KR20100092389A - Positive displacement type pump and positive displacement type fluid machine with the same - Google Patents

Abstract

PURPOSE: A volumetric pump and a volumetric fluid machine comprising the same are provided to continuously perform pump operation and to reduce the manufacturing cost. CONSTITUTION: A volumetric pump comprises a casing(30c), a rolling cylinder(30b), a rotational piston(30a), and a rotation regulator. The casing comprises a casing room(30c4). The rolling cylinder comprises a pump groove and is installed in the casing room to be able to rotate. The rotational piston is loosely inserted into the pump groove to divide the pump groove to form two pump rooms. The rotational piston or the rolling cylinder makes pivotal movement or rotational movement by a driving source. The rolling cylinder is installed to be able to rotate around the rotary shaft of the cylinder with respect to the pivotal shaft of the piston which is the central shaft of the pivotal movement of the pivotal piston and the pump groove is linearly extended by crossing with the rotary shaft of the cylinder which is the central shaft of the rotational movement of the rolling cylinder. The pivotal piston is arranged to be able to reciprocate in the pump groove and to make rotation with respect to the pump casing and the pivotal radius of the pivotal piston is the same to the eccentricity of the rolling cylinder approximately. The rotation regulator maintains the pivoting of the pivotal piston and the rotation of the rolling cylinder and regulates the pivoting speed of the pivotal piston as two times of the rotating speed of the rolling cylinder.

Description

Volumetric pumps and volumetric fluid machines with them {POSITIVE DISPLACEMENT TYPE PUMP AND POSITIVE DISPLACEMENT TYPE FLUID MACHINE WITH THE SAME}

FIELD OF THE INVENTION The present invention relates to a rolling cylindrical volumetric pump, in which a rolling cylinder rotates with a pivoting swing piston, and a volumetric fluid machine having the same, in particular a small-volume, low-cost manufacturing operation that facilitates smooth pump operation. A pump and a volumetric fluid machine having the same.

As a rolling cylinder type | mold volume pump, like the hermetic compressor disclosed by Unexamined-Japanese-Patent No. 11-125191 (patent document 1), it is the structure which provides two or more pistons from which a rotating phase differs in one rolling cylinder. there was.

[Patent Document 1] Japanese Patent Application Laid-Open No. 11-125191

The rolling cylinder volumetric pump has a turning amount of the turning piston at an arbitrary time when the gap of each part is 0 and the turning radius Ep of the turning piston rotating shaft and the eccentricity Es of the rotating shaft of the cylinder coincide. Between (ΔΦ p) and the amount of rotation (Φs) of the rolling cylinder, the relationship of the following equation 1 is always established. In addition, the rotation amount of a rolling cylinder shall hereinafter refer to the rotation amount of a stationary body mainly as a rotation amount. In some cases, however, the rotation amount is referred to as the rotation amount. Similarly for the speed and the central axis, the rotation speed is appropriately changed to the rotation speed and the rotation axis to the rotation axis.

[Equation 1]

By dividing both sides of the equation 1 by the time t, the equation 1 can be expressed by the following equation 2.

[Equation 2]

In addition, when turning piston turning speed is shown by ohmp and rolling cylinder rotational speed by ohms, Formula 1 can be expressed by following Formula 3.

[Equation 3]

When the relationship of Expression 1 or Expression 3 is not established for some reason, it is difficult to recover to the normal relationship, and as a result, the pump operation cannot be continued, and the volumetric pump is locked.

By the way, in general, the swinging motion of the swinging piston is realized by the rotation of the crankshaft whose swing radius is the eccentric amount. For this reason, the rolling cylindrical displacement pump has the following characteristics.

First, since the eccentric pin portion of the crankshaft and the rolling cylinder need to satisfy the above-described relations of Equation 1 or Equation 3, a configuration that enables relative rotation (relative rotational speed is? P / 2) is essential. Specifically, a configuration in which the external shape of the swinging piston is made into a cylindrical shape to enable rotation between the swinging piston outer circumferential surface and the pump groove of the rolling cylinder, or a configuration that enables rotation between the swinging piston inner peripheral surface and the crankshaft eccentric pin portion is required. Do.

Secondly, in the rolling cylindrical displacement pump, since the turning radius Ep of the turning piston and the eccentric amount Es (the eccentricity from the turning center of the turning piston) of the rolling piston are the same, turning only once during one turning of the turning piston The rotating shaft of the piston and the rotating cylinder coincide.

In the rolling cylindrical volumetric pump, these characteristics are present. Therefore, a situation in which the above-described equations 1 or 3 do not hold in principle occurs by a mechanism as described later. The occurrence frequency in this case is thought to be low, and it has been considered that countermeasures can be made with a simple response. By the way, in the actual apparatus, the situation where said Formula 1 or Formula 3 does not hold frequently arises because of the clearance gap formed between elements, and the element arrangement | assembly (assembly), and the fundamental countermeasure for avoiding the lock of a pump operation | movement. This indispensable became clear by experiment.

Below, the principal mechanism by which said Formula 1 or Formula 3 is not established is demonstrated first, and then the mechanism in an actual apparatus is demonstrated.

First, a generation mechanism in which Equation 1 or Equation 3 above is not established in principle will be described with reference to Fig. 17 showing the principle pump operation. Fig. 17 shows the positions of the swinging piston and the rolling cylinder every 45 degrees when the eccentric pin portion is turned once in the clockwise direction. Here, the angle (bold letters) on the outer side shown in the drawing is the amount of revolution of the eccentric pin part, but since it considers an ideal case where there is no gap between the eccentric pin part and the swing piston, it becomes equal to φp. Incidentally, the inner angle (thin letter) is the amount of revolution of the rotating piston rotating shaft (eccentric pin center axis) seen from the rotating shaft of the rolling cylinder, but it is equivalent to Φ s because it considers an ideal case where there is no gap between the turning piston and the pump groove. Become. At each of these timings, the rotational amount Φs of the rolling cylinder relative to the turning amount Φp of the turning piston is determined by the angle shown in FIG. Equation 1 or 3 holds.

The exception is the timing (Fig. 17 (I5)) in which the above-described cylinder rotating shaft (rotating shaft) coincides with the rotating piston rotating shaft as a second feature of the rolling cylindrical displacement pump. Since the rotating shafts of the two pump parts coincide, the rolling piston does not rotate (Ωp = 0) from the above-described first feature, and the rolling cylinder can rotate (Ωs ≠ 0). Thereby, the problem of (II) may arise through the situation of following (I).

(I) The pump stops at the timing when the rotating shaft of the rolling cylinder and the rotating shaft of the swinging piston coincide.

(II) Any rotating torque is applied to the rolling cylinder, and under the above situation (I), only the rolling cylinder rotates without the turning piston turning.

When the state changes according to this mechanism, the same situation as in Fig. 18 in which Equation 1 or Equation 3 above is not satisfied is satisfied. In addition, once in this situation, even if the turning piston attempts to turn, the force applied to the rolling cylinder is equal to the rolling cylinder of the rolling cylinder (regardless of the rotation angle of the rolling cylinder generated in the above (II)) as shown by the arrow in FIG. It occurs in the direction of passing through the axis of rotation. For this reason, the torque which rotates a rolling cylinder does not generate | occur | produce, and a rolling cylinder rotates and it cannot return naturally to a normal posture which Formula 1 or Formula 3 mentioned above hold | maintains. As a result, the volumetric pump is locked. In order to avoid this pump lock, it turned out that avoiding said (I) is effective and convenient. For example, what is necessary is just to determine the setting angle of a motor so that a pump may not stop in the state of said (I) using cogging of the motor which is a drive source.

Next, the mechanism which causes the independence of said Formula 1 or Formula 3 in the case of an actual apparatus is demonstrated. According to the above-described principle mechanism, the lock is indispensable in the generation of (I), but in fact, the lock frequently occurs even without the occurrence of the (I). That is, there is a lock generation mechanism that cannot be derived from the principle considerations. For this reason, the lock generation mechanism in an actual device is considered again. In other words, consider the mechanism by which the rolling cylinder causes rotation outside of Equation 1 or 3 above during the pump operation.

The main mechanism which causes the clearance between an eccentric pin part and a turning piston is demonstrated using FIG. 19, FIG. It is a figure explaining the effect which the clearance gap between an eccentric pin part and a turning piston has on pump operation. In Fig. 19, the gap is exaggerated and shown. (R3)-(R5) of FIG. 19 shows the same timing as (I3)-(I5) of FIG.

Since the revolving piston partitions the suction side pump chamber and the discharge side pump chamber, force is applied from the working fluid in the direction of the arrow in FIG. 19. For this reason, the turning piston is displaced in the direction of the force by the gap between the eccentric pin portion and the turning piston. Fig. 19 shows the displacement exaggerated, and shows the rotating piston rotating shaft in a black circle in that case. In addition, the central axis of the eccentric pin section in each case was also plotted in a rectangle. After (R5) in Fig. 19, since the eccentric pin portion moves to the lower left side, the black circle after this necessarily moves to the left side. As a result, the rotating shaft trajectory of the revolving piston does not pass the rotational axis of the rolling cylinder but passes the left side thereof. 20 shows an enlarged trajectory of the revolving piston rotating shaft.

In addition, as a force acting on the swinging piston, the reaction force of the force applied to the rolling cylinder to rotate the rolling cylinder, together with the force from the working fluid, acts in a direction perpendicular to the pump groove. Therefore, in FIG. 20, the rotation axis position of the turning piston which also added the displacement by this reaction force was plotted in a white circle. By reaction force from this rolling cylinder, it turns out that the rotating shaft trace of a turning piston shifts to the left side further. Since the linear direction connecting the rotating piston rotating shaft and the cylinder rotating shaft becomes the direction of the pump groove (i.e., the rotational phase of the rolling cylinder), the rolling cylinder is near this (near the position where the rotating shaft of the rotating piston and the rotating shaft of the rotating piston coincide). It can be seen that rotation outside of Equation 1 or 3 may occur (specifically, the rotation direction is suddenly reversed from the clockwise direction to the counterclockwise direction). In other words, it can be seen that the rotation between the eccentric pin portion and the swinging piston may cause rotation that deviates from the above-described equation 1 or equation 3 during the pump operation and may be locked.

On the contrary, from this consideration, the possibility of a normal pump operation without locking even when both do not coincide near the position where the rotational axis of the rolling cylinder and the rotational axis of the swinging piston coincide can also be explained as follows. That is, if the rotational axis of the rolling cylinder and the rotational axis trajectory of the swinging piston are smaller than the gap between the swinging piston and the pumping groove of the rolling cylinder, the position of the swinging piston is shifted from the central axis of the pumping groove without changing the direction of the pumping groove. The rotating shaft can pass through, and the pump operation can be continued.

Here, the cylinder rotating shaft described above and the rotating shaft of the swinging piston mean an instantaneous axis at the moment when the rotating shaft of the swinging piston passes a position shifted from the central axis of the pump groove.

As described above, in the case of an actual apparatus having a gap between pump components, a situation in which Equation 1 or Equation 3 above does not hold due to a gap that has been omitted in principle is considered. Thereby, the effect of correct | amending the situation where said Formula 1 or Formula 3 did not hold is also produced. Such a situation also occurs between other element gaps not described above, or due to an arrangement error of the element (the turning radius Ep of the turning piston rotating shaft and the eccentricity Es of the cylinder rotating shaft do not coincide). In this way, the above formula 1 or formula 3 is not established, and it is determined by a combination of various situations that the operation lock of the pump occurs between the element gaps or under the condition of element arrangement. It is extremely difficult.

Therefore, in order to avoid the operation lock of the rolling cylindrical displacement pump and to continue the smooth operation, it is essential to provide a rotation defining means which always establishes the above-described relationship of the equation 1 or equation 3 between the rolling cylinder and the turning piston. Do. At the same time, excessive restraint of the relationship of the above-described formula (1) or (3) is required, and rotation defining means under the same conditions as the following (A) and (B) is required.

(A) A mechanism other than the suspension rotating mechanism which defines a rolling cylinder and a swivel piston in relation of said Formula 1 or Formula 3 is provided.

(B) In the vicinity of the timing at which the rotational axis of the rolling cylinder coincides with the rotational axis of the swinging piston, the restricting action which defines the arrangement of the two cylinders in relation to the above-described equation 1 or 3 is the strongest, and as the deviation from the timing occurs, the regulation Action is alleviated.

In order to avoid the operation lock, the patent document 1 takes a countermeasure that two swing pistons having a 180-degree swing phase are provided in one rolling cylinder. This means that as the rotating shaft of one of the turning pistons approaches the rotating shaft of the rolling cylinder, the rotating shaft of the other turning piston is spaced most apart from the rotating shaft of the rolling cylinder, and the combination of the turning piston and the pump groove is mutually the condition (A ), (B) fulfills the role of rotation defining means, and is one of the books.

However, this technique requires two swinging pistons, which is not suitable for small capacity. In addition, since the number of parts increases, the machining cost is increased. In addition, there is a significant decrease in the granularity described below. The crankshaft has two eccentric pin portions which are eccentric in 180 degrees different directions, but it is almost impossible to insert them into a rolling cylinder in which two pump grooves are set in directions different by 90 degrees. For example, if the eccentric amount or diameter of the eccentric pin portion is designed to be extremely small compared to the pump groove width, it becomes geometrically possible, but the outer diameter and height of the swinging piston must be increased to secure the displacement, and the reliability of the eccentric pin portion is increased by increasing the load. However, a significant decrease in performance occurs. For this reason, after dividing a rolling cylinder, after inserting them into the eccentric pin part of a crankshaft, respectively, the complicated assembly process of integrating with high positional accuracy is needed, and there existed a problem of a raise of manufacturing cost.

SUMMARY OF THE INVENTION An object of the present invention is to provide a volumetric pump and a volumetric fluid machine using the same, which are suitable for a small capacity and can reduce the manufacturing cost while allowing the pump operation to be continued smoothly.

According to a first aspect of the present invention for achieving the above object, a pump casing having a casing chamber, a rolling cylinder having a pump groove disposed in the casing chamber so as to be rotatable, and loosely fitted in the pump groove A pivoting piston which defines two pump chambers by partitioning the pump groove, wherein the pivoting piston or the rolling cylinder is pivoted or rotated by a drive source, and the rolling cylinder is a piston pivot which is the central axis of the pivoting motion of the pivoting piston. With respect to the cylinder rotation axis whose eccentric m2 is Es, the pump groove extends in a straight line across the cylinder rotation axis which is the center axis of the rotational motion of the rolling cylinder, A piston reciprocates in the pump groove, while at the same time turning the turning radius to Ep relative to the pump casing. The rotating radius Ep of the rotating piston and the eccentricity Es of the rolling cylinder, which are arranged to rotate in rotation, are approximately equal to the number of the rotating pistons fitted to the rolling cylinder in a volumetric pump. It is set as one and the rotation provision means which maintains the rotation of the said rotation piston and the rotation of the said rolling cylinder, and prescribes the rotation speed of the said rotation piston at twice the rotation speed of the said rolling cylinder is provided.

The more preferable specific structural example in this 1st aspect of this invention is as follows.

(1) The prescribed degree of function of the rotation defining means is to reduce the case where the piston pivot shaft is rotated 180 degrees by the pivot angle compared with when the piston pivot shaft coincides with the cylinder rotation shaft.

(2) the swinging motion of the swinging piston is realized in the crankshaft whose eccentricity is Ec,

The rotating shaft of the rolling cylinder is realized in a bearing portion having an eccentric amount of Eb,

The clearance between the turning piston and the pump groove is equal to or less than twice the difference between the eccentricity Ec of the crankshaft Ec and the bearing eccentricity Eb.

(3) The rotation defining means includes rotation synchronizing means for synchronizing rotation of the pivoting piston with rotation of the rolling cylinder, and piston rotation defining means for defining the rotational speed of the pivoting piston at twice the rotational speed.

(4) The rotation synchronizing means is provided by providing a side flat portion in sliding contact with two sides of the pump groove, respectively, in the pivot piston.

(5) The piston rotation defining means is always turned on the turning trajectory of the turning piston rotating shaft so that a stop point at a position different from the cylinder rotating shaft always lies in a straight line fixed on the turning piston passing through the turning piston rotating shaft. A slider mechanism for defining the movement of the piston is provided.

(6) The stop point is provided at a position rotated 180 degrees from the position of the cylinder rotation shaft on the turning motion trajectory.

(7) the slider mechanism includes a floating slider and a rotation guide,

The floating slider is realized by a position fixing cylinder fixedly arranged at a position corresponding to the stop point of the pump casing,

The rotating guide is realized by forming a guide groove having a width equal to the diameter of the positioning cylinder in order to form the positioning cylinder and the sliding pair with the straight line as the center line on the pivoting piston.

(8) The floating slider has a fixed central axis and a slider member rotatably inserted in the fixed central axis.

(9) The guide groove is a passage for working fluid.

(10) The guide groove is extended to the side flat portion.

(11) Mounted in said volumetric fluid machine as a source of oil to each part of the volumetric fluid machine.

According to a second aspect of the present invention, there is provided a drive source, a crankshaft driven by the drive source, a displacement pump driven by the crankshaft, and the displacement pump includes a pump casing having a casing chamber, and a pump groove. And a rolling cylinder disposed in the casing chamber so as to be rotatable, and a turning piston loosely fitted in the pump groove to define two pump chambers to form two pump chambers, wherein the turning piston or the rolling cylinder includes: The rolling source is pivoted or rotated by the drive source, and the rolling cylinder is disposed so as to be rotatable about a cylinder rotational axis having an eccentric amount of Es with respect to the piston pivotal axis which is the central axis of the pivotal motion of the pivoting piston. Extends in a straight line across the cylinder axis of rotation, which is the central axis of rotation of the rolling cylinder, And the pivoting piston reciprocates in the pump groove, and is rotatably disposed so that the pivoting radius pivots with respect to the pump casing, and the pivoting radius Ep of the pivoting piston and the rolling cylinder. The amount of eccentricity Es is approximately equal in volumetric fluid machine, wherein the number of the pivoting pistons to be fitted to the rolling cylinder is one, and the pivoting of the pivoting piston and the rotation of the rolling cylinders are maintained. The provision of the rotation provision means which prescribes | requires turning speed twice the rotation speed of the said rolling cylinder is provided.

According to the present invention, it is possible to provide a volumetric pump and a volumetric fluid machine using the same, which are suitable for a small capacity and can reduce the manufacturing cost while allowing the pump operation to be continued smoothly.

1 is a longitudinal sectional view of a scroll compressor according to a first embodiment of the present invention.
FIG. 2 is a detailed enlarged view of the vicinity of a back pressure chamber of the scroll compressor of FIG. 1. FIG.
3 is an enlarged longitudinal sectional view of an oil supply pump of the scroll compressor of FIG.
4 is an enlarged view of the KK cross section of FIG.
5 is a VV cross-sectional view of FIG.
6A is an enlarged longitudinal sectional view of a first modification near the position fixing cylinder of the oil feed pump of the scroll compressor of FIG. 1;
6B is an enlarged longitudinal sectional view of a second modification near the position fixing cylinder of the oil feed pump of the scroll compressor of FIG. 1;
7 is a plan view of the base plate of the oil feed pump of the scroll compressor of FIG.
8 is a cross-sectional view in the case where the installation position of the position fixing cylinder and the turning piston side shape of the lubrication pump of the scroll compressor of FIG. 1 are changed.
9 is a partially exploded perspective view of the oil supply pump of the scroll compressor of FIG.
10 is an explanatory view of the operation of the oil supply pump of the scroll compressor of FIG. 1.
11 is an explanatory diagram of rotation synchronizing means and piston rotation defining means of an oil supply pump of the scroll compressor of FIG. 1;
12 is an enlarged longitudinal sectional view of a position fixing cylinder of a fuel pump of a scroll compressor according to a second embodiment of the present invention.
13 is an enlarged LL cross-sectional view of FIG. 12.
14 is an enlarged cross sectional view of a position fixing cylinder of a fuel pump of a scroll compressor according to a third embodiment of the present invention.
15 is an enlarged detail of the vicinity of a back pressure chamber of the scroll compressor according to the fourth embodiment of the present invention.
Fig. 16 is an enlarged longitudinal sectional view of an oil supply pump of a scroll compressor according to a fourth embodiment of the present invention.
17 is an explanatory view of the principle pump operation of the rolling cylindrical displacement pump.
18 is an explanatory diagram of a pump operation deviating from the principle in a rolling cylindrical displacement pump.
19 is an explanatory diagram showing an example of a pump operation deviating from the principle that actually occurs in a rolling cylindrical displacement pump.
FIG. 20 shows a pivoting piston center trajectory during pump operation that deviates from the principle that actually occurs in a rolling cylindrical displacement pump. FIG.
FIG. 21 is an explanatory diagram of a restricting operation diagram of a rolling cylinder rotation angle of a fuel pump of the scroll compressor of FIG. 1; FIG.
FIG. 22 is a view showing the trajectory of the swing piston when the clearance between the pump groove of the oil feed pump of the scroll compressor of FIG. 1 and the swing piston is twice the difference between Eb and Ec. FIG.

EMBODIMENT OF THE INVENTION Hereinafter, several embodiment when the volumetric pump of this invention is mounted as a bearing of the scroll compressor which is a volumetric fluid machine, or a lubrication pump (with working fluid as oil) to a compression chamber is used, using drawings, Explain. The same code | symbol in the drawing of each embodiment shows the same or equivalent. In addition, this invention includes making each embodiment more effective by combining suitably as needed. In addition, since the working fluid in the volumetric pump of this invention is limited to oil, it calls it oil now, and the name of a working fluid shall refer to working fluid in a scroll compressor.

(1st embodiment)

A first embodiment in which an oil storage unit is provided in a casing, and a volumetric pump according to the present invention is mounted as a lubrication pump in a scroll compressor in which the casing is at suction pressure will be described with reference to FIGS. 1 to 11. As the case where the so-called low pressure chamber type, in which the casing inside becomes the suction pressure, is employed, a case where a flammable gas is used as a working fluid is mentioned. For example, hydrocarbon fluids, such as propane and butane, correspond to it. This is an effective means to reduce the total amount of working fluid encapsulated throughout the device including the compressor, from a safety point of view.

First, the whole structure and operation | movement of the scroll compressor of this embodiment, the longitudinal cross-sectional view of the scroll compressor which concerns on 1st Embodiment of this invention of FIG. 1, and the enlarged view of the vicinity of the back pressure chamber of FIG. 2 (N part of FIG. 1) are used. Will be explained. In addition, the detail of the oil supply pump of this embodiment is demonstrated later using FIGS. 3-11.

A suction pipe 53 penetrates the side surface of the casing 8, and a working fluid of suction pressure is introduced into the casing 8 through the suction pipe 53. And this working fluid is guided from the suction port 2e opened to the side of the fixed scroll 2 to the compression chamber 100 formed between the fixed scroll 2 and the turning scroll 3. Since the compression chamber 100 reduces the volume while moving from the outer circumference to the inner circumference by the swing motion of the swing scroll 3, the working fluid in the compression chamber 100 is compressed. Here, the swing motion of the swing scroll 3 is realized by rotating the crankshaft 6 to which the swing scroll 3 is connected by the motor 7 and preventing the rotation by the Oldham ring 5.

In the fixed scroll 2, a bypass valve 22 for preventing overcompression or liquid compression is provided on the upper surface, and a discharge port 2d through which the compressed working fluid is discharged is formed. The fixed scroll 2 is screwed to the frame 4. Between the back surface of the revolving scroll 3 and the frame 4, a back pressure chamber 110 is formed to be an intermediate pressure (pressure between the suction pressure and the discharge pressure, hereinafter referred to as back pressure).

The crankshaft 6 has an axial position defined by the shaft flange portion 6h mounted on the frame shaft thrust protrusion 6p, while the upper portion is supported by the main bearing 24, and the lower portion of the crank shaft 25 is provided. Is supported. The eccentric pin part 6a provided in the upper end part of the crankshaft 6 is inserted in the turning bearing 23 of the turning scroll 3. Here, the sub-bearing 25 is composed of a ball 25a and a ball holder 25b. The ball holder 25b is welded to the minor shaft support part 50 fixed to the casing 8. By the structure which consists of this ball 25a and the ball holder 25b, the inclination of the subshaft support part 50 can be tolerated to some extent.

These bearings are supplied with oil pumped up by the oil supply pump 30 from the oil reservoir 125 under the casing 8 through the oil supply hole 6b of the crank shaft 6. The oil supplied to the slewing bearing 23 and the main bearing 24 enters the back pressure chamber 110 and then passes through the back pressure chamber outlet passage 135 through the frame 4 to the side of the frame 4. It is discharged and finally returned to the oil reservoir 125. Here, the back pressure control valve 26 is provided in the middle of the back pressure chamber outflow path 135. The back pressure control valve 26 maintains the pressure in the back pressure chamber 110 at a desired back pressure. By this back pressure, the turning scroll 3 is pressed by the fixed scroll 2 at the time of compression.

On the other hand, in order to improve the sealing property of the compression chamber 100, a small amount of oil is supplied from the slewing bearing chamber 115 to the compression chamber 100 through the compression chamber oil supply passage 130. This oil becomes discharge oil and is discharged from the discharge port 2d or the bypass valve 22 to the upper part of the fixed scroll 2 together with the working fluid.

The discharge oil separation oil conveyance cylinder 55 is screwed to the upper part of the fixed scroll 2, and the discharge chamber 120 is formed. In the upper part of the discharge oil separation oil conveyance cylinder 55, the discharge cover 51 which has the discharge pipe 52 which protruded further is screwed, and the oil separation chamber 90 and the oil conveyance chamber 95 are formed. .

The working fluid introduced into the discharge chamber 120 is led to the oil separation chamber 90 to separate the oil mixed into the working fluid, and then flows out of the compressor 1 through the discharge pipe 52. The oil separated in the oil separation chamber 90 flows into the oil transfer chamber 95. And the oil which flowed into the oil conveyance chamber 95 is the oil conveyance path 80 which connects the oil conveyance chamber 95 and the oil storage part 125, and the oil conveyance amount adjustment valve 70 provided in the middle. The oil reservoir 125 is returned to the oil storage unit 125 via. The oil conveyance amount adjustment valve 70 flows into the oil conveyance chamber 95 while always securing a small amount of oil in the oil conveyance chamber 95 in order to seal the discharge side and the suction side of both sides of the oil conveyance path 80. It is a valve that serves to return the flow rate and the same amount to the oil reservoir (125). This oil conveyance amount adjustment valve 70 is realized by an electromagnetic valve or the like which controls the opening degree by a sensor signal including information of a float valve and separation flow rate.

The oil supply pump 30 of this embodiment transfers the oil in the oil storage part 125 to each bearing part of the crankshaft 6 comprised by the sub bearing 25, the main bearing 24, and the turning bearing 23. As shown in FIG. In addition to the inherent role of supplying or supplying to the compression chamber 100 for improving the sealability of the compression chamber, it also plays a role of supplying the back pressure chamber 110 to generate back pressure. For this reason, the oil supply pump 30 is responsible for not only a flow volume but also a pressure boost.

In this embodiment, the separation oil back pressure chamber introduction path 500 which introduces a part of the separation oil which flowed into the oil conveyance chamber 95 into the back pressure chamber 110 is provided. And the separation oil branch means 501 which adjusts the flow volume which flows oil into this separation oil back pressure chamber introduction path 500 is provided. Even when the back pressure boosting amount by the oil supply pump 30 is insufficient, the back pressure can be increased by inserting the separation oil of the discharge pressure into the back pressure chamber 110, so that the deterioration of the compressor performance due to the insufficient back pressure can be avoided. It is effective. As the simplest means for realizing this separating oil branching means 501, there are branching pipes having different pipe diameters.

In addition, the separation oil branching means 501 and the back pressure control valve 26 may be integrated to provide a separation oil introduction back pressure control valve that operates to introduce the separation oil into the back pressure chamber 110 when the back pressure does not rise. . In this case, the separation oil is put into the back pressure chamber 110 only when the back pressure rise by the oil supply pump is not performed. For this reason, it is not necessary to always put high temperature separation oil into the back pressure chamber 110, and heating of the compression chamber 100 is suppressed and there exists an effect that a compressor performance improves.

Next, the detailed structure and operation | movement of the oil supply pump 30 are demonstrated in detail using FIGS. 3-11, 21, and 22. FIG. This oil supply pump 30 is a rolling cylinder type | mold type pump which uses the motor 7 which is the power source of the scroll compressor 1 as a drive source, and makes the turning piston 30a the drive side, and the rolling cylinder 30b the manual side.

First, the structure of the oil supply pump 30 is demonstrated using FIGS. 3 is a longitudinal cross-sectional view of the oil supply pump 30 (an enlarged detail of the portion M of FIG. 1, HH cross-sectional view of FIG. 4), FIG. 4 is a cross-sectional view (KK cross-sectional view of FIG. 3), and FIG. 3 and other longitudinal cross-sectional view of the oil supply pump 30 (VV sectional drawing of FIG. 4), FIG. 6A is an enlarged longitudinal cross-sectional view of the fixing part of the position fixing cylinder 30p separate from the base plate 30c1, FIG. 6B is FIG. 6A 8 is an enlarged longitudinal sectional view of the fixing portion of the base plate 30c1 and a separate position fixing cylinder 30p, FIG. 7 is a plan view of the base plate 30c1, and FIG. 8 is an installation position of the position fixing cylinder 30p. 9 is a cross-sectional view of the oil supply pump 30 in the case where the lateral shape of the swing piston 30a is changed.

A pump shaft portion 6f having a thin diameter is provided at the lower end portion of the crankshaft 6, and a pump eccentric portion 6f1 is provided below the pump shaft portion 6f. This pump eccentric part 6f1 is loosely fitted in the piston bearing hole 30a6 formed in the center part of the turning piston 30a. The swinging piston 30a is pivoted by the turning radius Ep by the eccentric rotation of the pump eccentric part 6f1 of the crankshaft 6. The oil supply hole 6b is opened in the lower end surface of the pump eccentric part 6f1.

The rolling cylinder 30b which is free to rotate which makes the axis | shaft eccentric by the eccentric amount Es substantially equal to the turning radius Ep with respect to the turning axis (alpha) of the turning piston 30a is provided. The rolling cylinder 30b has a pump groove 30b1 and is disposed in the casing chamber 30c4 of the pump casing 30c so as to be rotatable. The pump groove 30b1 crosses the cylinder rotation axis γ which is the central axis of the rotational movement of the rolling cylinder 30b and extends in a straight line shape. The swinging piston 30a is loosely fitted in the pump groove 30b1 and arranged to reciprocate, and the space on both sides of the swinging piston 30a is formed as two pump chambers 140 by partitioning the pump groove 30b1. Doing. The number of the turning pistons 30a fitted to the rolling cylinder 30b is one.

Here, the space | interval of the central axis (beta) of the pump eccentric part 6f1, and the rotating shaft (alpha) of the crankshaft 6 roughly defines the turning radius Ep of the turning piston 30a. By the way, since there is a clearance gap between the pump eccentric part 6f1 and the piston bearing hole 30a6, the radial force applied to the turning piston 30a generally (the same force as described in the problem to be solved by the above-described invention). 19 and 20), the turning piston 30a is displaced. For this reason, the space | interval Ec (notation of Ec) of the center axis (beta) of the pump eccentric part 6f1 and the crankshaft rotation axis (alpha) is the center axis of the piston bearing hole 30a6 (this is called the rotating shaft of a turning piston). Can be regarded and denoted as β ') and shifted from the turning radius Ep, which is the distance between the crankshaft rotational axis α.

As is apparent from Fig. 9, the swing piston 30a has a planar end 30a5 at one end face (lower end face) thereof, and at the same time, two parallel side flat parts 30h and two side flat parts. The side surface has two side cylindrical surfaces 30a4 which connect 30h. As is apparent from FIG. 8, the two side surface cylindrical surfaces 30a4 are shifted from each other to match the cylindrical surface forming the cylinder chamber 30c4. Thereby, since the volume of the pump chamber 140 formed is reduced to O in principle, dead volume is lost and the performance is improved.

The flat end portion 30a5 communicates with the lower surface lower space and the piston bearing hole 30a6, and has a guide groove 30g extending over the two side flat portions 30h. The width of this guide groove 30g is constant. As described above, the rolling cylinder 30b, which is freely rotatable, makes the axis eccentric from the piston pivot axis α as the rotation axis γ by an eccentric amount Es approximately equal to the pivot radius Ep of the swing piston 30a. It is. Thereby, as shown in FIG. 8, the cylinder rotating shaft (gamma) comes on the turning trace of the rotating piston rotating shaft (beta '). As a result, while the turning piston 30a is turning once, the timing at which the turning piston rotating shaft β 'and the cylinder rotating shaft γ coincide only once occurs.

The rolling cylinder 30b has the end plate part 30b4 in the axial upper part, and has the pump groove 30b1 in the axial lower part. The turning piston 30a is attached to the pump groove 30b1 so that the side flat part 30h of the turning piston 30a and the side surface of the pump groove 30b1 may slide and reciprocate. Thereby, the pump groove 30b1 is divided into two spaces, and each space becomes the pump chamber 140. The pump chamber 140 is spaced apart from the internal space of the casing 8 by the pump casing 30c.

The pump casing 30c is composed of a base plate 30c1 and a cover 30c2 which are respectively provided on the lower surface and the upper surface side of the rolling cylinder 30b, and a pump cylinder 30c3 that supports the rolling cylinder while being connected to them. . The base plate 30c1 has the suction indentation 30s1 provided in the upper surface and the pump suction hole 30s2 which penetrates the base plate 30c1, as shown in FIG. The pump suction flow path 30s which sucks up oil from this is comprised. The base plate 30c1 also has a pump discharge recess 30d1 and a pump discharge groove 30d2 provided on the upper surface thereof, and feeds oil into the oil supply hole 6b opened by the lower end of the crankshaft 6 by these. The pump discharge flow path 30d is formed.

In the center part of the base plate 30c1, the position fixing cylinder 30p which protrudes upwards is provided integrally with the base plate 30c1. This position fixing cylinder 30p is provided in the position which rotated 180 degrees from the cylinder rotating shaft (gamma) on the turning trace of the rotating shaft (beta ') of the turning piston 30a. 6A and 6B, the base plate 30c1 and the position fixing cylinders 30p-1 and 30p-2 separate from each other may be provided. The position fixing cylinder 30p-1 is comprised from the cylinder of the same diameter, and is pressed in and fixed to the hole of the base plate 30c1 from the top. The position fixing cylinder 30p-1 is comprised from the cylinder which has a flange in a lower end part, and is pressed in and fixed to the hole of the base plate 30c1 from below.

In the pump casing 30c of this embodiment, as can be seen from FIG. 3, FIG. 5, the upper pump casing 30c23 which integrated the cover 30c2 and the pump cylinder 30c3 is used. Thereby, there is an effect that the number of parts is reduced and assembly performance is improved. The upper pump casing 30c23 is formed with a minimum hole necessary for passing through the pump eccentric 6f1.

In actual assembling of the oil supply pump 30, as shown in FIG. 9, first, the pump eccentric part 6f1 is made to pass through the hole of the upper pump casing member 30c23, and then the rolling cylinder 30b and pivoting are performed. The piston 30a is assembled, and the upper pump casing 30c23 is temporarily fixed to the ball holder 25b (may be the part bearing support plate 50).

In this state, the base plate 30c1 is fixed to the upper pump casing 30c23 via the base plate fixing screw 30m while allowing the position fixing cylinder 30p to be inserted into the guide groove 30g. At this time, the knock pin is inserted so that the positioning hole 30i2 of the base plate 30c1 and the positioning hole 30i2 of the upper pump casing 30c23 fit, and the pump suction flow path 30s and the pump discharge flow path 30d are inserted. Increase the set position accuracy.

Thereafter, the pump cylinder fixing screw 30k is main body tightened while the crankshaft 6 is rotated to fix the pump cylinder 30c3 to the ball holder 25b. Thereby, since the positional accuracy of the pump cylinder 30c3 can be made high, the positional accuracy of the cylinder rotating shaft (gamma) is improved, the operation of the oil supply pump 30 can be made smooth, and the performance of an oil supply pump improves. It works. Here, the method for rotating the crankshaft 6 includes rotating the motor 7 at a low speed, sucking air from a suction port 2e with a vacuum pump, and rotating the swinging scroll member 3. It is available.

Next, the operation of the oil supply pump 30 will be described with reference to FIGS. 10, 11, 21 and 22. FIG. 10 is an operation explanatory diagram of the oil supply pump 30, which illustrates the pump operation during the pump chamber 140 in one stroke in the same cross section as in FIG. 4. FIG. 11 is an explanatory view of the operation when the rotational centers of the turning piston 30a and the rolling cylinder 30b coincide with each other, FIG. 21 is an explanatory view of a restrictive action diagram of the rolling cylinder rotation angle, and FIG. 22 is a pump groove 30b1. The clearance between the swing piston 30a is the trajectory of the swing piston 30a when the gap between the swing piston 30a is twice the difference between Eb and Ec.

As shown in FIG. 10, during one stroke of the pump chamber 140, the crankshaft 6 rotates two times (in the direction of a circular arrow). In addition, in FIG. 10, the cross-sectional change is shown every time the crankshaft 6 rotates 22.5 degrees, and hatching which shows the cross section of each component is abbreviate | omitted.

As described above, two pump chambers 140 are formed. Since these two pump chambers 140 are out of phase with each other, when one pump chamber is a suction stroke, the other pump chamber becomes a discharge stroke, but the operation change is the same. For this reason, the pump operation | movement is demonstrated, paying attention to one pump chamber (cross-hatching pump chamber of FIG. 10). In addition, when the pump chamber 140 is in a suction stroke, it is called the suction pump chamber 140s, and when it is in a discharge stroke, it is called discharge pump chamber 140d.

(1) to (10) in FIG. 10 (indicated by numerals in FIG. 10 as numerals given in parentheses) are suction strokes, and the oil in the oil storage part 125 is drawn in the suction flow path 30s. It is a process of sucking up to the suction pump chamber 140s through. 10 (11)-(16) is a process of discharging the oil of the discharge pump chamber 140d to the oil supply hole 6b by the discharge flow path 30d and the guide groove 30g in a discharge stroke. In principle (the configuration without the position fixing cylinder 30p, the guide groove 30g, and the side flat portion 30h newly installed in the present embodiment), the principle (the gap between each part is reduced to the limit and there is no assembly error) Pump operation and problems have already been described in the problem to be solved by the invention (see Figs. 17 and 18), and also the pump operation and problems of the actual device with gaps or errors in assembling each part Figs. As already explained in 20.

For this reason, after this, while using FIG. 10 or FIG. 21 centering on FIG. 11, it narrows down to the position fixing cylinder 30p, the guide groove 30g, and the side flat part 30h newly installed in this embodiment, and demonstrates. do. That is, it demonstrates that the position fixing cylinder 30p, the guide groove 30g, and the side flat part 30h become rotation defining means which satisfy | fills the said conditions (A) and (B). First, the mechanism newly satisfying the above conditions (A) and (B) can be configured by the elements newly installed in the present embodiment. Next, the above conditions (B ) Satisfies even more satisfactorily.

In order to satisfy condition (A) "the rotational speed of the turning piston 30a is always prescribed | regulated at twice the rotational speed of the rolling cylinder 30b", first, the rotational speed of the rolling cylinder 30b is set to the rotational piston 30a. By providing rotation synchronizing means for synchronizing with the rotating speed, the condition (A) is converted into the prescribed condition of only the turning piston 30a. Specifically, it changes to the content of providing the piston rotation defining means which defines the turning speed of the turning piston 30a at twice the rotating speed.

The rotation synchronization means of this embodiment is implemented by providing two side flat parts 30h on the side surface of the turning piston 30a so as to be in sliding contact with the pump groove 30b1 of the rolling cylinder 30b. This rotation synchronizing means is a seal | sticker part which partitions the suction pump chamber 140s and the discharge pump chamber 140d, and since it became surface seal from the conventional line seal, it also has the effect of improving sealing property.

The piston rotation defining means, which is another constituent means, is a guide groove (guide groove) passing a positioning cylinder 30p disposed on the swinging trajectory of the rotating shaft β 'of the rotating piston 30a through the rotating piston rotating shaft β'. Mounting angle is changed by the position of the fixed pin) (30g) to realize the slider mechanism to be inserted. It is apparent that the slider mechanism becomes the piston rotation defining means, even if the rotation defining means constituted by combining this slider mechanism and the synchronous rotating means can be operated as shown in FIG. 10. .

Thus, the mechanism which combined the side surface flat part 30h which implements a rotation synchronization means, the slider mechanism comprised from the positioning cylinder 30p which implements piston rotation defining means, and the guide groove 30g is the said condition (A Satisfies)

Next, what satisfies the condition (B) will be described. However, the prescribed function degree of the rotation defining means is defined before that. The prescribed degree of operation is an index that defines how precisely the target definition can be defined, and is defined as the inverse of the difference between the target value and the actual setting value. The target value of this embodiment is represented by said Formula 1, and the inverse of the error (rotation reaction force) of the target rotation amount and actual rotation amount of a rolling cylinder corresponding to the turning amount (phi) of a turning piston becomes a function. That is, the following equation 4 is obtained.

[Equation 4]

The main cause of the rolling cylinder rotational error (rotational reaction force) is considered to be the gap (reaction force) between the positioning cylinder 30p and the guide groove 30g. Therefore, let this rotation reaction force be (DELTA), and calculate | regulate the formula of a defined action degree. As a parameter of a defined action degree, as shown in FIG. 21, the setting angle of the position fixing cylinder 30p (it is an angle on the locus | trajectory of the turning piston rotating shaft (beta '), and (theta)) of the turning piston 30a, There is a turning amount Φp (with the same definition as in FIG. 17). In FIG. 21, the rotation reaction force (DELTA) is expanded and the diameter of the positioning cylinder 30p is reduced for description. From FIG. 21, the rotation reaction force (DELTA) calculates following Formula 5 from the reaction force (DELTA) s of the circumferential direction in radius L. As shown in FIG.

[Equation 5]

Where abs represents an absolute value.

In this expression 5, L and Δs are the following expressions 6 and 7.

[Equation 6]

[Equation 7]

Here, χ is also an angle formed by the line connecting the cylinder rotation shaft γ and the position fixing cylinder 30p and the guide groove 30g, as shown in FIG.

The guide groove 30g passes the cylinder rotation axis γ when the turning piston rotating shaft β 'comes to the position of the cylinder rotation axis γ, and the rotational speed of the guide groove 30g is synchronized with the rolling cylinder 30b. The rotational speed of the rolling cylinder 30b is half of the rotational speed of the swinging piston 30a, and it can be seen that χ is obtained by the following expression (8).

[Equation 8]

Substituting Equation 8 into Equation 7, and then substituting Equation 5 together with Equation 6 to obtain the rotational reaction force Δ, and finally by substituting Equation 4 to obtain the prescribed degree of action, the following Equation 9 is obtained.

[Equation 9]

Equation 9 calculates the prescribed degree of action at Φ p = π rad where the rotating piston rotating shaft β 'coincides with the cylinder rotation axis γ and φ p = 0 rad rotated by π rad (180 degrees) by Equation 9 (11).

[Equation 10]

[Equation 11]

Except for the case of θ = πred, the following inequality of Equation 12 is established.

[Equation 12]

From this, the prescribed functioning degree of the newly set rotation defining means is that the prescribed functioning degree by the drive system is low except for the case where the position fixing cylinder 30p coincides with the cylinder rotation axis γ (? =? High when the rotating shaft β 'coincides with the cylinder rotation shaft γ, and the specified action by the drive system is high (when the turning piston rotating shaft β' has rotated 180 degrees from the position coinciding with the cylinder rotating shaft γ). It can be seen that the time is lowered. This shows that the condition (B) is satisfied.

As mentioned above, even if the positioning cylinder 30p is set in any position different from the cylinder rotation shaft (gamma), the newly installed piston-cylinder rotation control means cooperates with a drive system, and replenishes when one prescribed action degree is low. On the contrary, when the degree of prescribed action is high, the degree of prescribed action is lowered so as not to be overcommitted, thereby achieving the effect of smooth pump operation.

Further, as is apparent from Equation 9, as the set angle θ is changed from πrad to Orad, it is found that the absolute value of the defined degree of action becomes large (when the turning amount Φ p is the same) and becomes maximum in Orad. Can be. That is, when the position fixing cylinder 30p is installed in the position which rotated 180 degrees on the trajectory of the rotating piston rotating shaft (beta ') from the position of the cylinder rotating shaft ((gamma)), a prescribed action will become the highest and rotational reaction force (DELTA) will be It can suppress and exhibits the effect that a pump operation can be made more smoothly.

The position fixing cylinder 30p is standing up in the pump discharge flow path 30d, as apparent from FIG.7 and FIG.9. The guide groove 30g is a discharge path connecting the pump discharge flow path 30d and the oil supply hole 6b. Therefore, since the sliding parts of the positioning cylinder 30p and the guide groove 30g are always located in the center of the oil passage, the lubricity is improved and the reliability is improved.

In this embodiment, the guide groove 30g is extended to the side flat part 30h. As described above, since the guide groove 30g is an oil flow path, the oil is present in the guide groove 30g smoothly. Accordingly, the guide groove 30g extending to the side flat portion 30h serves as an oil supply flow path to the side flat portion 30h, thereby improving the reliability by improving the lubricity between the side flat portion 30h and the pump groove 30g. In addition, with the improvement of sealing property, the performance shape effect by the leakage reduction from the discharge pump chamber 140d to the suction pump chamber 140s is exhibited. In addition, by extending the guide groove 30g to the side flat portion 30h, the movement of the cutting tool at the time of groove processing becomes constant, and there is an effect that the shape accuracy of the groove is improved.

Moreover, you may provide an affinity film, such as manganese phosphate, on the surface of the turning piston 30a, the rolling cylinder 30b, or the pump casing 30c. In this case, due to the affinity effect, each gap becomes small with operation, and there exists an effect that sealing property improves and pump performance improves.

In addition, in this embodiment, the turning piston 30a was made into the drive side and the rolling cylinder 30b was made into the passive side, On the contrary, the rolling cylinder 30b may be made into the drive side and the turning piston 30a may be made into the passive side. In this case, the revolving piston 30a performs crank support which can perform revolving motion.

In addition, in this embodiment, the pump groove 30b1 extends to the outer periphery of the rolling cylinder 30b. Thereby, since the movement of the cutting tool at the time of groove making becomes constant, there exists an effect that the shape precision of a groove improves. Moreover, since lubrication to a rolling cylinder outer peripheral gap is performed, the sealing property of an outer peripheral gap can be improved and there exists an effect that a pump performance improves. In addition, due to the penetrating pump groove 30b1, an end plate 30b4 is required, whereby the configuration of the fourth embodiment described later, which can suppress side gaps by pressing, can be realized. However, the present invention is not limited thereto, but may be a pump groove type of a conventional long hole shape without penetrating to the outer periphery.

In addition, in this embodiment, although the axial center of the two cylindrical surfaces 30a4 provided in the side surface of the turning piston 30a is shifted, the cylindrical surface which coincided with the axial center like the same cylindrical surface shown by the dashed-dotted line of FIG. You may make it. The reason is that since the working fluid of the present embodiment is an incompressible fluid called oil, the performance deterioration due to the dead volume is small, and the side shape can be processed by cutting the side flat portion 30h after the cylindrical surface is formed. There is an effect that the processing cost is reduced. In addition, in the case of the use of only oil transfer without boosting, the performance decrease due to the dead volume becomes negligible, so that the pivot piston 30a of the same cylindrical surface shown by the dashed-dotted line in FIG. Suitable.

In addition, in this embodiment, although the setting position of the position fixing cylinder 30p is set as the slide mechanism provided in the position which rotated 180 degrees from the position of the cylinder rotation shaft (gamma) on the turning trace of the turning piston rotating shaft (beta '), As described in the description of the prescribed operation degree, the moving position fixing cylinder 30p 'and the moving position fixing cylinder 30p' moved on the turning trajectory of the turning piston rotating shaft β 'to other positions. Even if the movement guide groove 30g 'which rotates the set angle by half of the movement angle is formed, the said slider mechanism used as the said piston rotation defining means can be comprised (refer FIG. 10). For example, as shown in FIG. 8, when the positioning cylinder 30p is set to the position which rotated 45 degrees, the guide groove rotates 22.5 degrees. In this case, as can be seen from the description of the above-described regulation action, the timing at which the regulation action is maximized is the same as before the movement of the position fixing cylinder (Fig. 10 (5)). However, as can be seen from the equation 9, the regulating action is smaller than before the positioning cylinder 30p moves. As a result, the cylinder rotation shaft γ is rotated on the turning trajectory of the rotating piston rotating shaft? What is necessary is just to form the movement position fixing cylinder 30p 'and the movement guide groove | channel 30g', when it cannot install in the position which rotated 180 degrees from the position of.

In the embodiment described so far, the turning piston 30a is close to the turning piston rotating shaft β ′ and the cylinder rotating shaft γ which are close to the rolling cylinder 30b which causes lock. The premise was that the turning radius Ep and the eccentricity Es of the rolling cylinder 30b were approximately equal (when the turning piston rotating shaft β 'passed the cylinder rotating shaft γ). That is, the case where the turning radius Ep and eccentricity Es differ is not considered. This is based on the following reasons.

The turning radius Ep is determined by the eccentricity Ec of the pump eccentric part 6f1 and the bearing clearance eccentricity in the piston bearing hole 30a6, and the eccentricity Es is the rolling cylinder 30b installed in the cylinder casing 30c3. ) Is determined by the amount of eccentricity Eb of the hole in which the holes are arranged and the amount of eccentricity of the bearing gap of the outer circumferential portion of the cylinder casing 30c3. The rotation defining means of this embodiment means for forcibly changing the bearing eccentricity in both so that the trajectory of the swinging piston rotating shaft β 'near the cylinder rotating shaft γ crosses the cylinder rotating shaft γ. In other words, That is, as a result of providing the rotation defining means, the turning radius Ep and the amount of eccentricity Es are equal to each other in the vicinity of the closest proximity of the turning piston rotating shaft β 'and the cylinder rotating shaft γ.

The change in the eccentricity of the bearing by the rotation defining means described in the shear lock becomes larger when the difference between the eccentricity Ec and the eccentricity Eb becomes large, and the load applied to the defining means or the pump eccentric portion 6f1 increases, resulting in a decrease in reliability. do. For this reason, in this embodiment, the clearance gap between the turning piston 30a and the pump groove 30g was made into twice the difference of the eccentricity Ec and the eccentricity Eb. Thereby, even if the turning radius Ep and the eccentric amount Es differ in the vicinity of the closest approach of the rotating piston rotating shaft β 'and the cylinder rotating shaft γ, as shown in FIG. 22, the turning piston 30a Although it shifts from the central axis in the pump groove 30b1, it does not rotate the rolling cylinder 30b abnormally. For this reason, the above-mentioned change of the bearing eccentricity is unnecessary, and the load applied to the defining means and the pump eccentric part 6f1 is not increased, and there exists an effect that reliability can be ensured.

When the clearance between the swing piston 30a and the pump groove 30g is more than twice the difference between the eccentricity Ec and the eccentricity Eb, the gap between the swing piston 30a and the pump groove 30b1 is enlarged. Therefore, sealing property falls and performance falls. For this reason, it turns out that what is necessary is just to make the clearance gap between the turning piston 30a and the pump groove 30g less than twice the difference between the eccentric amount Ec and the eccentric amount Eb.

According to the present embodiment, a pump element having a high sealing property can be configured with a simple element shape while reducing the processing cost, so that the risk of operating lock in a high performance rolling cylinder-type pump at low cost is 1 per cylinder. The configuration of the piston makes it avoidable. As a result, a volumetric pump suitable for a small capacity can be realized. In addition, compared to the plural cylinders as in the prior art, the configuration is simple and the assembly performance is improved, so that a volumetric pump with reduced processing cost can be realized.

(2nd embodiment)

Next, with respect to the scroll compressor 1 which mounts the volumetric pump which is 2nd Embodiment of this invention as the oil supply pump 30, the cross-sectional view in FIG. 12 and its LL which are enlarged longitudinal cross-sectional views of the positioning cylinder 30p are shown. It demonstrates using FIG. Since this 2nd Embodiment differs from 1st Embodiment in the following point, about another point, it is basically the same as 1st Embodiment, and the overlapping description is abbreviate | omitted.

In this 2nd Embodiment, the position fixing cylinder 30p is comprised by the center pin 30P3 fixed to the base plate 30c1 by press injection, adhesion | attachment, or electrodeposition, and the cylindrical roller 30p4. The flange part 30p31 is provided in the upper end part of the center pin 30p3. The roller 30p4 is fitted to the main body protrusion part 30p32 of the center pin 30P3, and is prevented from coming out from the center pin 30p3 by the flange part 30p31. In FIG. 12, FIG. 13, although the diameter gap of the roller 30p4 and the center pin 30p3 is shown large, it is actually 100 micrometers or less. For this reason, since the roller 30p4 rotates so that the relative speed of the location where the roller 30p4 and the guide groove 30g strongly slides may become small, it slides between the positioning cylinder 30p and the guide groove 30g. The state becomes good, poor movements such as rattling are suppressed, and the pump operation of the oil supply pump 30 is made smoother. In addition, wear of the positioning cylinder 30p and the guide groove 30g, falling off of the positioning cylinder 30p due to torsion, etc. can be avoided, and the effect of improving the reliability of the oil supply pump 30 is exhibited. do.

(Third embodiment)

Next, the scroll compressor 1 which mounts the volumetric pump which is 3rd Embodiment of this invention as the oil supply pump 30 is demonstrated using FIG. 14 which is the L-L cross-sectional correspondence of FIG.

In this 3rd Embodiment, since it is the same as that of 2nd Embodiment except having set it as the slider roller 30p5 which provided the roller flat surface 30p51 on the outer peripheral surface, description other than that is abbreviate | omitted. According to this third embodiment, the area of the sliding portion with the guide groove 30g is increased, so that the risk of abrasion is lowered, thereby providing an effect that a highly reliable oil supply pump 30 can be provided.

(4th embodiment)

Next, the scroll compressor 1 which mounts the volumetric pump which is 4th Embodiment of this invention as the oil supply pump 30 is demonstrated using FIG. 15, FIG.

This 4th Embodiment is a type which receives the axial position of the crankshaft 6 on the side of the oil supply pump 30 instead of the frame 4, except the vicinity of the back pressure chamber 110 and the oil supply pump 30, It is the same as that of the first to third embodiments. For this reason, it demonstrates mainly using FIG. 15 which is a detailed enlarged view of the back pressure chamber vicinity (corresponding to N part of FIG. 1), and FIG. 16 which is an enlarged longitudinal cross-sectional view of a lubrication pump (corresponding to M part of FIG. 1). Will be omitted.

In this fourth embodiment, the crankshaft flange portion 6h is spaced apart from the frame 4 (see FIG. 15), and the cover of the upper portion of the oil supply pump 30 in the first to third embodiments (see FIG. 5 ( 30c2)] (see FIG. 16), the shaft lower end surface 6z which is the end of the crankshaft 6 is supported by the end plate 30b4 of the rolling cylinder 30b. In the crankshaft 6, the entire area of the upper part faces the back pressure chamber 110, while the lower part of the crank shaft 6 faces the suction pressure which is the internal space pressure of the casing 8, so that a downward pressing force always acts. Therefore, the crankshaft 6 presses down the end plate 30b4. Therefore, the gap between the upper surface of the swing piston 30a, which is the side gap of the oil supply pump 30, and the bottom surface of the pump groove 30b1, the gap between the flat end portion 30a5 (lower surface of the swing piston) and the upper surface of the base plate 30c1, The gap between the rolling cylinder 30b and the upper surface of the base plate 30c1 can be reduced. Among these, the smallest gap becomes approximately zero, and slides. Which gap is slid is determined by the magnitude relationship between the depth of the pump groove 30b1 and the thickness of the swing piston 30a.

When the thickness of the turning piston 30a is made smaller than the depth of the pump groove 30b1, the clearance gap between the rolling cylinder 30b and the upper surface of the base plate 30c1 will slide. In this case, since the movement of the swinging piston 30a which rotates smoothly becomes smooth, there exists an effect that the pump operation of the oil supply pump 30 whole becomes smooth further. On the contrary, when the thickness of the swing piston 30a is made larger than the depth of the pump groove 30b1, the clearance between the upper surface of the swing piston 30a and the bottom surface of the pump groove 30b1 and the planar end 30a5 (lower surface of the swing piston) ) And the upper surface of the base plate 30c1 become the sliding surface. In this case, since two clearance gaps can be suppressed to a minimum value, sealing property can be improved further and the effect of providing a high performance oil supply pump is exhibited.

16, the oil supply pump 30 is fixed to the lower surface of the ball 25a which comprises the sub bearing 25 by the pump fixing screw 30n. Here, the lower surface of the ball 25a is perpendicular to the inner circumferential surface that is the bearing surface. The lower end face 6z of the shaft is assumed to be perpendicular to the central axis of the crankshaft 6. Thereby, since the shaft lower end surface 6z contacts the end plate 30b4 in the whole surface irrespective of the mounting posture of the crankshaft 6, the sealing property between the shaft lower end surface 6z and the end plate 30b4 is carried out. This improves and provides the effect of providing a high performance oil supply pump. In addition, since one-sided contact between the shaft lower end surface 6z and the end plate 30b4 is suppressed, the effect that the oil supply pump with high reliability can be provided is exhibited.

1: scroll compressor
2: fixed scroll member
3: turning scroll member
4: frame
5: Oldham Ring
6: crankshaft
6b: oil supply hole
6f: Pump shaft part
6f1: pump eccentric
7: motor
22: bypass valve
26: back pressure control valve
30: oil supply pump
30a: slewing piston
30a5: flat end
30a6: Piston Bearing Bore
30b: rolling cylinder
30b1: pump groove
30b4: end plate
30c: pump casing
30c4: casing room
30d: pump discharge flow path
30g: guide groove
30h: side flat part
30p: fixed position cylinder
30p4: Roller
30p5: slider roller
30s: pump suction flow path
100: compression chamber
105: suction chamber
110: back pressure chamber
120: discharge chamber
125: oil reservoir
140: pump chamber
140s: suction pump room
140d: discharge pump chamber
α: piston pivot shaft (crankshaft rotation shaft)
β: central axis of pump eccentric
β ': rotating piston rotating shaft
γ: cylinder rotation axis (cylinder rotation axis)

Claims (13)

A pump casing having a casing chamber,
A rolling cylinder having a pump groove and disposed rotatably in the casing chamber;
A turning piston loosely fitted in the pump groove to partition the pump groove to form two pump chambers;
The pivoting piston or the rolling cylinder is pivoted or rotated by a drive source,
The rolling cylinder is disposed so as to be rotatable about the cylinder rotation axis whose eccentricity is Es with respect to the piston pivot axis, which is the central axis of the pivot motion of the pivot piston,
The pump groove extends in a straight line cross the cylinder axis of rotation which is the central axis of the rotational movement of the rolling cylinder,
The pivoting piston is arranged to be capable of rotating in a reciprocating manner within the pump groove and at the same time pivoting the pivot radius with respect to the pump casing with respect to the pump casing,
The turning radius Ep of the turning piston and the amount of eccentricity Es of the rolling cylinder are approximately equal in volume pumps.
Let the number of the said turning pistons fitted in the rolling cylinder be one,
A volumetric pump, characterized in that a rotation defining means is provided for maintaining the turning of the turning piston and the rotation of the rolling cylinder and defining the turning speed of the turning piston at twice the rotation speed of the rolling cylinder. 2. The method of claim 1, wherein the prescribed functioning degree of the rotation defining means is smaller than the case where the piston pivot shaft is rotated 180 degrees, rather than when the piston pivot shaft coincides with the cylinder rotation shaft. A displacement pump. The crankshaft of Claim 1 or 2 WHEREIN: The turning motion of the said turning piston is implement | achieved with the crankshaft whose eccentricity is Ec,
The rotating shaft of the rolling cylinder is realized by a bearing portion having an eccentric amount of Eb,
The clearance pump of the said turning piston and the said pump groove was made into 2 times or less of the difference of the eccentricity Ec of the said crankshaft, and the eccentricity Eb of the said bearing part. The said rotation defining means is rotation rotation means which synchronizes rotation of the said turning piston and rotation of the said rolling cylinder, and piston rotation which prescribes the turning speed of the said turning piston at twice the rotation speed. A volumetric pump, comprising provision means. 5. The volumetric pump according to claim 4, wherein the rotation synchronizing means is provided with the side piston having sliding sides in contact with two side surfaces of the pump groove, respectively, in the turning piston. 5. The piston rotation defining means according to claim 4, wherein the piston rotation defining means is always on a straight line which is fixed on the pivoting piston passing through the pivoting piston rotational axis at a stop point at a position different from the cylinder rotational axis on the pivotal motion trajectory of the pivoting piston rotational shaft. A volumetric pump comprising a slider mechanism for defining the movement of the swing piston. The volumetric pump according to claim 6, wherein the stop point is provided at a position rotated 180 degrees from the position of the cylinder rotation shaft on the swing movement trajectory. 7. The slider mechanism of claim 6, wherein the slider mechanism comprises a floating slider and a rotation guide,
The floating slider is realized by a position fixing cylinder fixedly arranged at a position corresponding to the stop point of the pump casing,
The rotary guide is realized by forming a guide groove having a width equal to the diameter of the positioning cylinder in order to form the positioning cylinder and the sliding treatment on the pivoting piston with the straight line as a center line. . The volumetric pump according to claim 8, wherein the floating slider has a fixed central axis and a slider member rotatably inserted in the fixed central axis. The volumetric pump according to claim 8, wherein the guide groove is a passageway for a working fluid. The volumetric pump according to claim 10, wherein the guide groove is extended to the side flat portion. The volumetric pump according to claim 1 or 2, characterized in that it is mounted in said volumetric fluid machine as an oil supply source to each part of the volumetric fluid machine. A drive source, a crankshaft driven by the drive source, and a volumetric pump driven by the crankshaft,
The displacement pump includes a pump casing having a casing chamber, a rolling cylinder having a pump groove disposed in the casing chamber so as to be rotatable, and two pump chambers which are loosely fitted in the pump groove to partition the pump groove. With a turning piston to
The pivoting piston or the rolling cylinder is pivoted or rotated by the drive source,
The rolling cylinder is disposed so as to be rotatable about the cylinder rotation axis whose eccentricity is Es with respect to the piston pivot axis, which is the central axis of the pivot motion of the pivot piston,
The pump groove extends in a straight line cross the cylinder axis of rotation which is the central axis of the rotational movement of the rolling cylinder,
The pivoting piston is arranged to be capable of rotating in a reciprocating manner within the pump groove and at the same time pivoting the pivot radius with respect to the pump casing with respect to the pump casing,
The turning radius Ep of the turning piston and the eccentricity Es of the rolling cylinder are approximately equal in volumetric fluid machines,
Let the number of the said turning pistons fitted in the rolling cylinder be one,
A volumetric fluid machine comprising: a rotation defining means for maintaining the swinging of the swinging piston and the rotation of the rolling cylinder, and defining the swinging speed of the swinging piston at twice the rotational speed of the rolling cylinder.

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