US20050089410A1

US20050089410A1 - Rotating fluid machine

Info

Publication number: US20050089410A1
Application number: US10/768,002
Authority: US
Inventors: Kenji Matsumoto; Naoki Itoh
Original assignee: Individual
Current assignee: Honda Motor Co Ltd
Priority date: 2003-02-03
Filing date: 2004-02-02
Publication date: 2005-04-28
Also published as: JP2004239069A

Abstract

A rotary valve of a rotating fluid machine is constructed by bringing, into contact on sliding faces, a moving valve plate and a stationary valve plate, and high pressure working medium passages and low pressure working medium passages penetrate mating faces of a valve body and the stationary valve plate. By sealing the outer circumference of the mating faces with a first sealing member and sealing the circumference of the low pressure working medium passages with a second sealing member inside the first sealing member in the radial direction, the pressure of a high pressure working medium leaking out of the high pressure working medium passages to the mating faces is caused to act on substantially the whole area of the mating faces, thereby preventing local deformation of the stationary valve plate and ensuring close contact between the sliding faces.

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention
The present invention relates to a rotating fluid machine provided with a casing, a rotor rotatably supported by the casing, a working section provided on the rotor, and a rotary valve, provided between the casing and the rotor, for controlling the supply and discharge of the working medium to and from the working section.
2. Description of the Related Art
In a rotary valve for rotating fluid machines of this kind, a moving valve plate provided on the rotor and a stationary valve plate provided on a valve body engaged with the casing to be unable to rotate and movable in the direction of the axis of the rotor are brought into contact with each other on sliding faces orthogonal to the axis, and the rotation of the moving valve plate relative to the stationary valve plate causes steam of high temperature and high pressure to be successively supplied to or discharged from a group of axial piston cylinders provided on the rotor.
The device described in Japanese Patent Laid-Open No. 2002-256805 has two sealing members on the mating faces of the valve body and the stationary valve plate so as to surround the high pressure working medium passages and low pressure working medium passages penetrating the mating faces. These sealing members obstruct a high pressure working medium from leaking out of the high pressure working medium passages to be discharged outside the mating faces or into the low pressure working medium passages.
In the above-described conventional device, with regard the mating faces of the valve body and the stationary valve plate, a high pressure acts in the radially inner direction on the annular sealing members surrounding the high pressure working medium passages, and a low pressure acts in the radially inner direction on the annular sealing members surrounding the low pressure working medium passages. Therefore, there is a possibility that the high pressure and the low pressure act locally on the stationary valve plate to invite flexural deformation, whereby the close contact between its sliding face and that of the moving valve plate is lost, so that high temperature high pressure steam leaks.
In view of the problems noted above, the present invention has an object to eliminate the pressure imbalance on the mating faces of the valve body and the stationary valve plate of the rotary valve in the rotating fluid machine, thereby ensuring close contact between the sliding faces and the moving valve plate.

SUMMARY OF THE INVENTION

In order to achieve the object stated above, according to the present invention, there is proposed a rotating fluid machine comprising: a casing; a rotor rotatably supported by the casing; a working section disposed on the rotor; and a rotary valve, provided between the casing and the rotor, for controlling the supply and discharge of a working medium to and from the working section, the rotary valve being constructed by bringing, into contact on sliding faces, a moving valve plate provided on the rotor and a stationary valve plate provided on a valve body, and high pressure working medium passages and low pressure working medium passages penetrating mating faces of the valve body and the stationary valve plate, wherein the outer circumference of the mating faces is sealed with an annular first sealing member and the circumference of the low pressure working medium passages is sealed with an annular second sealing member inside the first sealing member in the radial direction.
With the configuration described above, as the arrangement of the annular first sealing member on the outer circumference of the mating faces of the valve body and the stationary valve plate of the rotary valve of the rotating fluid machine encloses both the high pressure working medium passages and the low pressure working medium passages penetrating the mating faces inside in their radial direction, the pressure of a high pressure working medium leaking out of the high pressure working medium passages to the mating faces is caused to act on substantially the whole area of the mating faces, thereby preventing local deformation of the stationary valve plate and ensuring close contact between the sliding faces. Also, since the annular second sealing member is arranged around the low pressure working medium passages inside the first sealing member in the radial direction, the second sealing member prevents the high pressure working medium on the mating faces from being discharged to the low pressure working medium passages. As the second sealing member encloses only a narrow area around the low pressure working medium passages, there is no fear that low pressure locally deforms the stationary valve plate.
Further, according to a second feature of the present invention, in addition to the first feature, the first and second sealing members are C-shaped seals each having an opening in its section, the opening of the first sealing member is arranged inward in the radial direction, and the opening of the second sealing member is arranged outward in the radial direction.
With the configuration described above, as the opening of the first sealing member comprising a C-shaped seal is arranged inward in the radial direction, the sealing performance of the first sealing member can be increased by the pressure of a high pressure working medium leaking from the high pressure working medium passages to the mating faces, thereby preventing the high pressure working medium from being discharged outside from the mating faces. Further, as the opening of the second sealing member comprising a C-shaped seal is arranged outward in the radial direction, the sealing performance of the second sealing member can be increased by the pressure of a high pressure working medium leaking to the mating faces, thereby preventing the high pressure working medium from being discharged out of the mating faces to the low pressure working medium passages.
A group of axial piston cylinders 56 in an embodiment correspond to the working section in the invention; sealing members 89 and 90 in the embodiment correspond to the first sealing member and the second sealing member, respectively, in the invention; first and second steam passages P1 and P2 in the embodiment correspond to the high pressure working medium passages in the invention; and the fifth and sixth steam passages P5 and P6 in the embodiment correspond to the low pressure working medium passages in the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a vertical sectional view of an expander according to a preferred embodiment.
FIG. 2 is a sectional view taken on line 2-2 in FIG. 1.
FIG. 3 is a view taken on line 3-3 in FIG. 1.
FIG. 4 is an enlarged view of Part 4 in FIG. 1.
FIG. 5 is an enlarged view of Part 5 in FIG. 1.
FIG. 6 is an exploded perspective view of a rotor.
FIG. 7 is a sectional view taken on line 7-7 in FIG. 4.
FIG. 8 is a sectional view taken on line 8-8 in FIG. 4.
FIG. 9 is an enlarged view of Part 9 in FIG. 4.
FIG. 10 is a sectional view taken on line 10-10 in FIG. 5.
FIG. 11 is a sectional view taken on line 11-11 in FIG. 5.
FIG. 12 is a sectional view taken on line 12-12 in FIG. 5.
FIG. 13 is a sectional view taken on line 13-13 in FIG. 5.
FIG. 14 is a view in arrowed direction 14 in FIG. 13.
FIG. 15 is a view in arrowed direction 15 in FIG. 13.
FIG. 16 is an enlarged view of Part 16 in FIG. 5.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

A preferred embodiment of the present invention will be described below with reference to the accompanying drawings.
As shown in FIG. 1 through FIG. 9, an expander E according to an embodiment is used in, for example, a Rankine cycle system. It converts thermal energy and pressure energy of high-temperature high-pressure steam as a working medium into mechanical energy and supplies the converted energy. The casing 11 of the expander E is provided with a casing body 12, a front cover 15 connected to the front opening of the casing body 12 with a plurality of bolts 14 . . . with a sealing member 13 therebetween, a rear cover 18 fitted to the rear opening of the casing body 12 with a plurality of bolts 17 . . . with a sealing member 16 therebetween, and an oil pan 21 fitted to the bottom opening of the casing body 12 with a plurality of bolts 20 . . . with a sealing member 19 therebetween.
A rotor 22 is arranged to be rotatable around an axis L extending in the middle of the casing 11 in the back and forth directions, and supported in front by combined angular bearings 23 f and 23 r disposed on the front cover 15 and on the back by a radial bearing 24 disposed on the casing body 12. A swash plate holder 28 is integrally formed on the rear face of the front cover 15. A swash plate 31 is rotatably supported by this swash plate holder 28 via an angular bearing 30. The axis of the swash plate 31 is inclined relative to the axis L of the rotor 22 at a fixed angle.
The rotor 22 is provided with an output shaft 32 supported on the front cover 15 with the combined angular bearings 23 f and 23 r, three sleeve supporting flanges 33, 34 and 35 formed integrally with one another on the rear part of the output shaft 32 via notches 57 and 58 of a predetermined width (see FIG. 4 and FIG. 9), a rotor head 38 connected to the rear sleeve supporting flange 35 with a plurality of bolts 37 . . . via a metal gasket 36 integrally and supported on the casing body 12 by the radial bearing 24, and a thermally insulating cover 40 fitted onto the three sleeve supporting flanges 33, 34 and 35 from front and connected with a plurality of bolts 39 . . . to the front sleeve supporting flange 33.
Five sleeve supporting holes 33 a . . . , 34 a . . . and 35 a . . . are respectively bored in the three sleeve supporting flanges 33, 34 and 35 around the axis L at 72° intervals. Five cylinder sleeves 41 . . . are fitted into the respective sleeve supporting holes 33 a . . . , 34 a . . . and 35 a . . . from behind. Formed at the rear end of each of the cylinder sleeves 41 is a flange 41 a, which is positioned in the axial direction in contact with the metal gasket 36 in a state in which it is fitted onto a stepped portion 35 b formed in the sleeve supporting hole 35 a of the rear sleeve supporting flange 35 (see FIG. 9). A piston 42 is slidably fitted within each of the cylinder sleeves 41, the front end of the piston 42 is in contact with a dimple 31 a formed in the swash plate 31, and a steam expansion chamber 43 is partitioned between the rear end of the piston 42 and the rotor head 38.
A plate-shaped bearing holder 92 is laid over the front face of the front cover 15 with a sealing member 91 therebetween and fixed with bolts 93 . . . . A pump body 95 is laid over the front face of the bearing holder 92 with a sealing member 94 therebetween and fixed with bolts 96 . . . . The combined angular bearings 23 f and 23 r are positioned between the stepped portion of the front cover 15 and the bearing holder 92, and fixed in the direction of the axis L.
A shim 97 of a predetermined thickness is placed between a flange 32 d formed in the output shaft 32 supporting the combined angular bearings 23 f and 23 r and the inner races of the combined angular bearings 23 f and 23 r. The inner races of the combined angular bearings 23 f and 23 r are fastened with nuts 98 screwed onto the outer circumference of the output shaft 32. As a result, the output shaft 32 is positioned in the direction of the axis L relative to the combined angular bearings 23 f and 23 r, namely with respect to the casing 11.
The combined angular bearings 23 f and 23 r are attached in mutually reverse orientations, and support the output shaft 32 not only in the radial direction but also immovably in the direction of the axis L. Thus, one combined angular bearing 23 f is arranged to restrict the forward movement of the output shaft 32, while the other combined angular bearing 23 r is arranged to restrict the backward movement of the output shaft 32.
As the combined angular bearings 23 f and 23 r are used as a bearing for the front part of the rotor 22, one of the loads arising toward the opposite ends of the axis L in the expansion chambers 43 . . . in a predetermined operating state of the expander E is transmitted via the rotor 22 to the inner races of the combined angular bearings 23 f and 23 r, and the other load is transmitted via the swash plate 31 and the swash plate holder 28 of the front cover 15 to the outer races of the combined angular bearings 23 f and 23 r. These two loads compress the swash plate holder 28 of the front cover 15 held between the angular bearing 30 supporting the swash plate 31 and the combined angular bearings 23 f and 23 r supporting the rotor 22, resulting in an enhanced rigidity of the mechanism. Moreover, the integral configuration of the swash plate holder 28 with the front cover 15 as in this embodiment of the invention makes the structure more rigid and simpler.
Further, by incorporating the angular bearing 30 supporting the swash plate 31 and the combined angular bearings 23 f and 23 r supporting the rotor 22 into the front cover 15, it is possible to accomplish the assembling process in the units of “the rotor 22 and the piston 42 . . . ”, “assembly of the front cover 15” and “the pump body 95” thereby improving the efficiency of procedures such as rearrangement of the piston 42 . . . and the replacement of an oil pump 49.
The radial bearing 24 supporting the rotor head 38 which constitutes the rear end of the rotor 22 is an ordinary ball bearing supporting only the load in the radial direction. To enable the rotor head 38 to slide in the direction of the axis L relative to the radial bearing 24, a gap a is formed between the rotor head 38 and the inner race of the radial bearing 24 (see FIG. 5).
An oil passage 32 a extending on the axis L is formed within the output shaft 32 integral with the rotor 22. The front end of the oil passage 32 a branches in radial directions to communicate with an annular groove 32 b on the outer circumference of the output shaft 32. In a radially inner position of the sleeve supporting flange 34 at the center of the rotor 22, an oil passage blocking member 45 is screwed into the inner circumference of the oil passage 32 a with a sealing member 44 therebetween. A plurality of oil holes 32 c . . . extend from the nearby oil passage 32 a outward in the radial direction, and open in the outer circumferential face of the output shaft 32.
A trochoidal oil pump 49 is arranged between a concave 95 a formed in the front face of the pump body 95 and a pump cover 48 fixed with a plurality of bolts 47 . . . to the front face of the pump body 95 with a sealing member 46 therebetween, and includes an outer rotor 50 rotatably fitted into the concave 95 a, and an inner rotor 51 fixed to the outer circumference of the output shaft 32 to engage with the outer rotor 50. The inner space of the oil pan 21 communicates with the intake port 53 of the oil pump 49 via an oil pipe 52 and the oil passage 95 b of the pump body 95. The discharge port 54 of the oil pump 49 communicates with the annular groove 32 b of the output shaft 32 via the oil passage 95 c of the pump body 95.
The piston 42 slidably fitted into the cylinder sleeve 41 consists of an end portion 61, a middle portion 62 and a top portion 63. The end portion 61 is a member having a spherical portion 61 a in contact with the dimple 31 a of the swash plate 31, and is welded onto the tip of the middle portion 62. The middle portion 62 is a cylindrical member having a large-capacity hollow space 62 a, and has in the outer circumferential part near the top portion 63 a smaller diameter part 62 b slightly reduced in diameter. A plurality of oil holes 62 c . . . are formed to penetrate the smaller diameter part 62 b in the radial direction. A plurality of spiral oil grooves 62 d . . . are formed in the outer circumferential part ahead of the smaller diameter part 62 b. The top portion 63 facing the expansion chambers 43 is formed integrally with the middle portion 62. A thermally insulating space 65 (see FIG. 9) is formed between a partition wall 63 a formed inside the space and a lid member 64 fitted and welded onto its rear end face. Fitted to the outer circumference of the top portion 63 are two compression rings 66 and one oil ring 67. An oil ring groove 63 b into which the oil ring 67 is fitted, communicates via a plurality of oil holes 63 c . . . with the hollow space 62 a of the middle portion 62.
The end portion 61 and the middle portion 62 of the piston 42 are made of high carbon steel, and the top portion 63, of stainless steel. The end portion 61 undergoes induction quenching, and the middle portion 62, plain quenching. As a result, the piston 42 obtains a high surface stress resistance in the end portion 61 which is in contact with the swash plate 31 under a high surface stress, a wear resistance in the middle portion 62 which is in sliding contact with the cylinder sleeves 41 under poor lubricating conditions, and a heat and corrosion resistance in the top portion 63 which faces the expansion chambers 43 to be exposed to high temperature and high pressure.
An annual groove 41 b (see FIG. 6 and FIG. 9) is formed in the outer circumference of the middle portion of each cylinder sleeve 41, and a plurality of oil holes 41 c . . . are formed in this annual groove 41 b. Irrespective of the mounting position of the cylinder sleeve 41 in the rotating direction, the oil holes 32 c . . . formed in the output shaft 32 and oil holes 34 b . . . (see FIG. 4 and FIG. 6) formed in the middle sleeve supporting flange 34 of the rotor 22 communicate with the annual groove 41 b. A space 68 formed between the thermally insulating cover 40 and the sleeve supporting flanges 33 and 35 respectively before and behind the rotor 22 communicates with the inner space of the casing 11 via oil holes 40 a . . . (see FIG. 4 and FIG. 7) formed in the thermally insulating cover 40.
An annular lid member 69 is welded onto the front side of the rotor head 38 connected with the bolts 37 . . . to the rear face of the sleeve supporting flange 33 in the front side of the rotor 22, or onto the expansion chambers 43 . . . An annular thermally insulating space 70 (see FIG. 9) is defined on the back or rear face of the lid member 69. The rotor head 38 is positioned in the rotating direction by a knock pin 55 relative to the rear sleeve supporting flange 35.
The five cylinder sleeves 41 . . . and the five pistons 42 . . . constitute a group of axial piston cylinders 56 according to the present invention.
Next will be described with reference to FIG. 5 and FIG. 10 through FIG. 15 the structure of a rotary valve 71 for supplying and discharging steam to and from the five expansion chambers 43 . . . of the rotor 22.
As shown in FIG. 5, the rotary valve 71 arranged along the axis L of the rotor 22 is provided with a valve body 72, a stationary valve plate 73, and a moving valve plate 74. The moving valve plate 74, in a state of being positioned by a knock pin 75 in the rotating direction on the rear side of the rotor 22, is fixed with bolts 76 screwed onto the oil passage blocking member 45 (see FIG. 4). The bolts 76 also have a function to fix the rotor head 38 to the output shaft 32.
As is clear from FIG. 5, the stationary valve plate 73 in contact with the moving valve plate 74 via the flat sliding faces 77 is fixed to the center of the front face of the valve body 72 with a single bolt 78, and fixed to the outer circumference of the valve body 72 with an annular fixed ring 79 and a plurality of bolts 80. When it is fixed, a stepped portion 79 a formed on the inner circumference of the fixed ring 79 is pressed onto the outer circumference of the stationary valve plate 73 in a spigot-fit manner, and a stepped portion 79 b formed on the outer circumference of the fixed ring 79 is spigot-fitted onto the outer circumference of the valve body 72, thereby ensuring a coaxial relationship of the stationary valve plate 73 to the valve body 72. Further, a knock pin 81 for positioning the stationary valve plate 73 in the rotational direction is arranged between the valve body 72 and the stationary valve plate 73.
Therefore, as the rotor 22 turns, the moving valve plate 74 and the stationary valve plate 73 turn relative to each other in close contact with each other on the sliding faces 77. The stationary valve plate 73 and the moving valve plate 74 are made of a highly durable material, such as carbon or ceramic, and their durability can be further enhanced by affixing a member having excellent heat resistance, lubricating performance, corrosion resistance and wear resistance to the sliding faces 77, or by coating them with such a material.
The valve body 72 made of stainless steel is a stepped columnar member having a larger diameter part 72 a and a smaller diameter part 72 b. The outer circumferential faces of those larger diameter part 72 a and smaller diameter part 72 b are respectively fitted onto the supporting faces 18 a and 18 b having a circular section in the rear cover 18 with sealing members 82 and 83 therebetween to be slidable in the direction of the axis L.
As is clear from FIG. 5 and FIG. 15 referenced together, three supporting holes 18 c . . . are formed in the inner face of the rear cover 18 opposite a stepped portion 72 c of the valve body 72. The larger diameter parts 84 a . . . of three engaging pins 84 . . . are pressed into these supporting holes 18 c . . . while the smaller diameter parts 84 b . . . of those engaging pins 84 . . . are slidably fitted into three pin holes 72 d . . . formed in the stepped portion 72 c of the valve body 72. The three engaging pins 84 . . . are arranged at 120° intervals on the same circle so as to surround the axis L. While restricting the valve body 72 in the rotational direction with these engaging pins 84 . . . thereby preventing them from being dragged by and accompanying the rotor 22 in rotation, the valve body 72 is made movable in the direction of the axis L to secure the close contact between the sliding faces 77.
Referring again to FIG. 5, a plurality of preload springs 85 are supported by the rear cover 18 so as to surround the axis L, and the valve body 72 in which the stepped portion 72 c between the larger diameter part 72 a and the smaller diameter part 72 b is pressed by these preload springs 85 . . . is urged forward to bring the sliding faces 77 of the stationary valve plate 73 and the moving valve plate 74 into close contact with each other.
A steam feed pipe 86 connected to the rear face of the valve body 72 communicates with the sliding faces 77 via a first steam passage P1 formed within the valve body 72 and a second steam passage P2 formed in the stationary valve plate 73. Among the casing body 12, the rear cover 18 and the rotor 22, there is formed a steam discharge chamber 88 sealed with a sealing member 87. The steam discharge chamber 88 communicates with the sliding faces 77 via sixth and seventh steam passages P6 and P7 formed within the valve body 72 and a fifth steam passage P5 formed in the stationary valve plate 73.
As is clear from FIG. 5 when referenced together with FIG. 13 and FIG. 16, on the mating faces of the valve body 72 and the stationary valve plate 73, there is provided a sealing member 89 large enough to surround at the same time the connecting part between the high pressure first and second steam passages P1 and P2 and that between the low pressure fifth and sixth steam passages P5 and P6. The sealing member 89 is a so-called C-type seal having a C-shaped section, surrounding the most part of the mating faces of the valve body 72 and the stationary valve plate 73. The open portion of its C-shape is open inward in the radial direction. The mating faces are also provided with a smaller sealing member 90 surrounding only the connecting part between the low pressure fifth and sixth steam passages P5 and P6. The sealing member 90 also is a C-type seal, and the open portion of its C shape is open outward in the radial direction.
Therefore, even if high temperature high pressure steam leaks out from the connecting part between the high pressure first and second steam passages P1 and P2 to the mating faces between the valve body 72 and the stationary valve plate 73, the high temperature high pressure steam is blocked by the sealing member 89 covering the whole mating faces, thereby preventing the steam from leaking outward. In this case, as the C-shaped opening of the sealing member 89 comprising a C-type seal is arranged inward in the radial direction, even if the high temperature high pressure steam has a force to leak out from inside to outside in the radial direction, the steam pressure will expand the sealing member 89 to increase its sealing capability, thereby reliably preventing the high temperature high pressure steam from leaking out.
Also, as the sealing member 90 surrounds the connecting part between the low pressure fifth and sixth steam passages P5 and P6 opening to the mating faces of the valve body 72 and the stationary valve plate 73, any high temperature high pressure steam leaking out from the connecting part between the high pressure first and second steam passages P1 and P2 to the mating faces is prevented from short-circuiting to the low pressure fifth and sixth steam passages P5 and P6, and thus wasteful discharge of high temperature high pressure steam is prevented. In this case, as the C-shaped opening of the sealing member 90 comprising a C-type seal is arranged outward in the radial direction, even if the high temperature high pressure steam has a force to leak out from outside to inside in the radial direction, the steam pressure will expand the sealing member 90 to increase its sealing capability, thereby reliably preventing the high temperature high pressure steam from leaking out to the fifth and sixth steam passages P5 and P6.
As the sealing member 89 keeps high pressure on substantially the whole area of the mating faces, namely the area other than the inside of the smaller sealing member 90 surrounding the fifth and sixth steam passages P5 and P6, deformation is prevented by uniformly pressing the carbon-made stationary valve plate 73 having a relatively low rigidity, and bringing it into close contact with the moving valve plate 74 on its sliding face 77 without any gap, thereby preventing the leakage of the high temperature high pressure steam and uneven wear of the sliding faces 77.
The sliding faces 77 of the stationary valve plate 73 and the moving valve plate 74 of the rotary valve 71 are not always strictly orthogonal to the axis L, but may be slightly inclined as a consequence of machining error or uneven wear. In this state, if the stationary valve plate 73 and the moving valve plate 74 turn relative to each other while sliding on the sliding faces 77, the valve body 72 supported by the rear cover 18 with the sealing member 82 and 83 therebetween oscillate around the axis L using the smaller diameter part 72 b as a fulcrum within the compression margin of the sealing member 82 and 83.
In this case, as the valve body 72 is engaged with the rear cover 18 by the three engaging pins 84 . . . arranged at equal intervals around the axis L, the valve body 72 can oscillate around the axis L, so that it is possible to increase the compliance of the sliding faces 77, thereby preventing the leakage of the high temperature high pressure steam and suppressing further uneven wear of the sliding faces 77. In order to facilitate the oscillation of the valve body 72, it is preferable to bring the positions of the engaging pins 84 . . . as close as possible to the axis L. Moreover, as the engaging pins 84 . . . are disposed not on the valve body 72 side but on the rear cover 18 side, an increase in the inertial mass of the valve body 72 can be minimized to further enhance the compliance of the sliding faces 77.
Referring again to FIG. 5, five third steam passages P3 . . . arranged at equal intervals around the axis L penetrate the moving valve plate 74, and both ends of five fourth steam passages P4 . . . formed in the rotor 22 so as to surround the axis L communicate with the third steam passages P3 . . . and the expansion chambers 43 . . . , respectively. While the parts opening in the sliding faces 77 of the second steam passages P2 are circular, those opening in the sliding faces 77 of the fifth steam passage P5 are formed in an arcuate shape centering on the axis L.
Next will be described the operation of the expander E according to the embodiment configured as described above.
High temperature high pressure steam generated by heating water in an evaporator flows from the steam feed pipe 86, and reaches the sliding face 77 of the moving valve plate 74 via the first steam passage P1 formed in the valve body 72 of the rotary valve 71 and the second steam passage P2 formed in the stationary valve plate 73 integral with this valve body 72. The second steam passage P2 opening in the sliding face 77 momentarily communicates for a predetermined air intake period with the corresponding third steam passage P3 formed in the moving valve plate 74 turning integrally with the rotor 22. The high temperature high pressure steam is supplied from the third steam passage P3 via the fourth steam passage P4 formed in the rotor 22, into the expansion chamber 43 within the cylinder sleeve 41.
Even after the communication between the second steam passage P2 and the third steam passage P3 is cut off along with the rotation of the rotor 22, expansion of the expansion chamber 43 causes the piston 42 fitted into the cylinder sleeve 41 to be thrust forward from the top dead center to the bottom dead center, so that the end portion 61 at the front end of the piston presses the dimple 31 a in the swash plate 31. As a result, the reaction force which the piston 42 receives from the swash plate 31 gives a rotational torque to the rotor 22. Every time the rotor 22 turns a ⅕ round, high temperature high pressure steam is supplied to a newly adjacent expansion chamber 43 to drive the rotor 22 for continuous rotation.
While the piston 42 having reached the bottom dead center along with the rotation of the rotor 22 is pressed by the swash plate 31 to recede toward the top dead center, low temperature low pressure steam thrust out of the expansion chamber 43 is discharged, via the fourth steam passage P4 of the rotor 22, the third steam passage P3 of the moving valve plate 74, the sliding faces 77, the arcuate fifth steam passage P5 of the stationary valve plate 73 and the sixth and seventh steam passages P6 and P7 of the valve body 72, into the steam discharge chamber 88, and supplied therefrom to a condenser.
When the oil pump 49 provided on the output shaft 32 is actuated along with the rotation of the rotor 22, oil sucked from the oil pan 21 via the oil pipe 52, the oil passage 95 b of the pump body 95 and the intake port 53 is discharged from the discharge port 54, and is supplied via the oil passage 95 c of the pump body 95, the oil passage 32 a of the output shaft 32, the annular groove 32 b of the output shaft 32, the oil holes 32 c . . . of the output shaft 32, the annual groove 41 b of the cylinder sleeves 41 and the oil holes 41 c . . . of the cylinder sleeves 41 to a space between the smaller diameter part 62 b formed in the middle portion 62 of the piston 42 and the cylinder sleeves 41. Part of the oil held in the smaller diameter part 62 b flows through the spiral oil grooves 62 d . . . formed in the middle portion 62 of the piston 42 to lubricate the sliding face in contact with the cylinder sleeve 41, and another part of the oil lubricates the sliding faces of the compression rings 66 and the oil rings 67 provided on the top portions 63 of the piston 42 and of the cylinder sleeve 41.
It is inevitable for water generated by the condensation of part of the supplied high temperature high pressure steam to infiltrate from the expansion chambers 43 onto the sliding faces of the cylinder sleeves 41 and the pistons 42 to be mixed with oil. Therefore, the conditions of lubrication of the sliding faces are poor, but a sufficient oil film can be maintained to secure the required lubricating performance by supplying the required quantity of oil from the oil pump 49 through the inside of the output shaft 32 directly to the sliding faces of the cylinder sleeves 41 and the pistons 42. The size of the oil pump 49 can be therefore reduced.
The oil scraped off the sliding faces of the cylinder sleeves 41 and the pistons 42 by the oil ring 67 flows from the oil holes 63 c . . . formed in the bottom of the oil ring groove 63 b to the hollow spaces 62 a within the pistons 42. The hollow spaces 62 a communicate with the inside of the cylinder sleeves 41 via the plurality of oil holes 62 c . . . penetrating the middle portion 62 of each piston 42, and the inside of the cylinder sleeves 41 communicates via the plurality of oil holes 41 c . . . with the annual groove 41 b in the outer circumferences of the cylinder sleeves 41. Although the circumference of the annual groove 41 b is covered by the sleeve supporting flange 34 in the middle of the rotor 22, oil within the hollow spaces 62 a in the pistons 42 is urged outward in the radial direction by a centrifugal force, and discharged into the space 68 within the thermally insulating cover 40 through the oil holes 34 b in the sleeve supporting flange 34, because the oil holes 34 b are formed in the sleeve supporting flange 34, and the oil is then returned therefrom to the oil pan 21 through the oil holes 40 a . . . in the thermally insulating cover 40. Since the oil holes 34 b are in positions deviating farther than the outer end of the sleeve supporting flange 34 in the radial direction toward the axis L, the oil positioned outward from the oil holes 34 b in the radial direction is held by a centrifugal force in the hollow spaces 62 a of the pistons 42.
As described above, the oil held in the hollow spaces 62 a within the pistons 42 and the oil held in the smaller diameter part 62 b on the outer circumference of the pistons 42 are supplied from the smaller diameter part 62 b toward the top portion 63 in the expansion stroke in which the capacities of the expansion chambers 43 increase, and they are supplied from the smaller diameter part 62 b toward the end portion 61 in the compression stroke in which the capacities of the expansion chambers 43 decrease, thereby reliably lubricating the whole area of the pistons 42 in the axial direction. Moreover, the flow of oil within the hollow space 62 a of the pistons 42 enables the heat of the top portion 63 exposed to high temperature high pressure steam to be transmitted to the low temperature end portion 61, thereby avoiding a local temperature rise in the pistons 42.
When high temperature high pressure steam is supplied from the fourth steam passages P4 to the expansion chambers 43, the thermally insulating space 65 is formed between the top portion 63 and the middle portion 62 of each piston 42 facing the expansion chambers 43, and the thermally insulating space 70 is also formed in the rotor head 38 facing the expansion chambers 43. Therefore, the escape of heat from the expansion chambers 43 to the pistons 42 and the rotor head 38 can be minimized to contribute to improvement in the performance of the expander E. Furthermore, as the large capacity hollow space 62 a is formed within each piston 42, not only can the weight of the piston 42 be reduced but also can the thermal mass of the piston 42 be curtailed for a more effective suppression of the escape of heat from the expansion chambers 43.
As the metal gasket 36 is disposed between the rear sleeve supporting flange 35 and the rotor head 38 to seal the expansion chambers 43, the dead volume around the seals can be reduced as compared with a case in which the expansion chambers 43 are sealed by thick annular sealing members, thereby securing a large volume ratio (expansion ratio) for the expander E and enhancing the thermal efficiency to increase the output. Further, as the cylinder sleeves 41 are configured as separate members from the rotor 22, the material of the cylinder sleeves 41 can be selected in consideration of thermal conductivity, thermal resistance, strength, wear resistance and the like, without being restricted by the material of the rotor 22. Furthermore, only the worn or damaged cylinder sleeve 41 needs to be replaced, resulting in an improved economy.
Moreover, because the outer circumferential faces of the cylinder sleeves 41 are exposed through the two notches 57 and 58 formed in the outer circumferential face of the rotor 22 in the circumferential direction, not only can the weight of the rotor 22 be reduced but also can the thermal mass of the rotor 22 be curtailed to enhance thermal efficiency. Moreover, by causing the notches 57 and 58 to function as thermally insulating spaces, the escape of heat from the cylinder sleeves 41 can be suppressed. Furthermore, as the outer circumference of the rotor 22 is covered with the thermally insulating cover 40, the escape of heat from the cylinder sleeves 41 can be suppressed even more effectively.
As the rotary valve 71 supplies and discharges steam to and from the group of axial piston cylinders 56 via the flat sliding faces 77 between the stationary valve plate 73 and the moving valve plate 74, the leakage of steam can be effectively prevented, because the flat sliding faces 77 can be readily machined with high accuracy and permit easier control of clearances than cylindrical sliding faces do. Moreover, as preset loads are given to the valve body 72 by the plurality of preload springs 85 . . . to generate surface stresses on the sliding faces 77 of the stationary valve plate 73 and the moving valve plate 74, steam leaks from the sliding faces 77 can be suppressed even more effectively.
Further, as the valve body 72 of the rotary valve 71 is made of stainless steel providing a larger thermal expansion amount, and the stationary valve plate 73 fixed to the valve body 72 is made of carbon or ceramic providing a smaller thermal expansion amount, there is a possibility that the centering between them is displaced due to the difference in thermal expansion. However, as the fixed ring 79 is fixed to the valve body 72 with the plurality of bolts 80 . . . in a state in which the stepped portion 79 a on the inner circumference of the fixed ring 79 is pressed in and spigot-fitted onto the outer circumference of the stationary valve plate 73 and the stepped portion 79 b on the outer circumference of the fixed ring 79 is spigot-fitted onto the outer circumference of the valve body 72, it is possible to precisely center the stationary valve plate 73 relative to the valve body 72 by virtue of the aligning effect of spigot fitting, thereby preventing the expander E from deteriorating in performance by keeping the supply and discharge of steam in time. Moreover, the contact faces of the stationary valve plate 73 and the valve body 72 can be uniformly brought into close contact with each other with the fastening force of the bolts 80 . . . , thereby suppressing steam leakage from those contact faces.
Furthermore, since the rotary valve 71 can be attached to or detached from the casing body 12 by merely removing the rear cover 18 from the casing body 12, maintenance including repairs, cleaning and replacement can be significantly facilitated. Also, though the rotary valve 71 through which high temperature high pressure steam passes is increased in temperature, oil can be prevented from being heated by the high temperature of the rotary valve 71 to deteriorate the lubricating performance of the swash plate 31 and the output shaft 32 because the swash plate 31 and the output shaft 32 which require lubrication with oil are arranged on the other side of the rotor 22 than the rotary valve 71. The oil also performs the function to prevent overheating by cooling the rotary valve 71.
When assembling the expander E, it is necessary to adjust the magnitude of the dead volume between the bottom of the cylinder sleeves 41 (i.e., the lid member 69 supported by the rotor head 38) and the top of the pistons 42, namely the capacities of the expansion chambers 43 when the pistons 42 are at the top dead center. If the shim 97 intervening between the flange 32 d of the output shaft 32 and the inner races of the combined angular bearings 23 f and 23 r is thinned, the output shaft 32 will move forward (rightward in FIG. 1), resulting in a rightward shift of the rotor head 38 as well, but the dead volume will decrease because the pistons 42 are restricted by the swash plate 31 to be unable to move forward. Conversely, if the shim 97 is thickened, the rotor head 38 will move backward (leftward in FIG. 1) together with the output shaft 32, and accordingly the dead volume will increase. As a result, it is possible to adjust the dead volume as desired by merely replacing the shim 97, and the step otherwise needed for dead volume adjustment can be eliminated to achieve a substantial time saving.
Further, as a single shim 97 having a predetermined thickness is sandwiched between the flange 32 d of the output shaft 32 and the combined angular bearings 23 f and 23 r, to adjust the dead volume only by fastening with a single nut 98 the front cover 15 incorporating the angular bearing 30 supporting the swash plate 31 and the combined angular bearings 23 f and 23 r supporting the rotor 22 and the rotor 22 incorporating the pistons 42 . . . , the adjustment procedure is simplified as compared with the conventional adjustment procedure in which the thicknesses of two shims, front and rear, are individually adjusted. Moreover, since the rotor 22 incorporating the pistons 42 . . . can be kept assembled into the casing body 12 when adjusting the dead volume, the adjusted dead volume can be confirmed while directly watching the state of contact between the pistons 42 . . . and the swash plate 31.
When the position of the output shaft 32 relative to the combined angular bearings 23 f and 23 r is adjusted back and forth by varying the thickness of the shim 97, the position of the rotor head 38 at the rear end of the rotor 22 also shifts back and forth, but there is no problem in adjusting the position of the output shaft 32 because the rotor head 38 is slidable in the direction of the axis L relative to the inner race of the radial bearing 24 disposed between it and the casing body 12.
Then, when the pressure of high temperature high pressure steam supplied to the expansion chambers 43 urges the pistons 42 in the direction of being thrust out of the cylinder sleeves 41, the pressing force of the pistons 42 presses forward (rightward in FIG. 1) the outer race of the combined angular bearings 23 f and 23 r via the swash plate 31, the angular bearing 30, the swash plate holder 28 and the front cover 15, and the pressing force of the cylinder sleeves 41 reverse in direction to the suppressing force of the pistons 42 presses backward (leftward in FIG. 1) the inner race of the combined angular bearings 23 f and 23 r via the rotor head 38 and the output shaft 32. Thus, the load generated by the high temperature high pressure steam supplied to the expansion chambers 43 is cancelled within the combined angular bearings 23 f and 23 r, without being transmitted to the casing body 12.
While the rotor 22 constructed of the output shaft 32, the three sleeve supporting flanges 33, 34 and 35, the rotor head 38 and the thermally insulating cover 40 is made of a ferrous material whose thermal expansion is relatively small, the casing 11 which supports the rotor 22 via the combined angular bearings 23 f and 23 r and the radial bearing 24 is made of an aluminum-based material whose thermal expansion is relatively large. As a consequence, there arises a difference in the quantity of thermal expansion in the direction along the axis L between the high and low temperatures of the expander E.
The casing 11 which is greater in thermal expansion than the rotor 22 expands more than the rotor 22 and its size in relatively increases in the direction of the axis L when the temperature is high. Conversely, when the temperature is low, it contracts more and its size relatively decreases in the direction of the axis L. As the casing 11 and the rotor 22 are positioned in the direction of the axis L via the combined angular bearings 23 f and 23 r, the difference in thermal expansion between them is absorbed by the sliding contact of the rotor head 38 with the inner race of the radial bearing 24, so that an excessive load is prevented from acting in the direction of the axis L on the combined angular bearings 23 f and 23 r, the radial bearing 24 and the rotor 22. This not only contributes to an increase in the durability of the combined angular bearings 23 f and 23 r and of the radial bearing 24, but also to stabilization in support of the rotor 22, thereby facilitating its smooth rotation. Moreover, it is possible to prevent the fluctuation in dead volume between the top of the cylinder sleeves 41 and the top of the pistons 42 accompanying the change in temperature.
The reason is that, supposing that both ends of the rotor 22 are restrained by the casing 11 to be immovable in the axial direction, as the casing 11 tends to contract in the direction of the axis L relative to the rotor 22 when the temperature is low, the pistons 42 whose top is in contact with the swash plate 31 supported by the swash plate holder 28 which is part of the casing 11, are pressed backward, and the rotor head 38 supported by the casing 11 via the radial bearing 24 is pressed forward, so that the pistons 42 are pressed into the cylinder sleeves 41 and the dead volume decreases accordingly. Conversely, when the temperature is high, as the casing 11 tends to extend in the direction of the axis L relative to the rotor 22, the pistons 42 are drawn out from the inside of the cylinder sleeves 41, resulting in an increase in dead volume, which in turn invites an increase in the initial volume of high temperature high pressure steam in the normal operating state after the warming-up, i.e. a drop in thermal efficiency due to a decrease in the volume ratio (expansion ratio) of the expander E.
By contrast, in this embodiment of the invention, as the rotor 22 is supported in a floating state in the direction of the axis L relative to the casing 11, the gaps between the combined angular bearings 23 f and 23 r and the radial bearing 24 are prevented from widening and so are the preloads from decreasing, and the dead volume is prevent from fluctuating due to temperature change. This enables the volume ratio (expansion ratio) of the expander E to be prevented from fluctuating, thereby achieving a stable performance.
Especially, for the expander E which uses high temperature high pressure steam as the working medium, the above-described advantage is highly effective because the difference is wide between high temperature and low temperature. Furthermore, although the difference between high temperature and low temperature is particularly wide in the vicinity of the rotary valve 71 to which high temperature high pressure steam is supplied, the difference in thermal expansion between the casing 11 and the rotor 22 can be absorbed without problem because the rotor head 38 can be in sliding contact in the direction of the axis L with the radial bearing 24 arranged closer to the rotary valve 71.
Further, out of the stationary valve plate 73 and the moving valve plate 74 of the rotary valve 71, as the stationary valve plate 73 supported by the casing 11 is urged by the springing force of the preload springs 85 . . . toward the moving valve plate 74 supported by the rotor 22, the sealing performance of the sliding faces 77 of the stationary valve plate 73 and the moving valve plate 74 will not be affected even if the positional relationship between the casing 11 and the rotor 22 in the direction of the axis L varies along with temperature variations. Not only that, an excessive load is prevented from acting on the combined angular bearings 23 f and 23 r and the radial bearing 24, resulting in stabilization of the rotational plane of the rotor 22 and accordingly in an improvement in the sealing performance of the sliding faces 77, to reduce the quantity of leaked steam.
Although the preferred embodiment of the present invention has been described above, the invention may be modified in various ways without deviating from the subject matter.
For example, the rotating fluid machine according to the invention is not limited to the application to the expander E, and is also applicable to a compressor, a hydraulic pump, a hydraulic motor and the like.
Although the expander E in the embodiment is provided with the group of axial piston cylinders 56 as the working section, the structure of the working section is not limited thereto.

Claims

1. A rotating fluid machine comprising:

a casing;

a rotor rotatably supported by the casing;

a working section disposed on the rotor; and

a rotary valve, provided between the casing and the rotor, for controlling the supply and discharge of a working medium to and from the working section,

the rotary valve being constructed by bringing, into contact on sliding faces, a moving valve plate provided on the rotor and a stationary valve plate provided on a valve body, and high pressure working medium passages and low pressure working medium passages penetrating mating faces of the valve body and the stationary valve plate,

wherein the outer circumference of the mating faces is sealed with an annular first sealing member and the circumference of the low pressure working medium passages is sealed with an annular second sealing member inside the first sealing member in the radial direction.

2. The rotating fluid machine according to claim 1, wherein the first and second sealing members are C-shaped seals each having an opening in its section, the opening of the first sealing member is arranged inward in the radial direction, and the opening of the second sealing member is arranged outward in the radial direction.