JP2001336491A - Rotary fluid machine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To enable to efficiently lubricate each slide part of a rotary fluid machine. SOLUTION: Both end peripheries of an output shaft 23 integrated with a rotor 31 of a vane type expanding machine powered by high pressure steam are mounted in a floating state with a liquid membrane of a pressurized liquid phase fluid supplied from a supply port of the pressurized liquid phase fluid 129 through a passage of the pressurized liquid phase fluid W5, and mounted in the floating state with a liquid membrane of a pressurized liquid phase fluid supplied from the supply port of the pressurized liquid phase fluid 129 through supply ports of the pressurized liquid phase fluid W6, W7, W9, W10, W11 and W12. A reciprocating vane 42 radially mounted with a rotor 31 is mounted in the floating state with a liquid membrane of a pressurized liquid phase fluid supplied through a passage of the pressurized liquid phase fluid extending radially outward in inside the rotor 31 (not shown in figure).

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a rotary fluid machine for converting pressure energy of a high-pressure gas-phase working medium into mechanical energy and extracting it from an output shaft.

[0002]

2. Description of the Related Art JP-A-58-48706 discloses that a high-pressure gas-phase working medium generated by heating a liquid-phase working medium by an evaporator is expanded by an expander to extract mechanical energy.
A Rankine cycle device is described in which the resulting low-pressure gas-phase working medium is cooled by a condenser and returned to a liquid-phase working medium, and the liquid-phase working medium is again supplied to an evaporator by a pump. The above-mentioned conventional one uses a vane type expander as a rotary fluid machine constituting the expander.

[0003]

By the way, in the vane type expander, it is necessary to lubricate the bearing portion of the output shaft which rotates integrally with the rotor and the sliding portion of the vane supported by the rotor so as to be movable in the radial direction. However, since these sliding parts are exposed to severe conditions of high temperature and high pressure, seizure and wear may occur unless effective lubricating means is employed.

[0004] The present invention has been made in view of the above circumstances, and has as its object to efficiently lubricate each sliding portion of a rotary fluid machine.

[0005]

According to the first aspect of the present invention, there is provided a casing having a rotor chamber, a rotor housed in the rotor chamber, A plurality of vane piston units arranged radially around the rotation axis and reciprocally movable in a radial direction, wherein each vane piston unit includes a vane sliding in the rotor chamber, and a non-sliding movement of the vane. A low-pressure gas phase caused by the expansion of the high-pressure gas-phase working medium to operate the piston to rotate the rotor through a power conversion device and to cause a pressure drop of the high-pressure gas-phase working medium. In a rotary fluid machine that rotates the rotor via the vanes by expansion of a working medium, a hydrostatic bearing of an output shaft that rotates integrally with the rotor A pressurized liquid-phase fluid is supplied to the output shaft to support the output shaft under a static pressure, and a part of the pressurized liquid-phase fluid is located radially outside the rotor via a passage formed inside the rotor. The present invention proposes a rotary fluid machine characterized by supplying to a hydrostatic bearing.

According to the above configuration, the pressurized liquid-phase fluid is supplied to the hydrostatic bearing of the output shaft that rotates integrally with the rotor and another hydrostatic bearing located radially outside the rotor, so as to float. Because of the static pressure support, it is possible to avoid the occurrence of solid contact and to reliably prevent the occurrence of seizure or wear in the sliding portion. Further, since the pressurized liquid-phase fluid is supplied to the other hydrostatic bearing via a passage formed inside the rotor, the pressurized liquid-phase fluid is further pressurized by the centrifugal force accompanying the rotation of the rotor, so that an effective static liquid is supplied. Pressure support can be provided.

According to the second aspect of the present invention,
In addition to the configuration of the first aspect, a rotary fluid machine is proposed, wherein the another static pressure bearing is configured to statically support the vane in a slot-shaped space formed in the rotor.

According to the above construction, the pressurized liquid-phase fluid is further pressurized by centrifugal force accompanying the rotation of the rotor, so that the vane is effectively statically supported and floated in the slot-like space formed in the rotor, Seizure and wear can be prevented.

According to the third aspect of the present invention,
In addition to the configuration of claim 1, a rotary fluid machine is proposed in which the another hydrostatic bearing supports the side surface of the rotor with hydrostatic pressure on the inner surface of the casing.

According to the above construction, the pressurized liquid-phase fluid is further pressurized by the centrifugal force accompanying the rotation of the rotor, whereby the side surface of the rotor is effectively statically supported on the inner surface of the casing to float, and the seizure is prevented. Wear can be prevented.

According to the invention described in claim 4,
In addition to the configuration according to any one of claims 1 to 3, a rotary fluid machine is proposed in which the pressurized liquid-phase fluid is the same fluid as the gas-phase working medium.

According to the above configuration, since the pressurized liquid-phase fluid and the gas-phase working medium are the same fluid, not only a special pressurized liquid-phase fluid for the hydrostatic bearing is not required, but also the working medium is required. It is possible to avoid adverse effects caused by mixing different kinds of pressurized liquid-phase fluids.

The water and the pressurized liquid-phase fluid in the embodiments correspond to the liquid-phase working medium of the present invention, and the steam in the embodiments corresponds to the gas-phase working medium of the present invention. The fluid passage W14 corresponds to the passage of the present invention.

[0014]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Hereinafter, embodiments of the present invention will be described based on embodiments of the present invention shown in the accompanying drawings.

1 to 19 show an embodiment of the present invention. FIG. 1 is a schematic view of a waste heat recovery device for an internal combustion engine, and FIG. 2 is an expansion corresponding to a cross-sectional view taken along line 2-2 of FIG. 3 is a longitudinal sectional view of the expander corresponding to the sectional view taken along line 3-3 of FIG. 5, FIG. 4 is an enlarged sectional view around the rotation axis of FIG. 2, and FIG. 5 is a sectional view taken along line 5-6, FIG. 6 is a view taken along line 6-6 of FIG. 2, FIG. 7 is an explanatory view showing a sectional shape of the rotor chamber and the rotor, FIG. 8 is a view taken along line 8-8 of FIG. 8 is a sectional view taken along line 9-10 of FIG. 8, FIG. 10 is a sectional view taken along line 10-10 of FIG. 8, FIG. 11 is an enlarged view of a main part of FIG. 5, FIG. 13 is a view in the direction of arrow 13 in FIG. 12, and FIG.
14 is a sectional view taken along line 14-14, FIG. 15 is a sectional view taken along line 15-15 in FIG. 14, FIG. 16 is a sectional view taken along line 16-16 in FIG.
5 is an enlarged view around the rotation axis of FIG. 5, FIG. 18 is a graph showing the relationship between the phase of the rotor and the pressure PF of the slot-like space and the pressure PQ of the rotor chamber, and FIG. 19 is a diagram showing the relationship between the phase of the rotor and the load applied to the vane. It is a graph shown.

In FIG. 1, a waste heat recovery device 2 of an internal combustion engine 1 uses a waste heat of the internal combustion engine 1, for example, exhaust gas as a heat source, and operates in a high-pressure state in which the temperature is raised from a high-pressure working medium, for example, water. Evaporator 3 for generating steam, that is, high-temperature and high-pressure steam, expander 4 for generating an output by expansion of the high-temperature and high-pressure steam, and the temperature and pressure after expansion, which are discharged from the expander 4, have decreased. It has a condenser 5 for liquefying the steam, that is, the stepped-down steam, and a supply pump 6 for pressurizing and supplying the water from the condenser 5 to the evaporator 3 and a static pressure bearing described later.

The expander 4 has a special structure and is configured as follows.

2 to 6, the casing 7 is composed of first and second metal halves 8 and 9. Both halves 8, 9
Consists of a main body 11 having a substantially elliptical concave portion 10 and a circular flange 12 integral with the main body 11. A substantially elliptical rotor chamber 14 is formed by overlapping the two circular flanges 12 via a metal gasket 13. Is done.
The outer surface of the main body 11 of the first half 8 is covered by a deep bowl-shaped main body 16 of a shell-shaped member 15, and a circular flange 17 integral with the main body 16 is attached to the circular flange 12 of the first half 8 by a gasket. The three circular flanges 12 and 17 are superimposed via 18 and fastened by bolts 19 at a plurality of positions in the circumferential direction. Thereby, both the main bodies 11 of the shell-shaped member 15 and the first half body 8,
Between 16, a relay chamber 20 is formed.

The main body 11 of the two halves 8, 9 has hollow bearing cylinders 21, 22 projecting outward on their outer surfaces, and the hollow bearing cylinders 21, 22 have hollow hollows penetrating through the rotor chamber 14. A large diameter portion 24 of the output shaft 23 is rotatably supported via a static pressure bearing 25. Thereby, the axis L of the output shaft 23 passes through the intersection of the major axis and the minor axis in the rotor chamber 14 having a substantially elliptical shape. Also, the small diameter portion 26 of the output shaft 23
Is projected from a hole 27 in the hollow bearing cylinder 22 of the second half 9 to the outside and connected to a transmission shaft 28 via a spline connection 29. The space between the small diameter portion 26 and the hole portion 27 is sealed by two seal rings 30.

A circular rotor 31 is provided in the rotor chamber 14.
Is accommodated, and the central shaft mounting hole 32 and the large-diameter portion 24 of the output shaft 23 are in a fitting relationship, and a meshing coupling portion 33 is provided between the two 31 and 24. As a result, the rotation axis of the rotor 31 matches the axis L of the output shaft 23, so that "L" is shared as the sign of the rotation axis.

A plurality of rotors 31 extending radially from a shaft mounting hole 32 about a rotation axis L thereof,
Two slot-shaped spaces 34 are formed at equal intervals on the circumference. Each space 34 has a narrow circumferential width, and the rotor 3
A substantially U-shape is formed in an imaginary plane orthogonal to both end surfaces 35 so as to open in series at both end surfaces 35 and the outer peripheral surface 36 of the first.

As best shown in FIGS. 11 and 16, in each slot-like space 34 (see FIG. 4), first to twelfth vane piston units U1 to U12 having the same structure are provided.
Are mounted so as to be reciprocally movable in the radial direction as follows. In the substantially U-shaped slot-like space 34, a stepped hole 38 is formed in a portion 37 defining the inner peripheral side thereof.
A stepped cylinder member 39 made of ceramic (or carbon) is fitted therein. The end face of the small diameter portion a formed inside the cylinder member 39 in the radial direction abuts on the outer peripheral surface of the large diameter portion 24 of the output shaft 23, and the small diameter hole b penetrating the inside of the small diameter portion a in the radial direction is formed by the large diameter portion 24. Through hole c opening on the outer peripheral surface
(See FIG. 4). Outside the cylinder member 39, a pair of side plates 40 formed coaxially with the member 39 and arranged in plane symmetry are arranged. The pair of side plates 40 facing each other is the cylinder member 3.
9, a pair of side plates 4
The axis of 0 coincides with the axis of the cylinder member 39. The ceramic piston 41 is guided by the cylinder member 39 and a pair of side plates 40 opposed to each other so as to be movable in the radial direction, and the first to twelfth vane piston units U1 to U12 are connected to a pair of opposed side plates. It is guided in the slot-like space 34 formed between 40 so as to be movable in the radial direction.

As shown in FIG. 2 and FIG.
The cross section B of the rotor chamber 14 in the imaginary plane A including the rotation axis L is a pair of semicircular cross sections B1 having diameters g facing each other, and one of the two diameters g of the two semicircular cross sections B1 facing each other. It has a square cross section B2 formed by connecting the ends and the other opposing ends to each other, and has a substantially track shape for competition. In FIG. 7, the portion indicated by the solid line indicates the maximum cross section including the major axis, while the portion partially indicated by the two-dot chain line indicates the minimum cross section including the minor axis. The rotor 31 has a cross section D slightly smaller than the minimum cross section including the minor diameter of the rotor chamber 14, as shown by the dotted line in FIG.

As shown in FIGS. 2 and 8 to 10, the vane 42 has a substantially U-shaped (horshoe-shaped) vane body 43 and a substantially U-shaped plate attached to the vane body 43. It is composed of a seal member 44 (see FIG. 2).

The vane main body 43 is a plate-like member having a constant thickness. The vane main body 43 has a semicircular arc-shaped portion 46 corresponding to the inner peripheral surface 45 formed by the semicircular cross section B1 of the rotor chamber 14, and an opposing inner space formed by the square cross section B2. It has a pair of parallel portions 48 corresponding to the end surfaces 47. The vane body 4 extends from one opposing inner end surface 47 to the other opposing inner end surface 47 via the semicircular portion 46.
3, a U-shaped groove 52 opening outwardly, for example, PT
The seal member 44 made of FE is fitted and held. A pair of short shafts 51 are provided on the end side of each parallel portion 48 and project left and right outward.
Are respectively mounted with rollers 59 having a ball bearing structure. On the front and back sides of the vane main body 43, a square having a predetermined area for holding a pressurized liquid-phase fluid (pressurized water in the embodiment) supplied to the hydrostatic bearing is provided adjacent to each short axis 51 (implemented). In this example, a total of four shallow recesses 49 having a rectangular shape but an arbitrary shape) are formed.

The axis L1 in FIG.
A pair of plate-like projections 53 are provided on the axis L1 of the piston 9 and the piston 41, and the radial outer end of the piston 41 abuts on the axis L1. A blind pressurized liquid phase fluid passage W1 is formed extending between the pair of plate-shaped protrusions 53 along the axis L1 into the inside of the vane body 43, and the pressurized liquid phase bent at a right angle from the tip thereof. The fluid passage W2 opens on one side surface of the vane main body 43.

Next, the structure of the side plate 40 will be described with reference to FIGS.

The side plate 40 has a U-shaped outer shape similar to the vane body 43, and has an inner member 122 having a vane sliding surface 121 with which the vane body 43 slides. An outer member 123 is held, and an orifice forming member 124 supported between the two members 122 and 123 and protruding from the outer surface of the outer member 123 on the outer side in the radial direction of the rotor 31. Also, a side plate 4 corresponding to a radially inner side of the rotor 31.
No. 0 has an inclined surface 125 that is tapered with respect to the vane sliding surface 121, and the inclined surface 125 facilitates mounting of the rotor 31 in a narrow space radially inward. The inner member 122 of the side plate 40 has a vane sliding surface 121.
The piston 41 moved radially outward from the cylinder member 39 is accommodated in the piston guides 126 of the pair of side plates 40 in a non-contact manner (FIG. 11). And FIG. 16).

A total of eight pressurized liquid-phase fluid passages W3 and W4 extending radially from the orifice forming member 124 are formed on the mating surface of the inner member 122 and the outer member 123, and six of them are engraved. The leading ends of the pressurized liquid-phase fluid passages W3 are opened as the pressurized liquid-phase fluid discharge holes 127 on the vane sliding surface 121 of the inner member 122, and the remaining two pressurized liquid-phase fluid passages W4 are connected to the inner member 122. Is opened as the pressurized liquid-phase fluid discharge hole 128 (see FIG. 12). The orifice forming member 124 is provided with the eight pressurized liquid phase fluid passages W3 and W3.
An orifice function is exhibited for W4.

Each vane 42 is slidably accommodated in each slot-like space 34 of the rotor 31. At this time, both side surfaces of the vane main body 43 have a pair of side plates 4 facing each other.
0 and slides in the radial direction between the vane sliding surfaces 121. At this time, the inner end surfaces of the pair of protrusions 53 of the vane 42 can contact the outer end surface of the piston 41. Both rollers 59 provided on the vane 42 are rotatably engaged with the substantially elliptical annular grooves 60 formed on the opposed inner end faces 47 of the first and second halves 8 and 9, respectively. The distance between the annular groove 60 and the rotor chamber 14 is constant over their entire circumference. Further, the forward movement of the piston 41 is converted into the rotational movement of the rotor 31 by the engagement between the roller 59 and the annular groove 60 via the vane 42.

By the cooperation of the roller 59 and the annular groove 60,
As is clearly shown in FIG.
6, the semicircular tip surface 61 is formed from the inner peripheral surface 45 of the rotor chamber 14, and both parallel portions 48 are formed in the rotor chamber 1.
4 are always separated from the opposing inner end surfaces 47, thereby reducing friction loss. And 2
Since the trajectory is regulated by the annular groove 60 composed of a pair of strips, the vane 42 is caused to rotate by a small displacement angle in the axial direction via the roller 59 due to a right and left trajectory error, and the inner circumferential surface 45 of the rotor chamber 14 is rotated. Increase the contact pressure with At this time, in the vane body 43 having a substantially U-shaped plate shape (horse-shoe shape), the displacement in the radial direction of the contact portion with the casing 7 can be greatly reduced as compared with the rectangular (rectangular) vane. Further, as clearly shown in FIG.
The sealing member 44 mounted on the inner surface of the rotor chamber 14 tightly contacts the inner peripheral surface of the rotor chamber 14 to perform a sealing action. At this time, since the substantially U-shaped plate-shaped vane 42 has no inflection point with respect to the square (rectangular) vane, the adhesion is good.

The sealing action between the vane body 43 and the inner peripheral surface of the rotor chamber 14 is effected by the spring force of the sealing member 44 made of an elastic material, the centrifugal force acting on the sealing member 44 itself, and the high pressure side. The steam which has entered the U-shaped groove 52 of the vane main body 43 from the rotor chamber 14 of the first embodiment generates steam pressure which pushes up the seal member 44. As described above, since the sealing action is not affected by the excessive centrifugal force acting on the vane main body 43 according to the rotation speed of the rotor 31, the sealing surface pressure does not depend on the centrifugal force applied to the vane main body 43, It is possible to always achieve both good sealing performance and low friction performance.

In FIGS. 2 and 4, the large diameter portion 24 of the output shaft 23 has a thick portion 62 supported by the hydrostatic bearing 25 of the second half 9, and a first half extending from the thick portion 62. Body 8
And a thin portion 63 supported by the static pressure bearing 25. A hollow shaft 64 made of ceramic (or metal) is fitted into the thin portion 63 so as to rotate integrally with the output shaft 23. A fixed shaft 65 is disposed inside the hollow shaft 64, and the fixed shaft 65 includes a large-diameter solid portion 66 fitted to the hollow shaft 64 so as to fit within the axial thickness of the rotor 31, and an output shaft. A small-diameter solid portion 69 fitted through two seal rings 68 into a hole 67 existing in the thick portion 62 of the 23, and a thin-walled portion extending from the large-diameter solid portion 66 and fitted into the hollow shaft 64. And a hollow portion 70. A seal ring 71 is interposed between the outer peripheral surface of the end of the hollow portion 70 and the inner peripheral surface of the hollow bearing cylinder 21 of the first half 8.

In the main body 16 of the shell-shaped member 15, a hollow cylindrical body 7 coaxial with the output shaft 23 is provided on the inner surface of the central portion thereof.
The second end wall 73 is attached with a plurality of bolts 50 via a seal ring 74. The inner end side of the short outer cylindrical portion 75 extending inward from the outer peripheral portion of the end wall 73 is connected to the hollow bearing cylinder 21 of the first half 8 via a connecting cylinder 76. End wall 73
And a small and long inner tube 77
The inner end of the inner tube portion 77 is fitted in a stepped hole h in the large-diameter solid portion 66 of the fixed shaft 65 together with a short hollow connecting tube 78 protruding therefrom. The outer end portion of the inner tube 77 projects outward from the hole 79 of the shell-shaped member 15,
The inner end side of the first high-temperature and high-pressure steam introduction pipe 80 inserted into the inner pipe portion 77 from the outer end portion is fitted into the hollow connection pipe 78. A cap member 81 is provided on the outer end portion of the inner pipe portion 77.
Is screwed in, and the introduction pipe 8 is
The flange 83 of the holder cylinder 82 holding the
Is press-bonded to the outer end surface of the first through a seal ring 84.

As shown in FIGS. 2 to 4 and FIG. 11, a plurality of cylinder members 39 of the first to twelfth vane piston units U1 to U12 are formed on the hollow shaft 64 and the output shaft 23 in series. In the example, the rotary valve V that supplies high-temperature and high-pressure steam through the twelve through holes c and discharges the expanded first temperature-lowering pressure-lowering steam through the hole c from the cylinder member 39 is provided with the fixed shaft 65. The large-diameter solid portion 66 is provided as follows.

FIG. 17 shows each cylinder member 39 of the expander 4.
2 shows a structure of a rotary valve V for supplying and discharging steam at a predetermined timing. In the large-diameter solid portion 66, first and second holes 86 and 87 extending in opposite directions from a space 85 communicating with the hollow connection pipe 78 are formed. The first and second holes 86 and 87 are formed. The first and second concave portions 88 and 89 open on the outer peripheral surface of the large-diameter solid portion 66 and open on the bottom surfaces. 1st, 2nd
Carbon first and second seal blocks 92 and 93 having supply ports 90 and 91 are attached to the recesses 88 and 89, respectively.
Their outer peripheral surfaces rub against the inner peripheral surface of the hollow shaft 64. First,
In the second holes 86 and 87, short first and second coaxial holes are coaxial.
The supply pipes 94 and 95 are loosely inserted, and the first and second supply pipes 94 and 95 are inserted.
95, the first and second seal cylinders 9 fitted to the outer peripheral surface on the distal end side
The outer peripheral surfaces i, j of 6,97 are inside the supply ports 90,91 of the first and second seal blocks 92,93 and are fitted to the inner peripheral surfaces of the tapered holes k, m connected thereto. The large-diameter solid portion 66 includes first and second annular concave portions n and o surrounding the first and second supply pipes 94 and 95, and first and second annular concave portions n and o adjacent thereto.
The second blind hole-shaped concave portions p and q are the first and second seal blocks 9.
The first and second bellows-like elastic bodies are formed so as to face the first and second annular concave portions n and o, and one end sides of the first and second annular concave portions n and o are fitted to the outer peripheral surfaces of the first and second seal cylinders 96 and 97. 98,99
However, the first and second coil springs 100 and 101 are respectively housed in the first and second blind hole-shaped concave portions p and q, and the first and second bellows-like elastic bodies 98 and 99 and the first and second coil springs are provided.
First and second by the spring force of the coil springs 100 and 101
The seal blocks 92 and 93 are pressed against the inner peripheral surface of the hollow shaft 64.

In the large-diameter solid portion 66, two through holes c are always provided between the first coil spring 100 and the second bellows-like elastic body 99 and between the second coil spring 101 and the first bellows-like elastic body 98. The first and second concave discharge portions 102 and 103 communicating with each other, and the discharge portions 102 and 103
First and second discharge holes 104 and 105 are formed from 103 to extend in parallel with the introduction pipe 80 and open into the hollow portion r of the fixed shaft 65.

The first seal block 92 and the second seal block 93 are members of the same kind,
The one with the “first” character and the one with the “second” character are point-symmetric with respect to the axis of the fixed shaft 65.

The inside of the hollow portion r of the fixed shaft 65 and the hollow cylindrical body 7
The inside of the second outer cylinder portion 75 is a passage s for the first temperature-reduced and reduced-pressure steam, and the passage s communicates with the relay chamber 20 via a plurality of through holes t penetrating the peripheral wall of the outer cylinder portion 75.

As shown in FIGS. 2 and 6, the first half 8
In the outer peripheral portion of the main body 11, first and second introduction hole groups 107 and 108 including a plurality of introduction holes 106 arranged in the radial direction are formed near both ends of the minor diameter of the rotor chamber 14.
From the relay chamber 20, the first reduced-temperature and reduced-pressure steam is introduced into the rotor chamber 14 through the introduction hole groups 107 and 108. In the outer peripheral portion of the main body 11 of the second half 9, one end of the long diameter of the rotor chamber 14 and the second introduction hole group 108
Between the plurality of outlet holes 1 arranged in the radial and circumferential directions.
09 is formed, and between the other end of the major axis and the first introduction hole group 107, a second extraction hole group is formed of a plurality of radially and circumferentially arranged discharge holes 109. 111 are formed. These first and second lead-out hole groups 11
From 0 and 111, the second temperature-reduced pressure-reduced steam having a further reduced temperature and pressure is discharged to the outside due to expansion between the adjacent vanes 42.

In FIG. 5, the first and seventh vane piston units U1 and U7, which are point-symmetric with respect to the rotation axis L of the rotor 31, perform the same operation. This is because the second and eighth vane piston units U having a point symmetrical relationship
The same applies to 2, U8 and the like.

For example, referring to FIG.
4 is shorter than the minor axis position E of the rotor chamber 14 in FIG.
, The first vane piston unit U1 is located at the short-diameter position E, and no high-temperature high-pressure steam is supplied to the large-diameter cylinder hole f. And the vane 42 is in the retracted position.

When the rotor 31 is slightly rotated counterclockwise in FIG. 5 from this state, the supply port 90 of the first seal block 92 and the through hole c communicate with each other, and the high-temperature high-pressure steam from the introduction pipe 80 is reduced in diameter. It is introduced into the large-diameter cylinder hole f through the hole b. As a result, the piston 41 moves forward, and the forward movement is caused by the sliding movement of the vane 42 toward the long diameter position F of the rotor chamber 14, whereby the engagement of the roller 59 and the annular groove 60 integral with the vane 42 via the vane 42. Rotor 3 by combination
It is converted into one rotational motion. When the through-hole c is displaced from the supply port 90, the high-temperature and high-pressure steam expands in the large-diameter cylinder hole f to cause the piston 41 to still move forward, so that the rotor 31
Rotation continues. The expansion of the high-temperature and high-pressure steam ends when the first vane piston unit U1 reaches the long diameter position F of the rotor chamber 14. Thereafter, as the rotor 31 rotates, the first temperature-reduced pressure-reduced steam in the large-diameter cylinder hole f is reduced by the vane 42 causing the piston 41 to retreat, thereby causing the small-diameter hole b, the through hole c, and the first concave discharge portion 102 , First
It is discharged to the relay chamber 20 through the discharge hole 104, the passage s (see FIG. 4) and each through hole t, and then to FIGS.
As shown in FIG. 7, the gas is introduced into the rotor chamber 14 through the first introduction hole group 107 and further expanded between the adjacent vanes 42 to rotate the rotor 31. It is discharged from the group 110 to the outside.

As described above, the expansion of the high-temperature and high-pressure steam operates the piston 41 to rotate the rotor 31 through the vane 42, and the expansion of the low-temperature and low-pressure steam caused by the pressure drop of the high-temperature and high-pressure steam causes the rotor 31 to rotate through the vane 42. By rotating, an output is obtained from the output shaft 23.

In addition to the embodiment, the forward movement of the piston 41 is directly received by the roller 59 without passing through the vane 42 so that the forward movement of the piston 41 is converted into the rotational movement of the rotor 31. Can be converted into a rotational motion by the engagement of. Further, the vane 42 may be always separated from the inner peripheral surface 45 and the opposed inner end surface 47 of the rotor chamber 14 at substantially constant intervals by the cooperation of the roller 59 and the annular groove 60 as described above. , And each of the vane 42 and the roller 59 is specially formed in the annular groove 60.
You may work with

When the expander 4 is used as a compressor, the rotor 31 is rotated clockwise by the output shaft 23 in FIG. Rotor chamber 1 from 110 and 111
4 and the low-compressed air thus obtained is supplied from the first and second inlet hole groups 107 and 108 to the relay chamber 2.
0, each through hole t, passage s, first and second discharge holes 104, 10
5, supply the large-diameter cylinder hole f via the first and second concave discharge portions 102 and 103 and the through hole c, and operate the piston 41 by the vane 42 to convert the low-pressure air into high-pressure air;
The high-pressure air is introduced into the introduction pipe 80 through the hole c, the supply ports 90 and 91, and the first and second supply pipes 94 and 95.

Next, lubrication of each sliding portion of the expander 4 by a hydrostatic bearing using a pressurized liquid-phase fluid as a medium will be described.

As shown in FIG. 4, a pressurized liquid phase fluid supply pipe 130 is connected to a pressurized liquid phase fluid supply hole 129 formed in the hollow bearing cylinder 22 of the second half 9. The supply of the pressurized liquid-phase fluid to the pressurized liquid-phase fluid supply pipe 130 is performed by a supply pump 6 (see FIG. 1) that supplies water from the condenser 5 to the evaporator 3 under pressure. By using the supply pump 6 for supplying the pressurized liquid-phase fluid to the hydrostatic bearings of each part of the expander 4, a special pressurized liquid-phase fluid supply pump is not required, and the number of parts is reduced. .

The high-pressure pressurized liquid-phase fluid supplied from the pressurized liquid-phase fluid supply hole 129 is supplied to the static pressure bearing 25 on the second half 9 side.
Through the pressurized liquid phase fluid passage W5 formed in the
5 and the outer peripheral surface of the large-diameter portion 24 of the output shaft 23 reaches a sliding portion, and the outer peripheral surface of the output shaft 23 is supported in a floating state by a liquid film formed thereon. Lubrication is performed so that solid contact between the bearing 23 and the static pressure bearing 25 is prevented and seizure and wear do not occur. Pressurized liquid phase fluid supply hole 12
9 is a large-diameter portion 2 of the output shaft 23 through a plurality of pressurized liquid-phase fluid passages W6 formed in the large-diameter portion 24 of the output shaft 23 in the axial direction.
4 and a plurality of pressurized liquid-phase fluid passages W8 formed on the inner periphery of the large-diameter portion 24 of the output shaft 23. The pressurized liquid-phase fluid that has passed through the pressurized liquid-phase fluid passage W8 is supplied to the sliding portion between the outer peripheral surface of the large-diameter solid portion 66 of the fixed shaft 65 and the inner peripheral surface of the hollow shaft 64, and the fixed shaft 6
The lubricating portion between the outer peripheral surface of the small diameter solid portion 69 and the inner peripheral surface of the hole 67 of the output shaft 23 is lubricated.

The annular pressurized liquid-phase fluid passage W7 has a hollow shaft 6
A plurality of pressurized liquid-phase fluid passages W9 (FIG. 1)
7), and communicates with an annular pressurized liquid-phase fluid passage W10 symmetrically formed on the opposite side of the annular pressurized liquid-phase fluid passage W7 across the rotor 31. The pressurized liquid-phase fluid from the annular pressurized liquid-phase fluid passage W10 is supplied to a pressurized liquid-phase fluid passage W11 formed between the large-diameter portion 24 of the output shaft 23 and the hollow shaft 64. Static pressure bearing 25 on one half 8 side
Through the pressurized liquid-phase fluid passage W12 formed at the inner surface of the hydrostatic bearing 25 and the outer surface of the large-diameter portion 24 of the output shaft 23, and output by the liquid film formed there. Shaft 23
Of the output shaft 23 by supporting the outer peripheral surface of the
To prevent solid contact between the bearing and the static pressure bearing 25 and lubricate so as not to cause seizure or wear. Further, the pressurized liquid-phase fluid that has exited the pressurized liquid-phase fluid passage W11 lubricates a sliding portion between the outer peripheral surface of the fixed shaft 65 and the inner peripheral surface of the hollow shaft 64.
Of the hollow bearing cylinder 21 of the first half 8 is lubricated.

The pressurized liquid-phase fluid which has finished lubricating the sliding portion between the left and right static pressure bearings 25 and the output shaft 23 is formed between the rotor 31 and the first and second halves 8 and 9. Through the left and right pressurized liquid phase fluid passages W13, the first and second outlet groups 110,
111 (see FIG. 6) is discharged to the outside of the casing 7.

The pressurized liquid-phase fluid lubricated at the sliding portion between the outer peripheral surface of the large-diameter solid portion 66 of the fixed shaft 65 and the inner peripheral surface of the hollow shaft 64 is stored in the first fixed shaft 65. It is captured by seal grooves 55 and 56 (see FIGS. 4 and 17) formed on the outer peripheral surface of the fixed shaft 65 so as to surround the outer periphery of the second seal blocks 92 and 93, respectively. Since the pressure of the pressurized liquid-phase fluid is higher than the pressure of the steam discharged from the cylinder member 39, the leak of the steam can be sealed with the pressurized liquid-phase fluid in the seal grooves 55 and 56. Seal groove 55,
56 is the first and second concave discharge portions 102, 10 of the fixed shaft 65.
3 (see FIG. 17), the pressurized liquid-phase fluid is supplied from the first and second concave discharge portions 102 and 103 to the relay chamber 20 via the first and second discharge holes 104 and 105. , The first and second inlet groups 107 and 108, the rotor chamber 14, and the first and second outlet groups 110 and 1 therefrom.
It is discharged to the outside of the casing 7 through 11.

A part of the pressurized liquid-phase fluid passing through the pressurized liquid-phase fluid passage W13 flows into the slot-like space 34 in which the vane body 43 slides, and from there to between the pair of projections 53 of the vane body 43. The fluid flows into the open pressurized liquid phase fluid passages W1 and W2. The pressurized liquid-phase fluid passage W2 opened at the radially outer end of the vane body 43 opens into the rotor chamber 14 within a predetermined angle range in which the vane body 43 projects most from the rotor 31 (see FIG. 11). As shown in FIG. 18, the pressure in the slot-shaped space 34 is a constant value PF, and this pressure PF is determined by the opening degree of the orifice 54 (see FIG. 2) provided in the first half 8 and communicating with the relay chamber 20. It can be set arbitrarily. On the other hand, the pressure PQ in the rotor chamber 14 is
1, the pressurized liquid-phase fluid passage W2 of the vane body 43 is connected to the rotor chamber 14 in a range where the pressure PF of the slot-shaped space 34 exceeds the pressure PQ in the rotor chamber 14 (PF> PQ). If it is set to be open, the pressurized liquid-phase fluid accumulated in the slot-shaped space 34 can be discharged to the rotor chamber 14 through the pressurized liquid-phase fluid passages W1 and W2 of the vane body 43.

As is apparent from FIGS. 3 and 11, an annular pressurized liquid-phase fluid passage W7 formed on the outer periphery of the output shaft 23 is provided.
, The interior of the rotor 21 is divided into 12 pressurized liquid phase fluid passages W
14 extends radially, and the outer end of each of the side plates 40 serves as a pressurized liquid-phase fluid passage W15 branched into two branches.
Orifice forming member 124. Accordingly, the pressurized liquid-phase fluid supplied from the pressurized liquid-phase fluid supply pipe 130 (see FIG. 4) is supplied to the pressurized liquid-phase fluid passage W provided in the rotor 21.
14 and the pressurized liquid phase fluid passage W15 move radially outward while being further pressurized by centrifugal force, and
Is supplied to the orifice forming member 124 of the side plate 40 and is ejected from the orifice forming member 124 through the pressurized liquid phase fluid discharge passage 127 through the pressurized liquid phase fluid passage W3 of the side plate 40 to the vane sliding surface 121. A static pressure bearing is formed between the surface 121 and the vane 42 to support the vane 42 in a floating state, thereby preventing solid contact between the side plate 40 and the vane 42 to prevent seizure and wear. As described above, the pressurized liquid-phase fluid passages W14, W provided inside the rotor 31 with the pressurized liquid-phase fluid lubricating the vane sliding surface 121 of the vane 42.
By supplying the pressurized liquid-phase fluid through the pressurized liquid-phase fluid passage W3 provided inside the side plate 15 and the side plate 40, not only can the pressurized liquid-phase fluid be pressurized by centrifugal force, but also the temperature around the rotor 32 can be stabilized. Thus, the influence of thermal expansion can be reduced, and the set clearance can be maintained to minimize the leakage of steam.

As shown in FIG. 19, the circumferential load applied to each vane 42 (the load in the direction perpendicular to the plate-shaped vane body 43) is the same as that of the vane body 43 in the rotor chamber 14. This is a combined force of the load due to the differential pressure of the steam pressure applied to the surface and the circumferential component of the reaction force received by the roller 59 provided on the vane body 43 from the annular groove 60,
These loads change periodically according to the phase of the rotor 31. Therefore, the vane 42 which has received the offset load periodically exhibits a behavior of tilting between the pair of side plates 40 sandwiching the vane 42.

In this way, the vane body 4
When 3 is tilted, the gap between the six pressurized liquid-phase fluid discharge holes 127 opened in each side plate 40 and the vane body 43 changes, so that the liquid film in the portion where the gap has widened flows out, In addition, since it is difficult to supply the pressurized liquid-phase fluid to the portion where the gap is narrowed, no pressure is applied to the sliding portion, and the vane body 43 is connected to the vane sliding surface 121 of the side plate 40.
Wear may occur due to direct contact with the surface. However, according to this embodiment, the orifice 112 communicating with each of the six pressurized liquid fluid passages W3 connected to the six pressurized liquid fluid discharge holes 127 by the orifice forming member 124 provided in the side plate 40. Is formed, so that the above-mentioned problem is solved.

That is, when the gap between the pressurized liquid-phase fluid discharge hole 127 and the vane body 43 is widened, the supply pressure of the pressurized liquid-phase fluid is constant, and therefore, the constant pressure generated before and after the orifice 112 in a steady state. The differential pressure before and after the orifice 112 increases due to the orifice effect due to the increase in the flow rate of the pressurized liquid-phase fluid due to the increase in the amount of outflow from the gap, and the pressure in the gap decreases. As a result, a force for narrowing the widened gap and restoring the gap is generated. Further, when the gap between the pressurized liquid-phase fluid discharge hole 127 and the vane body 43 is narrowed, the amount of outflow from the gap decreases, and the differential pressure across the orifice 112 decreases. As a result, the pressure in the gap decreases. Is increased, and a force for restoring the narrowed gap is generated.

As described above, even if the gap between the pressurized liquid-phase fluid discharge hole 127 and the vane body 43 changes due to the load applied to the vane 42, the gap is supplied to the gap according to the change in the size of the gap. Since the orifice 112 automatically adjusts the pressure of the pressurized liquid-phase fluid, the clearance between the vane body 43 and the side plate 40 can be stably maintained at a desired size. Accordingly, the liquid film is always held between the vane body 43 and the side plate 40 to support the vane body 43 in a floating state, and the vane body 43 comes into solid contact with the vane sliding surface 121 of the side plate 40 to cause abrasion. Can be reliably avoided.

Since the pressurized liquid-phase fluid is held in two recesses 49 formed on both sides of the vane body 43,
The recess 49 serves as a pressure reservoir to suppress a pressure drop due to leakage of the pressurized liquid-phase fluid. As a result, the vane body 43 sandwiched between the vane sliding surfaces 121 of the pair of side plates 40 is floated by the pressurized liquid-phase fluid, and the sliding resistance can be reduced to a state near zero.
When the vane 42 reciprocates, the relative position of the vane 42 in the radial direction with respect to the rotor 21 changes. However, the recess 49 is provided not on the side plate 40 but on the vane main body 43 side. Is provided in the vicinity of the roller 59 on which the reciprocating motion is applied, so that the reciprocating vane 42 is always kept in a floating state, and the sliding resistance can be effectively reduced.

The vane 4 with respect to the side plate 40
The pressurized liquid-phase fluid that has lubricated the sliding surface of No. 2 moves radially outward due to centrifugal force, and slides between the seal member 44 provided on the outer peripheral surface of the end of the vane body 32 and the inner peripheral surface of the rotor chamber 14. Lubricate the parts. Then, the lubricated pressurized liquid-phase fluid flows from the rotor chamber 14 to the first and second outlet groups 110 and 111.
(See FIG. 6).

As is apparent from FIG. 3, annular grooves 131, 132 centering on the axis L are formed on the inner peripheral surfaces of the first and second halves 8, 9 of the casing 7, and these annular grooves are formed. Annular ring seals 134 and 135 having a pair of seal members 133 are fitted inside the grooves 131 and 132. The pressurized liquid-phase fluid discharge holes 128 of the pressurized liquid-phase fluid passage W4 provided in the vane body 43 face the inner ends of the ring seals 134 and 135 facing the pair of parallel portions 48 of the vane body 43. Open. On the other hand, the pressurized liquid-phase fluid supply port 137 provided in the second half 9 is connected to the pressurized liquid-phase fluid passages W16 and W16.
17, and communicate with the pressure chamber 136 at the outer end of the annular grooves 131 and 132 via W18. The pressurized liquid-phase fluid supply port 137 that supplies the pressurized liquid-phase fluid to the ring seals 134 and 135 is
It is a separate system independent of the pressurized liquid-phase fluid supply hole 129 (see FIG. 4) for supplying a pressurized liquid-phase fluid to each part of the expander 4.

The ring seals 134, 135 are directed toward both end faces 35 of the rotor 31 with the pressurized liquid-phase fluid supplied to the pressure chamber 136 as urging means formed at the bottom of the annular grooves 131, 132. Pressing and ring seal 13
The pressurized liquid-phase fluid discharge hole 128 of the vane body 43 is provided on the sliding surface between the inner peripheral surface of the vane body 43 and the parallel portion 48 of the vane body 43.
To form a hydrostatic bearing by supplying a pressurized liquid-phase fluid from the ring seals 134 and 135 floating inside the annular grooves 131 and 132 so that both end faces 35 of the rotor 31
Can be sealed so that the rotor chamber 14
It is possible to prevent the inside steam from leaking from both end surfaces 35 of the rotor 31. At this time, the ring seal 13
4, 135 and both end surfaces 35 of the rotor 31 are separated by the liquid film of the pressurized liquid-phase fluid supplied from the pressurized liquid-phase fluid discharge hole 128, so that they do not make solid contact with each other, and even if the rotor 31 is inclined. Following this, the ring seals 134, 135 are inclined in the annular grooves 131, 132, so that a stable sealing performance can be secured while minimizing the frictional force.

Since the ring seals 134 and 135 are non-contact seals, not only the frictional force is extremely small, but also
The service life is semi-permanent as compared with the contact-type seal, and since there is a liquid film between the seal and the rotor 31, there is no risk of seizure. Since the ring-shaped seals 134 and 135 are provided, the slot-shaped space 34 around the fixed shaft 65 becomes a closed space.
By adjusting the pressure in the slot-shaped space 34 in the step (see FIG. 3), the amount of steam in the rotor chamber 14 leaking from both end surfaces 35 of the rotor 31 can be further reduced.
Ring seals 134 and 135 and both end surfaces 35 of rotor 31
The pressurized liquid-phase fluid that has lubricated the sliding portion is supplied to the rotor chamber 14 by centrifugal force, and is discharged therefrom through the first and second outlet groups 110 and 111 to the outside of the casing 7.

Instead of urging the ring seals 134, 135 toward the rotor 31 with the pressurized liquid-phase fluid, it is also possible to urge the ring seals with a resilient force of a metal spring or rubber.

In the present embodiment described above, the evaporator 3 that generates high-temperature and high-pressure steam by heating water with the heat energy of the exhaust gas of the internal combustion engine 1, and converts the high-temperature and high-pressure steam supplied from the evaporator 3 to a constant torque. An expander 4 that converts the shaft output into a shaft output, a condenser 5 that liquefies the temperature-reduced pressure-reduced steam discharged by the expander 4, and a supply pump 6 that supplies the water liquefied by the condenser 5 to the evaporator 3. In the Rankine cycle, a positive displacement expander 4 is employed. The positive displacement expander 4 is capable of recovering energy with high efficiency in a wide rotational speed range from low speed to high speed as compared with a non-positive displacement expander such as a turbine. It also has excellent follow-up and responsiveness to changes in the thermal energy of the exhaust gas (changes in the temperature and flow rate of the exhaust gas) due to the increase and decrease in the number of revolutions. Moreover, the expander 4
Since the first energy conversion means composed of the cylinder member 39 and the piston 41 and the second energy conversion means composed of the vane 42 are connected in series and arranged inside and outside in the radial direction, expansion is achieved. The efficiency of heat energy recovery by the Rankine cycle can be further improved while reducing the size and weight of the machine 4 to improve space efficiency.

Since water as a working medium for operating the expander 4 is also used as a pressurized liquid-phase fluid for lubrication, each lubricating portion of the expander 4 can be statically operated without requiring special lubricating oil. Lubrication by the pressure bearing can prevent seizure and wear. Moreover, since the lubricating oil does not mix with the water as the working medium, it is possible to avoid the adverse effect of mixing the water and the lubricating oil. Further, the support of the output shaft 23 that rotatably supports the rotor 31, the support of the vanes 42 in the slot-like space 34, and the support of the ring seals 134 and 135 on both end surfaces 35 of the rotor 31 are performed by the hydrostatic bearings. Thus, the solid contact of the sliding portion can be prevented, and the occurrence of seizure and wear can be reliably prevented.

Although the embodiments of the present invention have been described in detail, various design changes can be made in the present invention without departing from the gist thereof.

For example, in the embodiment, water as a working medium for operating the expander 4 is also used as the pressurized liquid-phase fluid for the hydrostatic bearing, but the expander is used as the pressurized liquid-phase fluid for the hydrostatic bearing. Oil and the like different from the working medium of 4 can be used,
As a result, the lubricity and sealing properties between the members such as the vane 42, the rotor 31, the casing 7, and the output shaft 23 can be further improved. In this case, the working medium of the expander 4 may be mixed with a pressurized liquid-phase fluid for a hydrostatic bearing, such as oil, which is different from the working medium, but a separating device for separating the working medium and the pressurized liquid-phase fluid. There is no hindrance if it is provided. If the working medium of the expander 4 is also used as the pressurized liquid-phase fluid for the hydrostatic bearing as in the above-described embodiment, a device for separating the working medium and the pressurized liquid-phase fluid becomes unnecessary.

[0069]

As described above, according to the first aspect of the present invention, the static pressure bearing of the output shaft, which rotates integrally with the rotor, and the other static pressure bearing located radially outside the rotor. Since the pressurized liquid-phase fluid is supplied and supported in a floating state by static pressure, the occurrence of solid contact can be avoided and the occurrence of seizure or wear in the sliding portion can be reliably prevented. Further, since the pressurized liquid-phase fluid is supplied to the other hydrostatic bearing via a passage formed inside the rotor, the pressurized liquid-phase fluid is further pressurized by the centrifugal force accompanying the rotation of the rotor, so that an effective static liquid is supplied. Pressure support can be provided.

According to the second aspect of the present invention,
By further pressurizing the pressurized liquid-phase fluid with the centrifugal force accompanying the rotation of the rotor, the vane is effectively supported by the static pressure in the slot-like space formed in the rotor and floated, thereby preventing the occurrence of seizure and wear. Can be.

According to the third aspect of the present invention,
By further pressurizing the pressurized liquid-phase fluid with the centrifugal force associated with the rotation of the rotor, the side surface of the rotor is effectively supported by the static pressure on the inner surface of the casing and floated, and the occurrence of seizure and wear can be prevented. .

According to the fourth aspect of the present invention,
Since the pressurized liquid phase fluid and the gas phase working medium are the same fluid,
Not only does a special pressurized liquid-phase fluid for the hydrostatic bearing become unnecessary, but also the adverse effect caused by mixing different kinds of pressurized liquid-phase fluids into the working medium can be avoided.

[Brief description of the drawings]

FIG. 1 is a schematic diagram of a waste heat recovery device for an internal combustion engine.

FIG. 2 is a longitudinal sectional view of the expander corresponding to the sectional view taken along line 2-2 of FIG. 6;

FIG. 3 is a longitudinal sectional view of the expander corresponding to a sectional view taken along line 3-3 in FIG. 5;

FIG. 4 is an enlarged sectional view around the rotation axis of FIG. 2;

FIG. 5 is a sectional view taken along line 5-5 of FIG. 2;

FIG. 6 is a view taken along line 6-6 in FIG. 2;

FIG. 7 is an explanatory diagram showing a cross-sectional shape of a rotor chamber and a rotor.

FIG. 8 is a view taken along line 8-8 in FIG. 6;

9 is a view in the direction of arrow 9 in FIG. 8;

FIG. 10 is a sectional view taken along line 10-10 of FIG. 8;

FIG. 11 is an enlarged view of a main part of FIG. 5;

FIG. 12 is a view taken along line 12-12 of FIG. 6;

FIG. 13 is a view in the direction of arrow 13 in FIG. 12;

14 is a sectional view taken along line 14-14 in FIG.

FIG. 15 is a sectional view taken along line 15-15 of FIG. 14;

FIG. 16 is a sectional view taken along line 16-16 of FIG. 11;

FIG. 17 is an enlarged view around the rotation axis of FIG. 5;

FIG. 18 is a graph showing a relationship between the pressure of the slot-like space and the pressure of the rotor chamber, PQ, with respect to the phase of the rotor.

FIG. 19 is a graph showing a relationship between a rotor phase and a load applied to a vane.

[Explanation of symbols]

 Reference Signs List 7 casing 14 rotor chamber 23 output shaft 25 static pressure bearing 31 rotor 32 slot-like space 41 piston 42 vane L axis U1-U12 vane piston unit W14 pressurized liquid phase fluid passage (passage)

 ──────────────────────────────────────────────────続 き Continued on the front page (72) Inventor Takeshi Baba 1-4-1, Chuo, Wako-shi, Saitama Prefecture Inside Honda R & D Co., Ltd. (72) Kensuke Honma 1-4-1, Chuo, Wako-shi, Saitama Inside the Honda R & D Co., Ltd. (72) Inventor Hiroyoshi Taniguchi 1-4-1 Chuo, Wako-shi, Saitama F-term inside the Honda R & D Co., Ltd. (Reference) 3H040 AA10 BB05 BB11 CC06 CC09 DD07 DD09 DD13 DD25 DD31 3H076 AA03 AA10 AA16 AA37 BB17 BB21 CC28 CC31 CC36 CC61 CC91

Claims (4)

[Claims] 1. A casing (7) having a rotor chamber (14), a rotor (31) housed in the rotor chamber (14), and a rotor (31) provided around the rotation axis (L). A plurality of vane piston units (U1 to U) which are arranged radially and are reciprocally movable in the radial direction.
12), and each vane piston unit (U1 to U)
12) is composed of a vane (42) sliding in the rotor chamber (14) and a piston (41) abutting on the non-sliding side of the vane (42). Actuates the piston (41) to rotate the rotor (31) via a power converter and expands the vane (42) by expansion of the low-pressure gas-phase working medium due to the pressure drop of the high-pressure gas-phase working medium. A rotary fluid machine that rotates the rotor (31) via the rotor (31), by supplying a pressurized liquid-phase fluid to a static pressure bearing (25) of an output shaft (23) that rotates integrally with the rotor (31). 23) and a passage (W14) formed inside the rotor (31) while part of the pressurized liquid phase fluid is formed inside the rotor (31).
A rotary fluid machine, wherein the fluid is supplied to another hydrostatic bearing located radially outside of the rotor (31) via a rotary shaft. 2. The method according to claim 1, wherein the other hydrostatic bearing includes a rotor (31).
The rotary fluid machine according to claim 1, characterized in that the vane (42) is statically supported in the slot-shaped space (34) formed in the rotary fluid machine. 3. The rotary fluid machine according to claim 1, wherein the other hydrostatic bearing is configured to support the side surface of the rotor (31) by hydrostatic pressure on an inner surface of the casing (7). . 4. The pressurized liquid-phase fluid is the same fluid as the gas-phase working medium.
The rotary fluid machine according to any one of the above.