpin eccentric with respect to the rotating shaft, and the compressor part between the main and sub bearings supporting the crankshaft. And a first cylinder and a second cylinder partitioned by a partition plate having a through hole having an inner diameter larger than the outer diameter of the first or second crank pins, these first and second pistons, and the crankshaft. As a plural cylinder rotary compressor, A connection portion provided on an intermediate shaft that connects the first and second crank pins, extending in the eccentric direction of these crank pins, and integrally formed on each crank, and formed on one crank pin side; The thickness of the partition plate is set thinner than the distance between the surface opposite to this crank pin and the reference line of the surface opposite the partition plate of the other crank pin. A main and secondary bearing in which the motor unit and the compressor unit are connected by a crankshaft in the sealed container, the crankshaft having a first crankpin and a second crankpin eccentric with respect to the rotating shaft, and wherein the compressor unit supports the crankshaft; First and second cylinders provided between the main and secondary bearings and partitioned by partition plates having a through hole having an inner diameter larger than an outer diameter of the first or second crank pin; A multi-cylinder rotary compressor having first and second pistons eccentrically interlocked with the rotation of the crankshaft in the two cylinders, and the crankshaft, Among the intermediate shafts configured between the first crank pin and the second crank pin, a vertical cross section with respect to the rotation axis of the portion accommodated in the through hole of the partition plate is the first crank pin or the second crank pin. There is a part included in the vertical section with respect to the rotation axis of one crank pin, and not included in the part of the vertical section with respect to the rotation axis of the other crank pin, and the maximum value of the contour and the distance from the rotation axis core is A multi-cylinder rotary compressor comprising a connecting portion smaller than the inner radius of the through hole of the plate. An electric motor part and a compressor part are connected by a crank in a hermetic container, and the crankshaft includes a first crankpin and a second crankpin eccentric with respect to a rotating shaft, and between the first crankshaft and the second crankshaft. A first and a second cylinder having an intermediate shaft in the compressor section, the first and second cylinders being partitioned by the main and secondary bearings supporting the crankshaft, and partition plates provided between the main and secondary bearings, and the first and second cylinders. A two-cylinder rotary compressor comprising first and second rollers eccentric with rotation of the crankshaft and the crankshaft therein, The intermediate shaft has a stepped portion in the axial direction, and the radial cross section partitioned at the stepped portion of the intermediate shaft is configured to be larger than the overlapping portion of the radial cross section of the first crank pin and the second crank pin. 2-cylinder rotary compressor. An electric motor part and a compressor part are connected by a crankshaft in a hermetic container, and the crankshaft includes a first crankpin and a second crankpin eccentric with respect to a rotating shaft, and the first crankshaft and the second crankshaft First and second cylinders provided with intermediate shafts and partitioned by a main and secondary bearings for supporting the crankshaft, and the partition plate provided between the main and secondary bearings, and the first and second cylinders. As a multi-cylinder rotary compressor provided with the 1st, 2nd roller which eccentrically rotates with the said crankshaft, and the said crankshaft, The radial cross section of the intermediate shaft is larger than the overlapping portion of the radial cross section of the first crank pin and the second crank pin, and each of the centers of the intermediate shaft of the first crank pin side and the second crank pin side A multi-cylinder rotary compressor characterized by being eccentric in the rectangular direction of the crank pin. 8. The plurality of cylinders according to claim 7, wherein the stepped portion between the first crankpin side and the second crankpin side of the intermediate shaft is close to the first crankshaft provided on the main bearing side. Rotary compressor. An electric motor part and a compressor part are connected by a crankshaft in a hermetic container, and the crankshaft includes a first crankpin and a second crankpin eccentric with respect to a rotating shaft, and the first crankshaft and the second crankshaft First and second cylinders provided with intermediate shafts and partitioned by a main and secondary bearings for supporting the crankshaft, and the partition plate provided between the main and secondary bearings, and the first and second cylinders. As a multi-cylinder rotary compressor provided with the 1st, 2nd roller which eccentrically rotates with the said crankshaft, and the said crankshaft, The radial cross section of the intermediate shaft is larger than the overlapping portion of the radial cross section of the first crank pin and the second crank pin, and the intermediate shaft is processed concentrically with the first crank pin and / or the second crank pin. Multi-cylinder rotary compressor characterized in that. An electric motor part and a compressor part are connected by a crankshaft in a hermetically sealed container, and the crankshaft includes a first crankpin and a second crankpin eccentric with respect to a rotational axis, and the first crankshaft and the second crankshaft In the first and second cylinders and the first and second cylinders having an intermediate shaft therebetween, the compressor section being partitioned by a main and secondary bearing for supporting the crankshaft, and partition plates provided between the main and secondary bearings. A multi-cylinder rotary compressor comprising first and second rollers eccentric with rotation of the crankshaft, and the crankshaft, A plurality of cylinders, characterized in that the radial cross section of the intermediate shaft is larger than the overlapping portion of the radial cross section of the first crank pin and the second crank pin, and the first crank pin and the second crank pin are in contact with each other. Rotary compressor.