U.S. patent application number 10/489641 was filed with the patent office on 2005-01-27 for rotary fluid machine.
Invention is credited to Endoh, Tsuneo, Ichikawa, Hiroshi, Kimura, Yasunari, Takahashi, Tsutomu.
Application Number | 20050019161 10/489641 |
Document ID | / |
Family ID | 19111885 |
Filed Date | 2005-01-27 |
United States Patent
Application |
20050019161 |
Kind Code |
A1 |
Ichikawa, Hiroshi ; et
al. |
January 27, 2005 |
Rotary fluid machine
Abstract
A rotary fluid machine is provided that includes a rotor
rotatably housed within a casing, a hollow rotating shaft (113)
that rotates integrally with the rotor, and a fixed shaft (102)
that is relatively rotatably fitted into the inner periphery of the
rotating shaft (113); wherein the fixed shaft (102) is floatingly
supported in the casing via a fixed shaft support spring (95)
having an alignment action. When rotational runout of the rotor is
transmitted to the fixed shaft (102) via the rotating shaft (113),
the alignment action of the fixed shaft support spring (95)
suppresses the rotational runout of the rotor, and any increase in
the frictional resistance in a sliding section between the rotating
shaft (113) and the fixed shaft (102) and the occurrence of
abnormal wear can be prevented effectively, and the leakage of
steam from a rotary valve (V) provided on the sliding surfaces of
the rotating shaft (113) and the fixed shaft (102) can be
reduced.
Inventors: |
Ichikawa, Hiroshi; (Saitama,
JP) ; Endoh, Tsuneo; (Saitama, JP) ;
Takahashi, Tsutomu; (Saitama, JP) ; Kimura,
Yasunari; (Saitama, JP) |
Correspondence
Address: |
BIRCH STEWART KOLASCH & BIRCH
PO BOX 747
FALLS CHURCH
VA
22040-0747
US
|
Family ID: |
19111885 |
Appl. No.: |
10/489641 |
Filed: |
September 22, 2004 |
PCT Filed: |
September 20, 2002 |
PCT NO: |
PCT/JP02/09719 |
Current U.S.
Class: |
415/216.1 |
Current CPC
Class: |
F02B 53/00 20130101;
F04C 2230/602 20130101; F04C 2240/601 20130101; F01B 13/068
20130101; F01C 11/006 20130101; F04B 1/047 20130101; F01B 13/061
20130101; F01C 21/08 20130101; F01C 21/18 20130101; F04C 2240/605
20130101; F01C 1/3446 20130101 |
Class at
Publication: |
415/216.1 |
International
Class: |
F01D 025/00 |
Foreign Application Data
Date |
Code |
Application Number |
Sep 21, 2001 |
JP |
2001-289387 |
Claims
What is claimed is:
1. A rotary fluid machine comprising a rotor (41) rotatably housed
within a casing (11), a hollow rotating shaft (113) that rotates
integrally with the rotor (41), and a fixed shaft (102) that is
relatively rotatably fitted into the inner periphery of the
rotating shaft (113), characterized in that the fixed shaft (102)
is floatingly supported in the casing (11) via resilient support
means (95) having an alignment action.
2. The rotary fluid machine according to claim 1, wherein a rotary
valve (V) for controlling supplying and discharging of a high
temperature gas-phase working medium is provided on sliding
surfaces of the rotating shaft (113) and the fixed shaft (102).
Description
FIELD OF THE INVENTION
[0001] The present invention relates to a rotary fluid machine for
interconverting the pressure energy of a gas-phase working medium
and the rotational energy of a rotor.
BACKGROUND ART
[0002] A rotary fluid machine disclosed in Japanese Patent
Application Laid-open No. 2000-320543 is equipped with a vane
piston unit in which a vane and a piston are combined; the piston,
which is slidably fitted in a cylinder provided radially in a
rotor, interconverts the pressure energy of a gas-phase working
medium and the rotational energy of the rotor via a power
conversion device comprising an annular channel and a roller, and
the vane, which is radially and slidably supported in the rotor,
interconverts the pressure energy of the gas-phase working medium
and the rotational energy of the rotor.
[0003] In such a rotary fluid machine, a rotary valve for supplying
and discharging a high temperature gas-phase working medium is
formed between the outer peripheral face of a fixed shaft fixed to
a casing and the inner peripheral face of a hollow rotating shaft
by fitting and rotatably supporting the rotating shaft, which
rotates integrally with the rotor, on the outer periphery of the
fixed shaft.
[0004] In order to maintain the sealing characteristics for the
gas-phase working medium in the rotary valve, since it is a mating
seal, it is necessary to precisely control the clearance between
the sliding surfaces of the fixed shaft and the rotating shaft.
However, since it is impossible to avoid the occurrence of some
degree of runout in the rotating rotor, if the above clearance is
set so as to be small, the frictional resistance between the
sliding surfaces of the fixed shaft and the rotating shaft is
higher, and there is the problem of interference with the rotation
of the rotor. Furthermore, if the clearance between the sliding
surfaces of the fixed shaft and the rotating shaft is set so as to
be appropriate when they are cold, the outer peripheral face of the
fixed shaft is worn due to a difference in thermal expansion in the
vicinity of the rotary valve where the high temperature gas-phase
working medium passes through, and since contact with the outer
peripheral face of the fixed shaft is uneven due to rotational
runout of the rotor, resulting in eccentric wear, there are the
problems of a degradation in the sealing characteristics for the
gas-phase working medium, an increase in the sliding resistance,
and degradation of the rotational behavior of the rotor.
DISCLOSURE OF THE INVENTION
[0005] The present invention has been accomplished under the
above-mentioned circumstances, and an object thereof is to lessen
the influence of rotational runout of a rotor of a rotary fluid
machine when a hollow rotating shaft provided integrally with the
rotor is rotatably supported on the outer periphery of a fixed
shaft fixed to a casing.
[0006] In order to achieve the above object, in accordance with a
first aspect of the present invention, there is proposed a rotary
fluid machine that includes a rotor rotatably housed within a
casing, a hollow rotating shaft that rotates integrally with the
rotor, and a fixed shaft that is relatively rotatably fitted into
the inner periphery of the rotating shaft, characterized in that
the fixed shaft is floatingly supported in the casing via resilient
support means having an alignment action.
[0007] In accordance with this arrangement, since the fixed shaft
is floatingly supported in the casing via the resilient support
means having the alignment action, that is, it is connected with
low rigidity and flexibly supported, and the hollow rotating shaft
is supported on the outer periphery of the fixed shaft, when
rotational runout of the rotor is transmitted to the fixed shaft
via the rotating shaft, the rotational runout of the rotor can be
suppressed by the alignment action of the resilient support means.
The tracking ability of seal surfaces can thereby be improved, and
the sealing characteristics can be enhanced by controlling the
clearance with high precision, thus avoiding effectively any
increase in the frictional resistance and any abnormal wear in a
sliding section between the rotating shaft and the fixed shaft.
[0008] Furthermore, in accordance with a second aspect of the
present invention, in addition to the first aspect, there is
proposed a rotary fluid machine wherein a rotary valve for
controlling supplying and discharging of a high temperature
gas-phase working medium is provided on sliding surfaces of the
rotating shaft and the fixed shaft.
[0009] In accordance with this arrangement, since the rotary valve
for controlling supplying and discharging of the high temperature
gas-phase working medium is provided on the sliding surfaces of the
rotating shaft and the fixed shaft, even when the outer peripheral
face of the fixed shaft is worn due to a difference in thermal
expansion between the fixed shaft and the rotating shaft,
rotational runout of the rotor can be suppressed by the alignment
action of the resilient support means, the amount of wear of the
outer peripheral face of the fixed shaft thus becomes uniform, and
it is therefore possible to precisely control the clearance between
the sliding surfaces of the fixed shaft and the rotating shaft when
they are hot. Moreover, once uniform contact when hot is formed by
the initial setting of the rotary fluid machine, the sealing
characteristics can always be maintained for subsequent
introduction of the gas-phase working medium, and the sealing
characteristics for the gas-phase working medium can be ensured by
maintaining a small and uniform clearance.
[0010] A fixed shaft support spring 95 of embodiments corresponds
to the resilient support means of the present invention, and steam
in the embodiments corresponds to the gas-phase working medium of
the present invention.
BRIEF DESCRIPTION OF THE DRAWINGS
[0011] FIG. 1 to FIG. 21D illustrate a first embodiment of the
present invention; FIG. 1 is a schematic view of a waste heat
recovery system of an internal combustion engine; FIG. 2 is a
longitudinal sectional view of an expander, corresponding a
sectional view along line 2-2 of FIG. 4; FIG. 3 is an enlarged
sectional view around the axis of FIG. 2; FIG. 4 is a sectional
view along line 4-4 of FIG. 2; FIG. 5 is a sectional view along
line 5-5 of FIG. 2; FIG. 6 is a sectional view along line 6-6 of
FIG. 2; FIG. 7 is a sectional view along line 7-7 of FIG. 5; FIG. 8
is a sectional view along line 8-8 of FIG. 5; FIG. 9 is a sectional
view along line 9-9 of FIG. 8; FIG. 10 is a sectional view along
line 10-10 of FIG. 3; FIG. 11 is an exploded perspective view of a
rotor; FIG. 12 is an exploded perspective view of a lubricating
water distribution section of the rotor; FIG. 13 is a schematic
view showing cross-sectional shapes of a rotor chamber and the
rotor; FIG. 14 is an enlarged view of an essential part of FIG. 3,
showing a rotary valve and a fixed shaft support spring; FIG. 15 is
an enlarged view of an essential part of FIG. 2, showing the outer
peripheral face of the fixed shaft; FIG. 16 is a sectional view
along line 16-16 of FIG. 14; FIG. 17A is an enlarged view of an
essential part of a first fixed shaft; FIG. 17B is a sectional view
along line 17B-17B of FIG. 17A; FIG. 18A is an enlarged view of a
nozzle member; FIG. 18B is a sectional view along line 18B-18B of
FIG. 18A; FIG. 19 is a sectional view along line 19-19 of FIG. 14;
FIG. 20A to FIG. 20D are diagrams for explaining the operation when
a fixed sleeve is shrink-fitted; and FIG. 21A to FIG. 21D are
graphs showing relationships between the thermal expansion of the
fixed shaft and that of the rotating shaft. FIG. 22 and FIG. 23
illustrate a second embodiment of the present invention; FIG. 22 is
a view corresponding to FIG. 14; and FIG. 23 is a sectional view
along line 23-23 of FIG. 22.
BEST MODE FOR CARRYING OUT THE INVENTION
[0012] A first embodiment of the present invention is explained
below with reference to FIG. 1 to FIG. 21D.
[0013] In FIG. 1, a waste heat recovery system 2 for an internal
combustion engine 1 includes an evaporator 3 that generates high
temperature, high pressure steam by vaporizing a high pressure
liquid (e.g. water) using as a heat source the waste heat (e.g.
exhaust gas) of the internal combustion engine 1, an expander 4
that generates an output by expansion of the steam, a condenser 5
that liquefies steam having decreased temperature and pressure as a
result of conversion of the pressure energy into mechanical energy
in the expander 4, and a supply pump 6 that pressurizes the liquid
(e.g. water) from the condenser 5 and resupplies it to the
evaporator 3.
[0014] As shown in FIG. 2 and FIG. 3, a casing 11 of the expander 4
is formed from first and second casing halves 12 and 13, which are
made of metal. The first and second casing halves 12 and 13 are
formed from main body portions 12a and 13a, which in cooperation
form a rotor chamber 14, and circular flanges 12b and 13b, which
are joined integrally to the outer peripheries of the main body
portions 12a and 13a, and the two circular flanges 12b and 13b are
joined together via a metal gasket 15. The outer face of the first
casing half 12 is covered with a transit chamber outer wall 16
having a deep bowl shape, and a circular flange 16a, which is
joined integrally to the outer periphery of the transit chamber
outer wall 16, is superimposed on the left face of the circular
flange 12b of the first casing half 12. The outer face of the
second casing half 13 is covered with an exhaust chamber outer wall
17 for housing a magnet coupling (not illustrated) for transmitting
the output of the expander 4 to the outside, and a circular flange
17a, which is joined integrally to the outer periphery of the
exhaust chamber outer wall 17, is superimposed on the right face of
the circular flange 13b of the second casing half 13. The
above-mentioned four circular flanges 12b, 13b, 16a, and 17a are
tightened together by means of a plurality of bolts 18 disposed in
the circumferential direction. A transit chamber 19 is defined
between the transit chamber outer wall 16 and the first casing half
12, and an exhaust chamber 20 is defined between the exhaust
chamber outer wall 17 and the second casing half 13. The exhaust
chamber outer wall 17 is provided with an outlet (not illustrated)
for guiding the decreased temperature, decreased pressure steam
that has finished work in the expander 4 to the condenser 5.
[0015] The main body portions 12a and 13a of the two casing halves
12 and 13 have hollow bearing tubes 12c and 13c projecting outward
in the lateral direction, and an outer sleeve 21 having a hollow
portion 21a is rotatably supported by these hollow bearing tubes
12c and 13c via a pair of bearing members 22 and 23. The axis L of
the outer sleeve 21 thus passes through the intersection of the
major axis and the minor axis of the rotor chamber 14, which has a
substantially elliptical shape. The outer sleeve 21, which is made
of metal, forms a rotating shaft 113 in cooperation with a ceramic
inner sleeve 85, which will be described later.
[0016] A seal block 25 is housed within a lubricating water supply
member 24 screwed onto the right-hand end of the second casing half
13, and secured by a nut 26. A small diameter portion 21b at the
right-hand end of the outer sleeve 21 is supported within the seal
block 25, a pair of seals 27 are disposed between the seal block 25
and the small diameter portion 21b, a pair of seals 28 are disposed
between the seal block 25 and the lubricating water supply member
24, and a seal 29 is disposed between the lubricating water supply
member 24 and the second casing half 13. A filter 30 is fitted in a
recess formed in the outer periphery of the hollow bearing tube 13c
of the second casing half 13, and is prevented from falling out by
means of a filter cap 31 screwed into the second casing half 13. A
pair of seals 32 and 33 are provided between the filter cap 31 and
the second casing half 13.
[0017] As is clear from FIG. 4 and FIG. 13, a circular rotor 41 is
rotatably housed within the rotor chamber 14, which has a
pseudo-elliptical shape. The rotor 41 is fitted onto and joined
integrally to the outer periphery of the outer sleeve 21, and the
axis of the rotor 41 and the axis of the rotor chamber 14 coincide
with the axis L of the outer sleeve 21. The shape of the rotor
chamber 14 viewed in the axis L direction is pseudo-elliptical, and
is similar to a rhombus having four rounded corners, the shape
having a major axis DL and a minor axis DS. The shape of the rotor
41 viewed in the axis L direction is a perfect circle having a
diameter DR that is slightly smaller than the minor axis DS of the
rotor chamber 14.
[0018] The cross-sectional shapes of the rotor chamber 14 and the
rotor 41 viewed in a direction orthogonal to the axis L are all
racetrack-shaped. That is, the cross-sectional shape of the rotor
chamber 14 is formed from a pair of flat faces 14a extending
parallel to each other at a distance d, and arc-shaped faces 14b
having a central angle of 180.degree. that are smoothly connected
to the outer peripheries of the flat faces 14a and, similarly, the
cross-sectional shape of the rotor 41 is formed from a pair of flat
faces 41a extending parallel to each other at the distance d, and
arc-shaped faces 41b having a central angle of 180.degree. that are
smoothly connected to the outer peripheries of the flat faces 41a.
The flat faces 14a of the rotor chamber 14 and the flat faces 41a
of the rotor 41 are in contact with each other, and a pair of
crescent-shaped spaces are formed between the inner peripheral face
of the rotor chamber 14 and the outer peripheral face of the rotor
41 (see FIG. 4).
[0019] The structure of the rotor 41 is now explained in detail
with reference to FIG. 3 to FIG. 6, and FIG. 11.
[0020] The rotor 41 is formed from a rotor core 42 that is formed
integrally with the outer periphery of the outer sleeve 21, and
twelve rotor segments 43 that are fixed so as to cover the
periphery of the rotor core 42 and form the outer shell of the
rotor 41. Twelve ceramic (or carbon) cylinders 44 are mounted
radially in the rotor core 42 at 30.degree. intervals and fastened
by means of clips 45 to prevent them falling out. A small diameter
portion 44a is projectingly provided at the inner end of each of
the cylinders 44, and a gap between the base end of the small
diameter portion 44a and the inner sleeve 85 is sealed via a C seal
46. The extremity of the small diameter portion 44a is fitted into
the outer peripheral face of the hollow inner sleeve 85, and a
cylinder bore 44b communicates with first and second steam passages
S1 and S2 within a fixed shaft 102 via twelve third steam passages
S3 running through the small diameter portion 44a and the rotating
shaft 113. A ceramic piston 47 is slidably fitted within each of
the cylinders 44. When the piston 47 moves to the radially
innermost position, it retracts completely within the cylinder bore
44b, and when it moves to the radially outermost position, about
half of the whole length projects outside the cylinder-bore
44b.
[0021] Each of the rotor segments 43 is a hollow wedge-shaped
member having a central angle of 30.degree., and has two recesses
43a and 43b formed on the faces thereof that are opposite the pair
of flat faces 14a of the rotor chamber 14, the recesses 43a and 43b
extending in an arc shape with the axis L as the center, and
lubricating water outlets 43c and 43d open in the middle of the
recesses 43a and 43b. Furthermore, four lubricating water outlets
43e and 43f open on the end faces of the rotor segments 43, that
is, the faces that are opposite vanes 48, which will be described
later.
[0022] The rotor 41 is assembled as follows. The twelve rotor
segments 43 are fitted around the outer periphery of the rotor core
42, which is preassembled with the cylinders 44, the clips 45, and
the C seals 46, and the vanes 48 are fitted in twelve vane channels
49 formed between adjacent rotor segments 43. At this point, in
order to form a predetermined clearance between the vanes 48 and
the rotor segments 43, shims having a predetermined thickness are
disposed on opposite faces of the vanes 48. In this state, the
rotor segments 43 and the vanes 48 are tightened inward in the
radial direction toward the rotor core 42 by means of a jig so as
to precisely position the rotor segments 43 relative to the rotor
core 42, and each of the rotor segments 43 is then provisionally
retained on the rotor core 42 by means of provisional retention
bolts 50 (see FIG. 8). Subsequently each of the rotor segments 43
and the rotor core 42 are co-machined so as to make two knock pin
holes 51 run therethrough, and four knock pins 52 are press-fitted
in the two knock pin holes 51 so as to join each of the rotor
segments 43 to the rotor core 42.
[0023] As is clear from FIG. 8, FIG. 9, and FIG. 12, a through hole
53 running through the rotor segment 43 and the rotor core 42 is
formed between the two knock pin holes 51, and recesses 54 are
formed at opposite ends of the through hole 53. Two pipe members 55
and 56 are fitted within the through hole 53 via seals 57 to 60,
and an orifice-forming plate 61 and a lubricating water
distribution member 62 are fitted into each of the recesses 54 and
secured by a nut 63. The orifice-forming plate 61 and the
lubricating water distribution member 62 are prevented from
rotating relative to the rotor segments 43 by two knock pins 64
running through knock pin holes 61a of the orifice-forming plate 61
and fitted into knock pin holes 62a of the lubricating water
distribution member 62, and a gap between the lubricating water
distribution member 62 and the nut 63 is sealed by an O ring
65.
[0024] A small diameter portion 55a formed in an outer end portion
of one of the pipe members 55 communicates with a sixth water
passage W6 within the pipe member 55 via a through hole 55b, and
the small diameter portion 55a also communicates with a radial
distribution channel 62b formed on one side face of the lubricating
water distribution member 62. The distribution channel 62b of the
lubricating water distribution member 62 extends in six directions,
and the extremities thereof communicate with six orifices 61b, 61c,
and 61d of the orifice-forming plate 61. The structures of the
orifice-forming plate 61, the lubricating water distribution member
62 and the nut 63 provided at the outer end portion of the other
pipe member 56 are identical to the structures of the
above-mentioned orifice-forming plate 61, lubricating water
distribution member 62, and nut 63.
[0025] Downstream sides of the two orifices 61b of the
orifice-forming plate 61 communicate with the two lubricating water
outlets 43e, which open so as to be opposite the vane 48, via
seventh water passages W7 formed within the rotor segments 43;
downstream sides of the two orifices 61c communicate with the two
lubricating water outlets 43f, which open so as to be opposite the
vane 48, via eighth water passages W8 formed within the rotor
segment 43; and downstream sides of the two orifices 61d
communicate with the two lubricating water outlets 43c and 43d,
which open so as to be opposite the rotor chamber 14, via ninth
water passages W9 formed within the rotor segment 43.
[0026] As is clear from reference in addition to FIG. 5, an annular
channel 67 is defined by a pair of O rings 66 on the outer
periphery of the cylinder 44, and the sixth water passage W6 formed
within said one of the pipe members 55 communicates with the
annular channel 67 via four through holes 55c running through the
pipe member 55 and a tenth water passage W10 formed within the
rotor core 42. The annular channel 67 communicates with sliding
surfaces of the cylinder bore 44b and the piston 47 via an orifice
44c. The position of the orifice 44c of the cylinder 44 is set so
that it stays within the sliding surface of the piston 47 when the
piston 47 moves between top dead center and bottom dead center.
[0027] As is clear from FIG. 3 and FIG. 9, the first water passage
W1 formed in the lubricating water supply member 24 communicates
with the small diameter portion 55a of said one of the pipe members
55 via a second water passage W2 formed in the seal block 25, third
water passages W3 formed in the small diameter portion 21b of the
outer sleeve 21, an annular channel 68a formed in the outer
periphery of a water passage forming member 68 fitted in the center
of the outer sleeve 21, a fourth water passage W4 formed in the
outer sleeve 21, a pipe member 69 bridging the rotor core 42 and
the rotor segments 43, and fifth water passages W5 formed so as to
bypass the knock pin 52 on the radially inner side of the rotor
segment 43.
[0028] As shown in FIG. 7, FIG. 9, and FIG. 11, twelve vane
channels 49 are formed between adjacent rotor segments 43 of the
rotor 41 so as to extend in the radial direction, and the
plate-shaped vanes 48 are slidably fitted in the respective vane
channels 49. Each of the vanes 48 has a substantially U-shaped form
comprising parallel faces 48a following the parallel faces 14a of
the rotor chamber 14, an arc-shaped face 48b following the
arc-shaped face 14b of the rotor chamber 14, and a notch 48c
positioned between the parallel faces 48a. Rollers 71 having a
roller bearing structure are rotatably supported on a pair of
support shafts 48d projecting from the parallel faces 48a.
[0029] A U-shaped synthetic resin seal 72 is retained in the
arc-shaped face 48b of the vane 48, and the extremity of the seal
72 projects slightly from the arc-shaped face 48b of the vane 48
and comes into sliding contact with the arc-shaped face 14b of the
rotor chamber 14. Two recesses 48e are formed on each side of the
vane 48, and these recesses 48e are opposite the two radially inner
lubricating water outlets 43e that open on the end faces of the
rotor segment 43. A piston receiving member 73, which is provided
so as to project radially inward in the middle of the notch 48c of
the vane 48, abuts against the radially outer end of the piston
47.
[0030] As is clear from FIG. 4, two pseudo-elliptical annular
channels 74 having a similar shape to that of a rhombus with its 4
apexes rounded are provided in the flat faces 14a of the rotor
chamber 14 defined by the first and second casing halves 12 and 13,
and the pair of rollers 71 of each of the vanes 48 are rollably
engaged with these annular channels 74. The distance between these
annular channels 74 and the arc-shaped face 14b of the rotor
chamber 14 is constant throughout the whole circumference.
Therefore, when the rotor 41 rotates, the vane 48 having the
rollers 71 guided by the annular channels 74 reciprocates radially
within the vane channel 49, and the seal 72 mounted on the
arc-shaped face 48b of the vane 48 slides along the arc-shaped face
14b of the rotor chamber 14 with a constant amount of compression.
This enables direct physical contact between the rotor chamber 14
and the vanes 48 to be prevented and vane chambers 75 defined
between adjacent vanes 48 to be reliably sealed while preventing
any increase in the sliding resistance or the occurrence of
wear.
[0031] As is clear from FIG. 2, a pair of circular seal channels 76
are formed in the flat faces 14a of the rotor chamber 14 so as to
surround the outside of the annular channels 74. A pair of ring
seals 79 equipped with two O rings 77 and 78 are slidably fitted in
the circular seal channels 76, and the seal surfaces are opposite
the recesses 43a and 43b (see FIG. 4) formed in each of the rotor
segments 43. The pair of ring seals 79 are prevented from rotating
relative to the first and second casing halves 12 and 13 by knock
pins 80.
[0032] As is clear from FIG. 2, FIG. 3, FIG. 10, and FIG. 14, an
opening 16b is formed at the center of the transit chamber outer
wall 16; a boss portion 81a of a spring support member 81 and a
boss portion 82a of a fixed sleeve support member 82 disposed on
the axis L are tightened together to the inner face of the opening
16b by a plurality of bolts 83, and the fixed sleeve support member
82 is secured to the first casing half 12 by means of a nut 84. The
inner sleeve 85, which is formed in a cylindrical shape using a
material having a small coefficient of thermal expansion such as
ceramic, is fixed in the hollow portion 21a of the outer sleeve 21,
which is made of metal, by shrink-fitting, and a fixed sleeve 86 is
relatively rotatably fitted into the inner peripheral face of the
inner sleeve 85. The fixed sleeve 86 is formed from an inner sleeve
87 made of a material having small coefficient of thermal expansion
such as ceramic and an outer sleeve 88 made of metal, the outer
sleeve 88 being united with the outer periphery of the inner sleeve
87 by shrink-fitting, and the left-hand end of the fixed sleeve 86
is supported by the fixed sleeve support member 82 via an Oldham
coupling 89 that allows relative movement in the radial direction.
A gap between the fixed sleeve 86 and the first casing half 12 is
sealed by a seal 90 at a position close to the Oldham coupling
89.
[0033] Disposed within the hollow fixed sleeve 86 are a steam
supply pipe 91, a first fixed shaft 92, a second fixed shaft 93, a
third fixed shaft 94, and a fixed shaft support spring 95. The
steam supply pipe 91, which is disposed on the axis L, runs through
the boss portion 81a of the spring support member 81 and is secured
by a nut 97. The first fixed shaft 92 is a pipe-shaped member
having the right-hand end thereof closed, and the right-hand end of
the steam supply pipe 91 is fitted into an open portion at the
left-hand end of the first fixed shaft 92. The inner sleeve 87 of
the fixed sleeve 86 has a thick portion 87a projecting radially
inward, the second fixed shaft 93, which is a pipe-shaped member
having a central portion thereof closed, is held between the inner
periphery of the thick portion 87a and the outer periphery of the
first fixed shaft 92, and seals 98 and 99 are disposed between the
thick portion 87a of the inner sleeve 87 and the second fixed shaft
93. A threaded portion at the right-hand end of the second fixed
shaft 93 is screwed into the inner peripheral face of the third
fixed shaft 94, which is a pipe-shaped member having the right-hand
end thereof closed, and two seals 100 and 101 provided at the
right-hand end of the third fixed shaft 94 are in intimate contact
with the inner peripheral face of the inner sleeve 87 of the fixed
sleeve 86 and the inner peripheral face of the outer sleeve 21 of
the rotating shaft 113.
[0034] The fixed sleeve 86, the first fixed shaft 92, the second
fixed shaft 93, and the third fixed shaft 94 form the fixed shaft
102 of the present invention.
[0035] As is most clearly shown in FIG. 14 and FIG. 19, the fixed
shaft support spring 95 disposed around the outer periphery of the
steam supply pipe 91 provides a connection between a cylindrical
spring portion 81b forming a multicylindrical support portion
extending rightward from the boss portion 81a of the spring support
member 81 and a cylindrical spring portion 93a similarly forming a
multicylindrical support portion and extending leftward from the
central portion of the second fixed shaft 93. That is, the fixed
shaft support spring 95 comprises seven cylindrical springs 103a,
103b, and 103c; 104a, 104b, and 104c; and 105, which are arranged
concentrically with the axis L as the center; the three cylindrical
springs 103a, 103b, and 103c are fitted around the outer periphery
of the cylindrical spring portion 81b of the spring support member
81 so that there are gaps therebetween and are welded to each other
at the ends; the three cylindrical springs 104a, 104b, and 104c are
fitted around the outer periphery of the cylindrical spring portion
93a of the second fixed shaft 93 so that there are gaps
therebetween and are welded to each other at the ends; and opposite
ends of the cylindrical spring 105 on the outermost peripheral side
are welded to the cylindrical springs 103c and 104c, which are on
the inside thereof.
[0036] As is clear from FIG. 10 and FIG. 14, two collars 106 are
fitted around the second fixed shaft 93, which is sandwiched
between the first fixed shaft 92 and the inner sleeve 87, and two
nozzle members 107 are fitted in the thick portion 87a of the inner
sleeve 87. The first steam passage S1, which communicates with the
steam supply pipe 91, is formed in the center of the first fixed
shaft 92 in the axial direction, and the two second steam passages
S2, which pass through the interiors of the collars 106 and the
nozzle members 107, run radially through the first fixed shaft 92,
the second fixed shaft 93, and the fixed sleeve 86 with a phase
difference of 180.degree.. As described above, the twelve third
steam passages S3 run through the small diameter portions 44a of
the twelve cylinders 44 retained at intervals of 30.degree. in the
rotor 41 fixed to the rotating shaft 113 and the inner sleeve 85 of
the rotating shaft 113, and radially inner end portions of these
third steam passages S3 are opposite the radially outer end
portions of the second steam passages S2 so as to be able to
communicate therewith.
[0037] A pair of notches 86a are formed on the outer peripheral
face of the thick portion 87a of the fixed sleeve 86 with a phase
difference of 180.degree., and these notches can communicate with
the third steam passages S3. The notches 86a and the transit
chamber 19 communicate with each other via four fourth steam
passages S4 formed axially in the fixed sleeve 86, a fifth steam
passage S5 formed within the fixed sleeve 86 and the fixed sleeve
support member 82, and through holes 82b opening on the outer
periphery of the boss portion 82a of the fixed sleeve support
member 82.
[0038] As shown in FIG. 2 and FIG. 4, a plurality of radially
aligned intake ports 108 are formed in the first casing half 12 and
the second casing half 13 at positions that are advanced by
15.degree. in the direction of rotation R of the rotor 41 relative
to the minor axis of the rotor chamber 14. The interior space of
the rotor chamber 14 communicates with the transit chamber 19 by
means of these intake ports 108. Furthermore, a plurality of
exhaust ports 109 are formed in the second casing half 13 at
positions that are retarded by 15.degree. to 75.degree. in the
direction of rotation R of the rotor 41 relative to the minor axis
of the rotor chamber 14. The inner space of the rotor chamber 14
communicates with the exhaust chamber 20 by means of these exhaust
ports 109. These exhaust ports 109 open in shallow depressions 13d
formed within the second casing half 13 so that the seals 72 of the
vanes 48 are not damaged by the edges of the exhaust ports 109.
[0039] The second steam passages S2 and the third steam passages
S3, and the notches 86a of the fixed sleeve 86 and the third steam
passages S3, form a rotary valve V, which provides periodic
communication therebetween by rotation of the rotating shaft 113
relative to the fixed shaft 102 (see FIG. 10).
[0040] As is clear from FIG. 17A and FIG. 17B, a plurality of
notches 92a are formed in a left-hand end outer peripheral portion
of the first fixed shaft 92, and convex portions 92b formed between
the notches 92a are in intimate contact with the cylindrical spring
93a of the fixed shaft support spring 95. Even when the temperature
of the first fixed shaft 92, through which high temperature, high
pressure steam passes, increases, by making only the convex
portions 92b come into contact with the cylindrical spring 93a, the
heat transmitted to the fixed shaft support spring 95 can be
minimized.
[0041] As is clear from FIG. 18A and FIG. 18B, an annular channel
107a is formed on the outer periphery of the nozzle member 107,
which is fitted in the inner sleeve 87, and a plurality of notches
107b are formed in an end portion of the nozzle member 107. This
enables transmission to the inner sleeve 87 of heat of the nozzle
member 107, through which high temperature, high pressure steam
passes, to be minimized.
[0042] As is clear from FIG. 14 to FIG. 16, a plurality (twelve in
the embodiment) of annularly disposed port holes 88d are formed at
two positions of the outer sleeve 88 on either side of the rotary
valve V, and two annularly disposed port channels 87d communicating
with the port holes 88d are formed in the inner sleeve 87. The port
holes 88d and the port channels 87d communicate with the transit
chamber 19 via two passages 87b formed in the axis L direction on
the mating surfaces of the inner sleeve 87 and the outer sleeve 88,
an annular channel 87c formed in the inner sleeve 87, and a through
hole 88a formed in the outer sleeve 88. Segmented spiral channels
88b extending in a spiral shape are formed axially outside the two
lines of port holes 88d of the outer peripheral face of the outer
sleeve 88. The directions of inclination of the spiral channels 88b
on either side of the two lines of port holes 88d are opposite to
each other. Two abraded powder collecting channels 88c are formed
axially inside the two lines of port holes 88d on the outer
peripheral face of the outer sleeve 88.
[0043] As is clear from FIG. 2, pressure chambers 110 are formed at
the rear face of the ring seals 79 fitted in the circular seal
channels 76 of the first and second casing halves 12 and 13. An
eleventh water passage W11 formed in the first and second casing
halves 12 and 13 communicates with the two pressure chambers 110
via a twelfth water passage W12 and a thirteenth water passage W13,
which are formed from pipes, and the ring seals 79 are urged toward
the side face of the rotor 41 by virtue of water pressure applied
to the two pressure chambers 110.
[0044] The eleventh water passage W11 communicates with the outer
peripheral face of the annular filter 30 via a fourteenth water
passage W14, which is a pipe, and the inner peripheral face of the
filter 30 communicates with a sixteenth water passage W16 formed in
the second casing half 13 via a fifteenth water passage W15 formed
in the second casing half 13. Water supplied to the sixteenth water
passage W16 lubricates sliding surfaces between the outer sleeve 88
of the fixed shaft 102 and the inner sleeve 85 of the rotating
shaft 113. Water supplied to the outer periphery of the bearing
member 23 from the inner peripheral face of the filter 30 via a
seventeenth water passage W17 lubricates the outer peripheral face
of the outer sleeve 21 of the rotating shaft 113 through an orifice
penetrating the bearing members 23, and also forms a hydrostatic
bearing to support the rotating shaft 113 in a floating state,
thereby reducing the frictional force and preventing seizing. On
the other hand, water supplied to the outer periphery of the
bearing members 22 from the eleventh water passage W11 via an
eighteenth water passage W18, which is a pipe, lubricates the outer
peripheral face of the outer sleeve 21 of the rotating shaft 113
through an orifice penetrating the bearing member 22, and also
lubricates the sliding surfaces between the outer sleeve 88 of the
fixed shaft 102 and the inner sleeve 85 of the rotating shaft
113.
[0045] Operation of the present embodiment having the
above-mentioned arrangement is now explained.
[0046] Operation of the expander 4 is first explained. In FIG. 3,
high temperature, high pressure steam from the evaporator 3 is
supplied to the steam supply pipe 91, the first steam passage S1
passing through the center of the fixed shaft 102, and the pair of
second steam passages S2 and S2 passing radially through the fixed
shaft 102. In FIG. 10, when the inner sleeve 85 that rotates
integrally with the rotor 41 and the outer sleeve 21 in the
direction shown by the arrow R reaches a predetermined phase
relative to the fixed shaft 102, the pair of third steam passages
S3 that are present on the advanced side in the direction of
rotation R of the rotor 41 relative to the position of the minor
axis of the rotor chamber 14 are made to communicate with the pair
of second steam passages S2, and the high temperature, high
pressure steam of the second steam passages S2 is supplied to the
interiors of a pair of the cylinders 44 via the third steam
passages S3 and pushes the pistons 47 radially outward. In FIG. 4,
when the vanes 48 pushed by the pistons 47 move radially outward,
since the pair of rollers 71 provided on the vanes 48 are engaged
with the annular channels 74, the forward movement of the pistons
47 is converted into rotational movement of the rotor 41.
[0047] Even after the communication between the second steam
passages S2 and the third steam passages S3 is blocked as a result
of the rotation of the rotor 41, the high temperature, high
pressure steam within the cylinders 44 continues to expand, thus
making the pistons 47 move further forward and thereby enabling the
rotor 41 to continue to rotate. When the vanes 48 reach the
position of the major axis of the rotor chamber 14, the third steam
passages S3 communicating with the corresponding cylinders 44 also
communicate with the pair of notches 86a formed on the outer
peripheral face of the fixed sleeve 86, the pistons 47 are pushed
by the vanes 48 whose rollers 71 are guided by the annular channels
74 and move radially inward, and the steam within the cylinders 44
accordingly passes through the third steam passages S3, the notches
86a, the fourth passages S4, the fifth passage S5, and the through
holes 82b, and is supplied to the transit chamber 19 as a first
decreased temperature, decreased pressure steam. The first
decreased temperature, decreased pressure steam is the high
temperature, high pressure steam that has been supplied from the
steam supply pipe 91, has finished work of driving the pistons 47
and, as a result, has a decreased temperature and pressure. The
thermal energy and the pressure energy of the first decreased
temperature, decreased pressure steam are lower than those of the
high temperature, high pressure steam, but are still sufficient for
driving the vanes 48.
[0048] The first decreased temperature, decreased pressure steam
within the transit chamber 19 is supplied to the vane chambers 75
within the rotor chamber 14 via the intake ports 108 of the first
and second casing halves 12 and 13, and further expands therein to
push the vanes 48, thus rotating the rotor 41. A second decreased
temperature, decreased pressure steam that has finished the work
and accordingly has a further decreased temperature and pressure is
discharged from the exhaust ports 109 of the second casing half 13
into the exhaust chamber 20, and is supplied therefrom to the
condenser 5.
[0049] In this way, the expansion of the high temperature, high
pressure steam enables the twelve pistons 47 to operate in turn to
rotate the rotor 41 via the rollers 71 and the annular channels 74,
and the expansion of the first decreased temperature, decreased
pressure steam, which is the high temperature, high pressure steam
whose temperature and pressure have decreased, enables the rotor 41
to rotate via the vanes 48, thereby providing an output from the
rotating shaft 113.
[0050] Lubrication of the vanes 48 and the pistons 47 of the
expander 4 with water is now explained.
[0051] Lubricating water is supplied using the supply pump 6 (see
FIG. 1) for supplying water under pressure from the condenser 5 to
the evaporator 3, and a portion of the water discharged by the
supply pump 6 is supplied to the first water passage W1 of the
casing 11 for the purpose of lubrication. Such use of the supply
pump 6 for supplying water to the hydrostatic bearing of each
section of the expander 4 eliminates the need for a special pump
and enables the number of components to be reduced.
[0052] In FIG. 3 and FIG. 8, the water that has been supplied to
the first water passage W1 of the lubricating water supply member
24 flows into the small diameter portion 55a of one of the pipe
members 55 via the second water passages W2 of the seal block 25,
the third water passages W3 of the outer sleeve 21, the annular
channel 68a of the water passage forming member 68, the fourth
water passage W4 of the outer sleeve 21, and the fifth water
passages W5 formed in the pipe member 69 and the rotor segment 43,
and the water that has flowed into the small diameter portion 55a
flows into the small diameter portion 56a of the other pipe member
56 via the through hole 55b of said one of the pipe members 55, the
sixth water passage W6 formed in the pipe members 55 and 56, and
the through hole 56b formed in the other pipe member 56.
[0053] A portion of the water that has passed through the six
orifices 61b, 61c, and 61d of the orifice-forming plate 61 from the
small diameter portions 55a and 56a of the pipe members 55 and 56
via the distribution channel 62b of the lubricating water
distribution member 62 issues from the four lubricating water
outlets 43e and 43f that open on the end faces of the rotor segment
43, and another portion of the water issues from the lubricating
water outlets 43c and 43d within the arc-shaped recesses 43a and
43b formed on the side faces of the rotor segment 43.
[0054] In this way, the water issuing from the lubricating water
outlets 43e and 43f on the end faces of each of the rotor segments
43 into the vane channel 49 supports the vane 48 in a floating
state by forming a hydrostatic bearing between the vane channel 49
and the vane 48, which is slidably fitted in the vane channel 49,
thus preventing physical contact between the end face of the rotor
segment 43 and the vane 48 and thereby preventing the occurrence of
seizing and wear. Supplying the water for lubricating the sliding
surfaces of the vane 48 via the water passages provided in a radial
shape within the rotor 41 in this way not only enables the water to
be pressurized by virtue of centrifugal force but also enables the
temperature of the periphery of the rotor 41 to be stabilized, thus
lessening the effect of thermal expansion and thereby minimizing
the leakage of steam by maintaining a preset clearance.
[0055] Since water is retained in the recesses 48e, two of which
are formed on each of the opposite faces of the vane 48, these
recesses 48e function as pressure reservoirs, thereby suppressing
any decrease in pressure due to leakage of water. As a result the
vane 48, which is held between the end faces of the pair of rotor
segments 43, is in a floating state due to the water, and the
sliding resistance can thereby be reduced effectively. Furthermore,
when the vane 48 reciprocates, the radial position of the vane 48
relative to the rotor 41 changes, and since the recesses 48e are
provided not on the rotor segment 43 side but on the vane 48 side
and in the vicinity of the rollers 71, where the largest load is
imposed on the vane 48, the reciprocating vane 48 can always be
kept in a floating state, and the sliding resistance can thereby be
reduced effectively.
[0056] The water that has lubricated the sliding surfaces of the
vane 48 that are opposite the rotor segments 43 moves radially
outward by virtue of centrifugal force and lubricates the sliding
section between the seal 72 provided on the arc-shaped face 48b of
the vane 48 and the arc-shaped face 14b of the rotor chamber 14.
Water that has finished lubricating is discharged from the rotor
chamber 14 via the exhaust ports 109.
[0057] In FIG. 2, by supplying water into the pressure chambers 110
at the bottom portions of the circular seal channels 76 of the
first casing half 12 and the second casing half 13 so as to urge
the ring seals 79 toward the side faces of the rotor 41, and making
the water issue from the lubricating water outlets 43c and 43d
formed within the recesses 43a and 43b of each of the rotor
segments 43 so as to form a hydrostatic bearing on the sliding
surfaces with the flat faces 14a of the rotor chamber 14, the flat
faces 41a of the rotor 41 can be sealed by the ring seals 79 that
are in a floating state within the circular seal channels 76 and,
as a result, the steam within the rotor chamber 14 can be prevented
from leaking through a gap with the rotor 41. In this process, the
ring seals 79 and the rotor 41 are isolated from each other by a
film of water supplied from the lubricating water outlets 43c and
43d and do not make physical contact with each other, and even if
the rotor 41 tilts, the damping effect of the ring seals 79
tracking the tilting within the circular seal channels 76 enables
stable sealing characteristics to be maintained while minimizing
the frictional force.
[0058] The water that has lubricated the sliding section between
the ring seals 79 and the rotor 41 is supplied to the rotor chamber
14 by virtue of centrifugal force, and discharged therefrom to the
exterior of the casing 11 via the exhaust ports 109.
[0059] Furthermore, in FIG. 5, water that has been supplied from
the sixth water passage W6 within the pipe member 55 to the sliding
surfaces between the cylinder 44 and the piston 47 via the tenth
water passage W10 within the rotor segments 43 and the annular
channel 67 of the outer periphery of the cylinder 44 exhibits a
sealing function by virtue of the viscous properties of the film of
water formed on the sliding surfaces, thereby preventing
effectively the high temperature, high pressure steam supplied to
the cylinder 44 from leaking past the sliding surfaces with the
piston 47. Since the water that is supplied to the sliding surfaces
between the cylinder 44 and the piston 47 through the interior of
the expander 4, which is in a high temperature state, is heated, it
is possible to minimize any decrease in output of the expander 4
that might be caused by this water cooling the high temperature,
high pressure steam supplied to the cylinder 44.
[0060] Moreover, since water, which is the same substance as steam,
is used as a medium for sealing, there will be no problem even when
the steam is contaminated with water. If the sliding surfaces of
the cylinder 44 and the piston 47 were sealed by an oil, since it
would be impossible to prevent the oil from contaminating the water
or steam, a special filter device for separating the oil would be
required. Furthermore, since a portion of the water for lubricating
the sliding surfaces of the vane 48 and the vane channels 49 is
separated for sealing the sliding surfaces of the cylinder 44 and
the piston 47, it is unnecessary to specially provide an extra
water passage for guiding the water to the sliding surfaces, thus
simplifying the structure.
[0061] In order to maintain the sealing characteristics for the
steam in the rotary valve V, it is necessary to precisely control
the clearance between the sliding surfaces of the rotating shaft
113 and the fixed shaft 102. When the expander 4 is cold, the fixed
shaft 102, through which the high temperature steam passes, first
expands thermally in the vicinity of the rotary valve V, the
rotating shaft 113 then thermally expands after a time lag, and the
difference in thermal expansion causes wear of the outer peripheral
face of the fixed shaft 102. During this process, if the fixed
shaft 102 is firmly fixed to the casing 11, rotational runout of
the rotor 41 results in uneven contact with the outer peripheral
face of the fixed shaft 102, thereby causing eccentric wear, and
giving rise to problems such as degradation of the sealing
characteristics for the steam in the rotary valve V, an increase in
the sliding resistance, and degradation in the rotational behavior
of the rotor 41.
[0062] However, in accordance with the present embodiment, since
the fixed shaft 102 is floatingly supported by the fixed shaft
support spring 95 relative to the casing 11, when the rotational
runout of the rotor 41 is transmitted to the fixed shaft 102 via
the rotating shaft 113, the alignment action arising from tracking
exhibited by the damping effect of the fixed shaft support spring
95 suppresses the rotational runout of the rotor 41, and any
increase in the frictional resistance in the sliding section
between the fixed shaft 102 and the rotating shaft 113 and the
occurrence of abnormal wear can be prevented effectively. In this
way, if the outer peripheral face of the fixed shaft 102 is
uniformly worn by the action of the fixed shaft support spring 95,
the clearance of the uniformly worn section of the fixed shaft 102
is uniformly reduced when the expander 4 is hot, and the sealing
characteristics of the rotary valve V can be ensured. Since the
left-hand end of the fixed shaft 102 is supported via the Oldham
coupling 89 in a non-rotatable but radially movable manner, the
alignment action of the fixed shaft 102 due to the tracking
exhibited by the damping effect of the fixed shaft support spring
95 can be exhibited without any problem.
[0063] Suppressing the thermal expansion of the fixed shaft 102 due
to the heat of the steam to a low level enables wear of the outer
peripheral face of the fixed shaft 102 in the vicinity of the
rotary valve V to be further reduced. In the present embodiment,
the fixed sleeve 86 is therefore formed by shrink-fitting the outer
sleeve 88, which is made of metal, around the outer periphery of
the inner sleeve 87, which is made of ceramic, etc. having a small
coefficient of thermal expansion.
[0064] That is, as shown in FIG. 20A, the outer diameter Do of the
inner sleeve 87 is larger than the inner diameter Di of the outer
sleeve 88 at room temperature, and the outer sleeve 88 is fitted
around the outer periphery of the inner sleeve 87 in a state, as
shown in FIG. 20B, in which the inner diameter Di' thereof is made
larger than the outer diameter Do of the inner sleeve 87 by heating
the outer sleeve 88, which is made of metal, so as to thermally
expand it. When the outer sleeve 88 is cooled so as to shrink it in
this state, the inner peripheral face of the outer sleeve 88 comes
into intimate contact with the outer peripheral face of the inner
sleeve 87 as shown in FIG. 20C, thus completing the shrink-fitting.
In a state in which the shrink-fitting is completed, the outer
sleeve 88, whose inner diameter should have decreased to Di (broken
line), is restrained by the inner sleeve 87, and the inner diameter
only decreases to an inner diameter D", which is larger than the
above Di (Di<Di"<D'), and the outer sleeve 88 is in a state
in which an internal stress acts on it in a tensile direction.
[0065] Therefore, as shown in FIG. 20D, when the outer sleeve 88
and the inner sleeve 87 are heated by steam, the thermal expansion
of the outer sleeve 88 is canceled by the internal stress in the
tensile direction, and the outer diameter of the outer sleeve 88
does not increase substantially. In practice, the outer diameter of
the outer sleeve 88 is controlled by the small amount of thermal
expansion of the inner sleeve 87, which is made of ceramic, etc.
having a small coefficient of thermal expansion, and increases
slightly due to being widened by the inner sleeve 87. In this way,
since the change due to thermal expansion in the outer diameter of
the fixed sleeve 86 having the outer sleeve 88, which is a collar
made of an easily stretched metal and is in sliding contact with
the inner sleeve 85 of the rotating shaft 113, can be suppressed by
shrink-fitting, wear of the outer peripheral face of the fixed
sleeve 86 can be minimized, thereby preventing the leakage of steam
from the rotary valve V.
[0066] Since the outer sleeve 88 of the fixed sleeve 86 is made of
metal, a coating of a low friction material, which is difficult to
apply to a ceramic sleeve, can be applied to the outer sleeve 88
and this, together with the structure of the shrink-fitting on the
rotating shaft 113 side, enables the frictional resistance between
the outer sleeve 88 and the inner sleeve 85 to be further reduced,
thus suppressing any increase in the clearance and reducing the
leakage of steam.
[0067] In the same way as for the fixed sleeve 86 of the
above-mentioned fixed shaft 102, the rotating shaft 113 is also
formed by uniting the outer sleeve 21, which is made of metal, with
the outer periphery of the ceramic inner sleeve 85 by
shrink-fitting, and the outer sleeve 21 is in a state in which an
internal stress acts in the tensile direction.
[0068] The effect of the shrink-fitting is now explained with
reference to FIG. 21A to FIG. 21D.
[0069] FIG. 21D corresponds to a conventional example in which both
the rotating shaft 113 and the fixed shaft 102 are made of metal,
and when high temperature steam is supplied to the rotary valve V
through the interior of the fixed shaft 102 when it is cold, the
fixed shaft 102 side first expands thermally to a large extent and
comes into contact with the inner peripheral face of the rotating
shaft 113, and wear of the sliding surfaces occurs between point a
and point b. This wear occurs only when running the expander 4 for
the first time after assembly. When, after time has elapsed, it is
hot, that is, when the temperatures of both the fixed shaft 102 and
the rotating shaft 113 are sufficiently high, the amount of
expansion of the rotating shaft 113 becomes larger than the amount
of expansion of the fixed shaft 102, and the clearance therebetween
gradually enlarges. In this way, in the conventional arrangement,
both the fixed shaft 102 and the rotating shaft 113 expand
thermally, thus generating wear of the sliding surfaces and
increasing the clearance when hot.
[0070] On the other hand, FIG. 21A shows the characteristics of the
present embodiment in which shrink-fitting is employed for both the
rotating shaft 113 and the fixed shaft 102. The radii of the
rotating shaft 113 and the fixed shaft 102 hardly change from when
they are cold to when they are hot, and the clearance between the
sliding surfaces thereof is always maintained substantially
constant.
[0071] FIG. 21B shows the characteristics when shrink-fitting is
employed only for the rotating shaft 113 side. The fixed shaft 102
side expands thermally accompanying the starting of the supply of
steam and comes into contact with the inner peripheral face of the
rotating shaft 113, which hardly expands at all, thereby generating
wear on the outer peripheral face of the fixed shaft 102. This wear
occurs only when running the expander 4 for the first time after
assembly, and once bedding in due to the wear is completed, the
clearance between the sliding surfaces is always maintained
substantially constant in subsequent running.
[0072] FIG. 21C shows the characteristics when shrink-fitting is
employed only for the fixed shaft 102 side. The rotating shaft 113
side expands thermally accompanying the starting of the supply of
steam and the clearance between itself and the rotating shaft 113,
which hardly expands at all thermally, gradually increases, but
since contact between the fixed shaft 102 and the rotating shaft
113 is avoided, wear will not be caused, and the sliding resistance
therebetween can be minimized.
[0073] As hereinbefore described, the maximum effect can be
obtained when shrink-fitting is employed for both the rotating
shaft 113 and the fixed shaft 102, and the expected effect can also
be obtained when shrink-fitting is employed for only one of the
rotating shaft 113 or the fixed shaft 102.
[0074] Even if an attempt is made to prevent the steam from leaking
from the rotary valve V as described above, it is impossible to
prevent a slight amount of steam from leaking past the sliding
surfaces of the rotating shaft 113 and the fixed shaft 102. This
leaked steam is captured by the port holes 88d and the port
channels 87d annularly formed on the outer peripheral face of the
fixed sleeve 86, and is supplied therefrom to the transit chamber
19 via the two passages 87b formed on the mating surfaces between
the inner sleeve 87 and the outer sleeve 88, the annular channel
87c formed in the inner sleeve 87, and the through hole 88a formed
in the outer sleeve 88. The steam that has been supplied to the
transit chamber 19 is combined with the first decreased
temperature, decreased pressure steam that has finished driving the
pistons 47, and is provided for driving the vanes 48. In this way,
the steam that has leaked from the rotary valve V is captured by
the port holes 88d and the port channels 87d and reused, thereby
contributing an improvement of the overall energy efficiency of the
expander 4.
[0075] When the outer sleeve 88, which is made of metal, of the
fixed sleeve 86 is worn due to sliding against the ceramic inner
sleeve 85 of the rotating shaft 113, the abraded powder thus formed
is collected by the abraded powder collecting channels 88c formed
on the outer peripheral face of the outer sleeve 88, and thereby
prevented from accumulating on the sliding surfaces of the fixed
sleeve 86 and the inner sleeve 85 of rotating shaft 113. It is
thereby possible to avoid any increase in the frictional resistance
and the occurrence of seizure of the sliding surfaces.
[0076] If the water that has been supplied from the sixteenth water
passage W16 and lubricated the sliding surfaces of the fixed sleeve
86 and the inner sleeve 85 of the rotating shaft 113 and the water
that has lubricated the outer peripheral face of the rotating shaft
113 through the orifice penetrating the bearing members 22 and 23
and has also lubricated the sliding surfaces of the fixed sleeve 86
and the inner sleeve 85 of the rotating shaft 113 were to flow into
the transit chamber 19 via the port holes 88d and the port channels
87d formed in the outer periphery of the fixed sleeve 86, the first
decreased temperature, decreased pressure steam within the transit
chamber 19 might be cooled, and the output of the expander 4 might
be degraded.
[0077] However, in accordance with the present embodiment, when the
water that lubricates the sliding surfaces of the fixed sleeve 86
and the inner sleeve 85 of the rotating shaft 113 flows from
opposite ends of the fixed sleeve 86 toward the port holes 88d and
the port channels 87d in the center, the spiral channels 88b formed
on the outer periphery of the outer sleeve 88 can exhibit an effect
of generating a pressure so as to push back the lubricating water
away from the port holes 88d and the port channels 87d. That is, as
a result of the relative rotation between the inner sleeve 85 of
the rotating shaft 113 and the fixed sleeve 86 the lubricating
water retained in the spiral channels 88b is pressurized by a
spring pump action and pushed back in a direction away from the
port holes 88d and the port channels.
[0078] If the spiral channels 88b were made to communicate with the
port holes 88d and the port channels 87d without being sectioned
into short lengths, there is the possibility that high pressure
lubricating water might pass through the interior of the spiral
channels 88b without being stopped and flow into the low pressure
port holes 88d and the port channels 87d, but this problem can be
solved by sectioning the spiral channels 88b into short
lengths.
[0079] Furthermore, the first water passage W1 and the eleventh
water passage W11 are independent from each other, and water is
supplied at a pressure that is required for each of the lubrication
sections. More specifically, the water that is supplied from the
first water passage W1 is mainly for floatingly supporting the
vanes 48 and the rotor 41 by means of a hydrostatic bearing as
described above, and it is required to have a high pressure that
can counterbalance variations in the load. In contrast, the water
that is supplied from the eleventh water passage W11 mainly
lubricates the surroundings of the fixed shaft 102 and the bearing
members 22 and 23 and also forms a hydrostatic bearing, and since
it is for sealing the high temperature, high pressure steam that
leaks from the third steam passages S3 and S3 past the outer
periphery of the fixed shaft 102 so as to reduce the influence of
thermal expansion of the fixed shaft 102, the rotating shaft 113,
the rotor 41, etc., it is required to have a pressure that is at
least higher than the pressure of the transit chamber 19.
[0080] Since there are provided in this way two water supply lines,
that is, the first water passage W1 for supplying high pressure
water and the eleventh water passage W11 for supplying lower
pressure water, problems caused when only one water supply line for
supplying high pressure water is provided can be eliminated. That
is, the problem of water having excess pressure being supplied to
the surroundings of the fixed shaft 102, thus increasing the amount
of water flowing into the transit chamber 19, and the problem of
the fixed shaft 102, the rotating shaft 113, the rotor 41, etc.
being overcooled, thus decreasing the temperature of the steam, can
be prevented, and as a result the output of the expander 4 can be
increased while reducing the amount of water supplied.
[0081] A second embodiment of the present invention is now
explained with reference to FIG. 22 and FIG. 23. The second
embodiment is different from the first embodiment with respect to
the structure of the fixed shaft support spring 95, and the
structures of the other parts are the same as those of the first
embodiment.
[0082] In the second embodiment, a second fixed shaft 93 extends
leftward so as to cover the outer periphery of a steam supply pipe
91, and the left-hand end of the second fixed shaft 93 is fitted in
and fixed to a boss portion 81a of a spring support member 81. A
plurality (eight in this embodiment) of slits 93b extending in the
axis L direction are formed in the second fixed shaft 93 adjacent
to the boss portion 81a of the spring support member 81, and the
section where these slits 93b are formed functions as a fixed shaft
support spring 95. The fixed shaft support spring 95 can easily be
elastically deformed in the radial direction by virtue of the slits
93b and, moreover, it can withstand a load in the axis L direction
without being deformed.
[0083] In order to prevent steam that has leaked from the section
where the right-hand end of the steam supply pipe 91 and the
left-hand end of the first fixed shaft 92 are fitted together from
passing through the slits 93b of the fixed shaft support spring 95
and leaking into the transit chamber 19, the outer periphery of the
fixed shaft support spring 95 is covered by a sealing tube 111 and
bellows 112. The right-hand end of the sealing tube 111 is held
between the second fixed shaft 93 and the inner sleeve 87, and the
left-hand end thereof extends to a middle section of the fixed
shaft support spring 95. The left-hand end of the bellows 112 is
welded to the boss portion 81a of the spring support member 81, and
the right-hand end thereof is welded to the right-hand end of the
sealing tube 111. Since the sealing tube 111 and the bellows 112
can easily flex in the radial direction, elastic deformation of the
fixed shaft support spring 95 is not inhibited.
[0084] This second embodiment can also achieve the same effects as
those obtained by the above-mentioned first embodiment.
[0085] Other than the embodiments described above, as an
arrangement for a power conversion device for converting the
forward movement of pistons 47 into the rotational movement of a
rotor 41, the forward movement of the pistons 47 can be directly
transmitted to rollers 71 without involving vanes 48, and can be
converted into rotational movement by engagement with annular
channels 74. Furthermore, as long as the vanes 48 are always spaced
from the inner peripheral face of a rotor chamber 14 by a
substantially constant gap as a result of cooperation between the
rollers 71 and the annular channels 74 as described above, the
pistons 47 and the rollers 71, and also the vanes 48 and the
rollers 71, can independently work together with the annular
channels 74.
[0086] When the expander 4 is used as a compressor, the rotor 41 is
rotated by the rotating shaft 113 in a direction opposite to the
arrow R in FIG. 4, outside air is drawn in by the vanes 48 from the
exhaust ports 109 into the rotor chamber 14 and compressed, and the
low pressure compressed air thus obtained is drawn in from the
intake ports 108 into the cylinders 44 via the transit chamber 19,
the through holes 82b, the fifth steam passages S5, the fourth
steam passages S4, the notches 86a of the fixed shaft 102 and the
third steam passages S3, and compressed there by the pistons 47 to
give high pressure compressed air. The high pressure compressed air
thus obtained is discharged from the cylinders 44 via the third
steam passages S3, the second steam passages S2, the first steam
passage S1, and the steam supply pipe 91. When the expander 4 is
used as a compressor, the steam passages S1 to S5 and the steam
supply pipe 91 are read instead as air passages S1 to S5 and air
supply pipe 91.
[0087] Although embodiments of the present invention are described
in detail above, the present invention can be modified in a variety
of ways without departing from the scope and spirit thereof.
[0088] For example, in the embodiments, the expander 4 is
illustrated as the rotary fluid machine, but the present invention
can also be applied to a compressor.
[0089] Furthermore, in the embodiments, steam and water are used as
the gas-phase working medium and the liquid-phase working medium,
but other appropriate working media can also be employed.
[0090] Industrial Applicability
[0091] The present invention can desirably be applied to an
expander employing steam (water) as a working medium, but can also
be applied to an expander employing any other working medium and a
compressor employing any working medium.
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