U.S. patent application number 10/500691 was filed with the patent office on 2005-10-06 for rotary type fluid machine.
Invention is credited to Honma, Kensuke, Nakino, Hiroyuki.
Application Number | 20050220643 10/500691 |
Document ID | / |
Family ID | 27606004 |
Filed Date | 2005-10-06 |
United States Patent
Application |
20050220643 |
Kind Code |
A1 |
Honma, Kensuke ; et
al. |
October 6, 2005 |
Rotary type fluid machine
Abstract
A rotary fluid machine is provided that includes a rotary valve
(61) for controlling the intake and discharge of a working medium
to and from an operating part (49, 57) formed from an axial piston
cylinder group, a steam supply pipe (77) that is disposed on an
axis (L) of a rotor (27) and supplies steam to the rotary valve
(61), the steam supply pipe (77) being provided separately from a
rotary valve main body (62), and gland packing sealing means (97)
disposed between the steam supply pipe (77) and the rotary valve
main body (62). Since the sealing means (97), which is flexible,
has the function of preventing movement in the direction of the
axis (L) of the steam supply pipe (77) from being transmitted to
the rotary valve (61), it is possible to ensure the intimacy of
contact of sliding surfaces (68) of the rotary valve (61) while
minimizing the leakage of steam past the outer periphery of the
steam supply pipe (77) by means of the sealing means (97), thereby
enabling the steam to be supplied and discharged reliably.
Inventors: |
Honma, Kensuke; (Wako-shi,
Saitama, JP) ; Nakino, Hiroyuki; (Wako-shi, Saitama,
JP) |
Correspondence
Address: |
BIRCH STEWART KOLASCH & BIRCH
PO BOX 747
FALLS CHURCH
VA
22040-0747
US
|
Family ID: |
27606004 |
Appl. No.: |
10/500691 |
Filed: |
March 4, 2005 |
PCT Filed: |
January 17, 2003 |
PCT NO: |
PCT/JP03/00332 |
Current U.S.
Class: |
417/505 ;
417/521; 418/269 |
Current CPC
Class: |
F04B 27/0808 20130101;
F01B 3/02 20130101; F04B 1/22 20130101; F04B 27/0843 20130101; F04B
1/205 20130101 |
Class at
Publication: |
417/505 ;
418/269; 417/521 |
International
Class: |
F04B 007/00; F04B
023/04 |
Foreign Application Data
Date |
Code |
Application Number |
Jan 21, 2002 |
JP |
2002-11015 |
Claims
1. A rotary fluid machine comprising: a casing (11); a rotor (27)
rotatably supported in the casing (11); an operating part (49, 57)
provided in the rotor (27); and a rotary valve (61) that is
provided between the casing (11) and the rotor (27) and controls
the intake and discharge of a working medium to and from the
operating part (49, 57) via sliding surfaces (68) that are
perpendicular to the axis (L) of the rotor (27); wherein a working
medium supply pipe (77) is provided separately from the rotary
valve (61), the working medium supply pipe being positioned on the
axis (L) of the rotor (27) and supplying the working medium to the
rotary valve (61), and sealing means (97) is disposed between the
working medium supply pipe (77) and the rotary valve (61), the
sealing means (97) having the function of preventing movement of
the working medium supply pipe (77) in the axial (L) direction of
the rotor (27) from being transmitted to the rotary valve (61).
2. The rotary fluid machine according to claim 1, wherein the
sealing means (97) is a gland packing.
3. The rotary fluid machine according to claim 2, wherein the
rotary fluid machine further comprises working medium recovery
means (94, 18e) for recovering working medium that has leaked past
the sealing means (97).
4. The rotary fluid machine according to claim 3, wherein the
working medium recovery means (94, 18e) returns the recovered
working medium to a downstream side of the operating part (49, 57).
Description
FIELD OF THE INVENTION
[0001] The present invention relates to a rotary fluid machine that
includes a casing, a rotor rotatably supported in the casing, an
operating part provided in the rotor, and a rotary valve that is
provided between the casing and the rotor and controls the intake
and discharge of a working medium to and from the operating part
via sliding surfaces that are perpendicular to the axis of the
rotor.
BACKGROUND ART
[0002] The rotary valve of this type of rotary fluid machine
generally includes a valve main body fixed to the casing so as to
be positioned on the axis of the rotor, and controls the supply and
discharge of the working medium via sliding surfaces of the fixed
valve main body and the rotating rotor. Supply of the working
medium to the rotary valve is carried out via a working medium
supply pipe fixed to the valve main body so as to be disposed on
the axis of the rotor, and the valve main body is resiliently
biased toward the rotor so that the working medium does not leak
past the sliding surfaces of the valve main body and the rotor.
[0003] However, in this conventional rotary fluid machine, since
the working medium supply pipe is fixed to the valve main body,
axial movement of the valve main body is restricted by the working
medium supply pipe, or vibration of the working medium supply pipe
is transmitted to the valve main body. The intimacy of contact
between the sliding surfaces of the valve main body and the rotor
is therefore degraded, and there is thus the problem that supply
and discharge of the working medium becomes inaccurate.
DISCLOSURE OF THE INVENTION
[0004] The present invention has been achieved under the
above-mentioned circumstances, and it is an object thereof to
ensure the intimacy of contact between the sliding surfaces of the
valve main body and the rotor of the rotary valve of the rotary
fluid machine.
[0005] In order to accomplish this object, in accordance with a
first aspect of the present invention, there is proposed a rotary
fluid machine that includes a casing, a rotor rotatably supported
in the casing, an operating part provided in the rotor, and a
rotary valve that is provided between the casing and the rotor and
controls the intake and discharge of a working medium to and from
the operating part via sliding surfaces that are perpendicular to
the axis of the rotor, the rotary fluid machine further including a
working medium supply pipe provided separately from the rotary
valve, the working medium supply pipe being positioned on the axis
of the rotor and supplying the working medium to the rotary valve,
and sealing means disposed between the working medium supply pipe
and the rotary valve, the sealing means having the function of
preventing movement of the working medium supply pipe in the axial
direction of the rotor from being transmitted to the rotary
valve.
[0006] In accordance with this arrangement, since the working
medium supply pipe that is disposed on the axis of the rotor and
supplies the working medium to the rotary valve is provided
separately from the rotary valve, and the sealing means disposed
between the working medium supply pipe and the rotary valve has the
function of preventing movement of the working medium supply pipe
in the axial direction of the rotor from being transmitted to the
rotary valve, it is possible to ensure the intimacy of contact
between the sliding surfaces of the rotary valve while minimizing,
with the sealing means, leakage of the working medium past the
outer periphery of the working medium supply pipe, thereby enabling
the working medium to be supplied and discharged reliably.
[0007] 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 the sealing means is a
gland packing.
[0008] In accordance with this arrangement, since the sealing means
disposed between the working medium supply pipe and the rotary
valve is formed from a gland packing, not only is the durability of
the sealing means against high temperature working medium
increased, but also it is possible to prevent axial movement of the
working medium supply pipe from being transmitted to the rotary
valve by allowing relative movement between the working medium
supply pipe and the rotary valve.
[0009] Moreover, in accordance with a third aspect of the present
invention, in addition to the second aspect, there is proposed a
rotary fluid machine wherein the rotary fluid machine further
includes working medium recovery means for recovering working
medium that has leaked past the sealing means.
[0010] In accordance with this arrangement, since the working
medium that has leaked past the sealing means is recovered by the
working medium recovery means, the necessity for replenishing the
working medium can be minimized.
[0011] Furthermore, in accordance with a fourth aspect of the
present invention, in addition to the third aspect, there is
proposed a rotary fluid machine wherein the working medium recovery
means returns the recovered working medium to a downstream side of
the operating part.
[0012] In accordance with this arrangement, since the working
medium that has leaked past the sealing means is returned to the
downstream side of the operating part via the working medium
recovery means, it is possible to prevent the recovered working
medium from affecting the performance of the operating part.
[0013] A first axial piston cylinder group 49 and a second axial
piston cylinder group 57 of an embodiment correspond to the
operating part of the present invention, a steam supply pipe 77 of
the embodiment corresponds to the working medium supply pipe of the
present invention, and a spring case 94 and a steam recovery
passage 18e correspond to the working medium recovery means of the
present invention.
BRIEF DESCRIPTION OF DRAWINGS
[0014] FIG. 1 to FIG. 13 illustrate one embodiment of the present
invention; FIG. 1 is a vertical sectional view of an expander, FIG.
2 is a sectional view along line 2-2 in FIG. 1, FIG. 3 is an
enlarged view of part 3 in FIG. 1, FIG. 4 is an enlarged sectional
view of part 4 in FIG. 1 (sectional view along line 4-4 in FIG. 8),
FIG. 5 is a view from arrowed line 5-5 in FIG. 4, FIG. 6 is a view
from arrowed line 6-6 in FIG. 4, FIG. 7 is a sectional view along
line 7-7 in FIG. 4, FIG. 8 is a sectional view along line 8-8 in
FIG. 4, FIG. 9 is a sectional view along line 9-9 in FIG. 4, FIG.
10 is a graph showing torque variations of an output shaft, FIG. 11
is an explanatory diagram showing the operation of an intake system
of a high-pressure stage, FIG. 12 is an explanatory diagram showing
the operation of a discharge system of the high-pressure stage and
an intake system of a low-pressure stage, and FIG. 13 is an
explanatory diagram showing the operation of a discharge system of
the low-pressure stage.
BEST MODE FOR CARRYING OUT THE INVENTION
[0015] An embodiment of the present invention is explained below
with reference to the attached drawings.
[0016] As shown in FIG. 1 to FIG. 3, a rotary fluid machine of the
present invention is, for example, an expander M used in a Rankine
cycle system, and the thermal energy and the pressure energy of
high-temperature, high-pressure steam as a working medium are
converted into mechanical energy and output. A casing 11 of the
expander M is formed from a casing main body 12, a front cover 15
fitted via a seal 13 in a front opening of the casing main body 12
and joined thereto via a plurality of bolts 14, and a rear cover 18
fitted via a seal 16 in a rear opening of the casing main body 12
and joined thereto via a plurality of bolts 17. An oil pan 19 abuts
against a lower opening of the casing main body 12 via a seal 20
and is joined thereto via a plurality of bolts 21. Furthermore, a
breather chamber dividing wall 23 is superimposed on an upper face
of the casing main body 12, a breather chamber cover 25 is further
superimposed on an upper face of the breather chamber dividing wall
23, and they are together secured to the casing main body 12 by
means of a plurality of bolts 26.
[0017] A rotor 27 and an output shaft 28 that can rotate around an
axis L extending in the fore-and-aft direction in the center of the
casing 11 are united by welding. A rear part of the rotor 27 is
rotatably supported in the casing main body 12 via an angular ball
bearing 29 and a seal 30, and a front part of the output shaft 28
is rotatably supported in the front cover 15 via an angular ball
bearing 31 and a seal 32. A swash plate holder 36 is fitted via two
seals 33 and 34 and a knock pin 35 in a rear face of the front
cover 15 and fixed thereto via a plurality of bolts 37, and a swash
plate 39 is rotatably supported in the swash plate holder 36 via an
angular ball bearing 38. The axis of the swash plate 39 is inclined
relative to the axis L of the rotor 27 and the output shaft 28, and
the angle of inclination is fixed.
[0018] Seven sleeves 41 formed from members that are separate from
the rotor 27 are arranged within the rotor 27 so as to surround the
axis L at equal intervals in the circumferential direction.
High-pressure pistons 43 are slidably fitted in high-pressure
cylinders 42 formed at inner peripheries of the sleeves 41, which
are supported by sleeve support bores 27a of the rotor 27.
Hemispherical parts of the high-pressure pistons 43 projecting
forward from forward end openings of the high-pressure cylinders 42
abut against seven dimples 39a recessed in a rear face of the swash
plate 39. Heat resistant metal seals 44 are fitted between the rear
ends of the sleeves 41 and the sleeve support bores 27a of the
rotor 27, and a single set plate 45 retaining the front ends of the
sleeves 41 in this state is fixed to a front face of the rotor 27
by means of a plurality of bolts 46. The sleeve support bores 27a
have a slightly larger diameter in the vicinity of their bases,
thus forming a gap a (see FIG. 3) between themselves and the outer
peripheries of the sleeves 41.
[0019] The high-pressure pistons 43 include pressure rings 47 and
oil rings 48 for sealing the sliding surfaces with the
high-pressure cylinders 42, and the sliding range of the pressure
rings 47 and the sliding range of the oil rings 48 are set so as
not to overlap each other. When the high-pressure pistons 43 are
inserted into the high-pressure cylinders 42, in order to make the
pressure rings 47 and the oil rings 48 engage smoothly with the
high-pressure cylinders 42, tapered openings 45a widening toward
the front are formed in the set plate 45.
[0020] As hereinbefore described, since the sliding range of the
pressure rings 47 and the sliding range of the oil rings 48 are set
so as not to overlap each other, oil attached to the inner walls of
the high-pressure cylinders 42 against which the oil rings 48 slide
will not be taken into high-pressure operating chambers 82 due to
sliding of the pressure rings 47, thereby reliably preventing the
oil from contaminating the steam. In particular, the high-pressure
pistons 43 have a slightly smaller diameter part between the
pressure rings 47 and the oil rings 48 (see FIG. 3), thereby
preventing effectively the oil attached to the sliding surfaces of
the oil rings 48 from moving to the sliding surfaces of the
pressure rings 47.
[0021] Since the high-pressure cylinders 42 are formed by fitting
the seven sleeves 41 in the sleeve support bores 27a of the rotor
27, a material having excellent thermal conductivity, heat
resistance, abrasion resistance, strength, etc. can be selected for
the sleeves 41. This not only improves the performance and the
reliability, but also machining becomes easy compared with a case
in which the high-pressure cylinders 42 are directly machined in
the rotor 27, and the machining precision also increases. When any
one of the sleeves 41 is worn or damaged, it is possible to
exchange only the sleeve 41 with an abnormality, without exchanging
the entire rotor 27, and this is economical.
[0022] Furthermore, since the gap a is formed between the outer
periphery of the sleeves 41 and the rotor 27 by slightly enlarging
the diameter of the sleeve support bores 27a in the vicinity of the
base, even when the rotor 27 is thermally deformed by the
high-temperature, high-pressure steam supplied to the high-pressure
operating chambers 82, this is prevented from affecting the sleeves
41, thereby preventing the high-pressure cylinders 42 from
distorting.
[0023] The seven high-pressure cylinders 42 and the seven
high-pressure pistons 43 fitted therein form a first axial piston
cylinder group 49.
[0024] Seven low-pressure cylinders 50 are arranged at
circumferentially equal intervals on the outer peripheral part of
the rotor 27 so as to surround the axis L and the radially outer
side of the high-pressure cylinders 42. These low-pressure
cylinders 50 have a larger diameter than that of the high-pressure
cylinders 42, and the pitch at which the low-pressure cylinders 50
are arranged in the circumferential direction is displaced by half
a pitch relative to the pitch at which the high-pressure cylinders
42 are arranged in the circumferential direction. This makes it
possible for the high-pressure cylinders 42 to be arranged in
spaces formed between adjacent low-pressure cylinders 50, thus
utilizing the spaces effectively and contributing to a reduction in
the diameter of the rotor 27.
[0025] The seven low-pressure cylinders 50 have low-pressure
pistons 51 slidably fitted thereinto, and these low-pressure
pistons 51 are connected to the swash plate 39 via links 52. That
is, spherical parts 52a at the front end of the links 52 are
swingably supported in spherical bearings 54 fixed to the swash
plate 39 via nuts 53, and spherical parts 52b at the rear end of
the links 52 are swingably supported in spherical bearings 56 fixed
to the low-pressure pistons 51 by clips 55. A pressure ring 78 and
an oil ring 79 are fitted around the outer periphery of each of the
low-pressure pistons 51 in the vicinity of the top face thereof so
as to adjoin each other. Since the sliding ranges of the pressure
ring 78 and the oil ring 79 overlap each other, an oil film is
formed on the sliding surface of the pressure ring 78, thus
enhancing the sealing characteristics and the lubrication.
[0026] The seven low-pressure cylinders 50 and the seven
low-pressure pistons 44 fitted therein form a second axial piston
cylinder group 57.
[0027] As hereinbefore described, since the front ends of the
high-pressure pistons 43 of the first axial piston cylinder group
49 are made in the form of hemispheres and are made to abut against
the dimples 39a formed in the swash plate 39, it is unnecessary to
connect the high-pressure pistons 43 to the swash plate 39
mechanically, thus reducing the number of parts and improving the
ease of assembly. On the other hand, the low-pressure pistons 51 of
the second axial piston cylinder group 57 are connected to the
swash plate 39 via the links 52 and the front and rear spherical
bearings 54 and 56, and even when the temperature and the pressure
of medium-temperature, medium-pressure steam supplied to the second
axial piston cylinder group 57 become insufficient and the pressure
of low-pressure operating chambers 84 becomes negative, there is no
possibility of the low-pressure pistons 51 becoming detached from
the swash plate 39 and causing knocking or damage.
[0028] Furthermore, when the swash plate 39 is secured to the front
cover 15 via the bolts 37, changing the phase at which the swash
plate 39 is secured around the axis L enables the timing of supply
and discharge of the steam to and from the first axial piston
cylinder group 49 and the second axial piston cylinder group 57 to
be shifted, thereby altering the output characteristics of the
expander M.
[0029] Moreover, since the rotor 27 and the output shaft 28, which
are united, are supported respectively by the angular ball bearing
29 provided on the casing main body 12 and the angular ball bearing
31 provided on the front cover 15, by adjusting the thickness of a
shim 58 disposed between the casing main body 12 and the angular
ball bearing 29 and the thickness of a shim 59 disposed between the
front cover 15 and the angular ball bearing 31, the longitudinal
position of the rotor 27 along the axis L can be adjusted. By
adjusting the position of the rotor 27 in the axis L direction, the
relative positional relationship in the axis L direction between
the high-pressure and low-pressure pistons 43 and 51 guided by the
swash plate 39, and the high-pressure and low-pressure cylinders 42
and 50 provided in the rotor 27 can be changed, thereby adjusting
the expansion ratio of the steam in the high-pressure and
low-pressure operating chambers 82 and 84.
[0030] If the swash plate holder 36 supporting the swash plate 39
were formed integrally with the front cover 15, it would be
difficult to secure a space for attaching and detaching the angular
ball bearing 31 or the shim 59 to and from the front cover 15, but
since the swash plate holder 36 is made detachable from the front
cover 15, the above-mentioned problem can be eliminated. Moreover,
if the swash plate holder 36 were integral with the front cover 15,
during assembly and disassembly of the expander M it would be
necessary to carry out cumbersome operations of connecting and
disconnecting the seven links 52, which are in a confined space
within the casing 11, to and from the swash plate 39 pre-assembled
to the front cover 15, but since the swash plate holder 36 is made
detachable from the front cover 15, it becomes possible to form a
sub-assembly by assembling the swash plate 39 and the swash plate
holder 36 to the rotor 27 in advance, thereby greatly improving the
ease of assembly.
[0031] Systems for supply and discharge of steam to and from the
first axial piston cylinder group 49 and the second axial piston
cylinder group 57 are now explained with reference to FIG. 4 to
FIG. 9.
[0032] As shown in FIG. 4, a rotary valve 61 is housed in a
circular cross-section recess 27b opening on a rear end face of the
rotor 27 and a circular cross-section recess 18a opening on a front
face of the rear cover 18. The rotary valve 61, which is disposed
along the axis L, includes a rotary valve main body 62, a
stationary valve plate 63, and a movable valve plate 64. The
movable valve plate 64 is fixed to the rotor 27 via a knock pin 66
and a bolt 67 while being fitted to the base of the recess 27b of
the rotor 27 via a gasket 65. The stationary valve plate 63, which
abuts against the movable valve plate 64 via flat sliding surfaces
68, is joined via a knock pin 69 to the rotary valve main body 62
so that there is no relative rotation therebetween. When the rotor
27 rotates, the movable valve plate 64 and the stationary valve
plate 63 therefore rotate relative to each other on the sliding
surfaces 68 in a state in which they are in intimate contact with
each other. The stationary valve plate 63 and the movable valve
plate 64 are made of a material having excellent durability, such
as a super hard alloy or a ceramic, and the sliding surfaces 68 can
be provided with or coated with a member having heat resistance,
lubrication, corrosion resistance, and abrasion resistance.
[0033] The rotary valve main body 62 is a stepped cylindrical
member having a large diameter part 62a, a medium diameter part
62b, and a small diameter part 62c; an annular sliding member 70
fitted around the outer periphery of the large diameter part 62a is
slidably fitted in the recess 27b of the rotor 27 via a cylindrical
sliding surface 71, the medium diameter part 62b and the small
diameter part 62c are fitted in an inner periphery 18a of the rear
cover 18 via seals 72 and 73, and a cylindrical part 62e extending
from the small diameter part 62c further extends to the interior of
a spring case 94 fixed to a rear face of the rear cover 18 via
bolts 93. The sliding member 70 is made of a material having
excellent durability, such as a super hard alloy or a ceramic. A
knock pin 74 implanted in the outer periphery of the rotary valve
main body 62 engages with a long hole 18b formed in the recess 18a
of the rear cover 18 in the axis L direction, and the rotary valve
main body 62 is therefore supported so that it can move in the axis
L direction but cannot rotate relative to the rear cover 18.
[0034] A plurality of preload springs 75 are supported within the
spring case 94 so as to surround the axis L, and a spring seat 95
receiving front ends of these preload springs 75 abuts against a
step 62d between the cylindrical part 62e and the small diameter
part 62c. The rotary valve main body 62 having the step 62d pushed
by the preload springs 75 is therefore biased forward so that the
sliding surfaces 68 of the stationary valve plate 63 and the
movable valve plate 64 are made to come into intimate contact with
each other. A steam supply pipe 77 having a heat-insulating space
77a in the interior thereof is fixed to the spring case 94 by a nut
96 so as to be positioned on the axis L. The steam supply pipe 77
is loosely inserted into the inner periphery 62f of the cylindrical
part 62e and the small diameter part 62c of the rotary valve main
body 62, and a tapered front end of the steam supply pipe 77 faces
an entrance end of a first steam passage P1 across a gap, the first
steam passage P1 being formed in the interior of the small diameter
part 62c of the rotary valve main body 62.
[0035] A plurality of annular sealing means 97 are disposed between
the outer periphery of the steam supply pipe 77 and the inner
periphery 62f of the cylindrical part 62e and the small diameter
part 62c of the rotary valve main body 62, and the rear end of the
annular sealing means 97 is fixed by a retaining member 98 screwed
into the inner periphery 62f. The sealing means 97 is a gland
packing formed from a material having excellent heat resistance
such as, for example, an inorganic fiber such as expanded graphite
carbon fiber, carbon fiber, or a metal fiber, or an organic fiber
such as a fluorine resin fiber or an aramid fiber, and can be
elastically deformed easily by an external force, thus allowing
relative movement between the rotary valve main body 62 and the
steam supply pipe 77.
[0036] The back of the sealing means 97 communicates with an
internal space of the spring case 94, and the internal space of the
spring case 94 communicates with a steam discharge chamber 90 via a
steam recovery passage 18e running through the cover 18.
[0037] A high-pressure stage steam intake route for supplying
high-temperature, high-pressure steam to the first axial piston
cylinder group 49 is shown in FIG. 11 by a mesh pattern. As is
clear from FIG. 11 together with FIG. 5 to FIG. 9, the first steam
passage P1 having its upstream end communicating with a pressure
chamber 76, to which the high-temperature, high-pressure steam is
supplied from the steam supply pipe 77, runs through the rotary
valve main body 62, opens on the surface at which the rotary valve
main body 62 is joined to the stationary valve plate 63, and
communicates with a second steam passage P2 running through the
stationary valve plate 63. In order to prevent the steam from
leaking past the surface of the rotary valve main body 62 which
joins the stationary valve plate 63, the joining surface is
equipped with a seal 81 (see FIG. 7 and FIG. 11), which seals the
outer periphery of a connecting part between the first and second
steam passages P1 and P2.
[0038] Seven third steam passages P3 (see FIG. 5) and seven fourth
steam passages P4 are formed respectively in the movable valve
plate 64 and the rotor 27 at circumferentially equal intervals, and
the downstream ends of the fourth steam passages P4 communicate
with the seven high-pressure operating chambers 82 defined between
the high-pressure cylinders 42 and the high-pressure pistons 43 of
the first axial piston cylinder group 49. As is clear from FIG. 6,
an opening of the second steam passage P2 formed in the stationary
valve plate 63 does not open evenly to the front and rear of the
top dead center (TDC) of the high-pressure pistons 43, but opens
displaced slightly forward in the direction of rotation of the
rotor 27, which is shown by the arrow R. This enables as long an
expansion period as possible, that is, a sufficient expansion
ratio, to be maintained, negative work, which would be generated if
the opening were set evenly to the front and rear of the TDC, to be
minimized and, moreover, the expanded steam remaining in the
high-pressure operating chambers 82 to be reduced, thus providing
sufficient output (efficiency).
[0039] A high-pressure stage steam discharge route and a
low-pressure stage steam intake route for discharging
medium-temperature, medium-pressure steam from the first axial
piston cylinder group 49 and supplying it to the second axial
piston cylinder group 57 are shown in FIG. 12 by a mesh pattern. As
is clear from FIG. 12 together with FIG. 5 to FIG. 8, an arc-shaped
fifth steam passage P5 (see FIG. 6) opens on a front face of the
stationary valve plate 63, and this fifth steam passage P5
communicates with a circular sixth steam passage P6 (see FIG. 7)
opening on a rear face of the stationary valve plate 63. The fifth
steam passage P5 opens from a position displaced slightly forward
in the direction of rotation of the rotor 27, which is shown by the
arrow R, relative to the bottom dead center (BDC) of the
high-pressure pistons 43 to a position displaced slightly backward
in the rotational direction relative to the TDC. This enables the
third steam passages P3 of the movable valve plate 64 to
communicate with the fifth steam passage P5 of the stationary valve
plate 63 over an angular range that starts from the BDC and does
not overlap the second steam passage P2 (preferably, immediately
before overlapping the second steam passage P2), and in this range
the steam is discharged from the third steam passages P3 to the
fifth steam passage P5.
[0040] Formed in the rotary valve main body 62 are a seventh steam
passage P7 extending in the axis L direction and an eighth steam
passage P8 extending in a substantially radial direction. The
upstream end of the seventh steam passage P7 communicates with the
downstream end of the sixth steam passage P6. The downstream end of
the eighth steam passage P8 communicates with a tenth steam passage
P10 running radially through the sliding member 70 via a ninth
steam passage P9 within a coupling member 83 disposed so as to
straddle the rotary valve main body 62 and the sliding member 70.
The tenth steam passage P10 communicates with the seven
low-pressure operating chambers 84 defined between the low-pressure
cylinders 50 and the low-pressure pistons 41 of the second axial
piston cylinder group 57 via seven eleventh steam passages P11
formed radially in the rotor 27.
[0041] In order to prevent the steam from leaking past the joining
surfaces of the rotary valve main body 62 and the stationary valve
plate 63, the outer periphery of a part where the sixth and seventh
steam passages P6 and P7 are connected is sealed by equipping the
joining surfaces with a seal 85 (see FIG. 7 and FIG. 12). Two seals
86 and 87 are disposed between the inner periphery of the sliding
member 70 and the rotary valve main body 62, and a seal 88 is
disposed between the outer periphery of the coupling member 83 and
the sliding member 70.
[0042] The interiors of the rotor 27 and the output shaft 28 are
hollowed out to define a pressure regulating chamber 89, and this
pressure regulating chamber 89 communicates with the eighth steam
passage P8 via a twelfth steam passage P12 and a thirteenth steam
passage P13 formed in the rotary valve main body 62, a fourteenth
steam passage P14 formed in the stationary valve plate 63, and a
fifteenth steam passage P15 running through the interior of the
bolt 67. The pressure of the medium-temperature, medium-pressure
steam discharged from the seven third steam passages P3 into the
fifth steam passage P5 pulsates seven times per revolution of the
rotor 27, but since the eighth steam passage P8, which is partway
along the supply of the medium-temperature, medium-pressure steam
to the second axial piston cylinder group 57, is connected to the
pressure regulating chamber 89, the pressure pulsations are damped,
steam at a constant pressure is supplied to the second axial piston
cylinder group 57, and the efficiency with which the low-pressure
operating chambers 84 are charged with the steam can be
enhanced.
[0043] Since the pressure regulating chamber 89 is formed by
utilizing dead spaces in the centers of the rotor 27 and the output
shaft 28, the dimensions of the expander M are not increased, the
hollowing out brings about a weight reduction effect and, moreover,
since the cuter periphery of the pressure regulating chamber 89 is
surrounded by the first axial piston cylinder group 49, which is
operated by the high-temperature, high-pressure steam, there is no
resultant heat loss in the medium-temperature, medium-pressure
steam supplied to the second axial piston cylinder group 57.
Furthermore, when the temperature of the center of the rotor 27,
which is surrounded by the first axial piston cylinder group 49,
increases, the rotor 27 can be cooled by the medium-temperature,
medium-pressure steam in the pressure regulating chamber 89, and
the resulting heated medium-temperature, medium-pressure steam
enables the output of the second axial piston cylinder group 57 to
be increased.
[0044] A steam discharge route for discharging the low-temperature,
low-pressure steam from the second axial piston cylinder group 57
is shown in FIG. 13 by a mesh pattern. As is clear from reference
to FIG. 13 together with FIG. 8 and FIG. 9, an arc-shaped sixteenth
steam passage P16 that can communicate with the seven eleventh
steam passages P11 formed in the rotor 27 is cut out in the sliding
surface 71 of the sliding member 70. This sixteenth steam passage
P16 communicates with a seventeenth steam passage P17 that is cut
out in an arc-shape in the outer periphery of the rotary valve main
body 62. The sixteenth steam passage P16 opens from a position
displaced slightly forward in the direction of rotation of the
rotor 27, which is shown by the arrow R, relative to the BDC of the
low-pressure pistons 51 to a position displaced rotationally
slightly backward relative to the TDC. This allows the eleventh
steam passages P11 of the rotor 27 to communicate with the
sixteenth steam passage P16 of the sliding member 70 over an
angular range that starts from the BDC and does not overlap the
tenth steam passage P10 (preferably, immediately before overlapping
the tenth steam passage P10), and in this range the steam is
discharged from the eleventh steam passages P11 to the sixteenth
steam passage P16.
[0045] The seventeenth steam passage P17 further communicates with
the steam discharge chamber 90 formed between the rotary valve main
body 62 and the rear cover 18 via an eighteenth steam passage P18
to a twentieth steam passage P20 formed within the rotary valve
main body 62 and a cutout 18d of the rear cover 18, and this steam
discharge chamber 90 communicates with a steam discharge hole 18c
formed in the rear cover 18.
[0046] As hereinbefore described, since the supply and discharge of
the steam to and from the first axial piston cylinder group 49 and
the supply and discharge of the steam to and from the second axial
piston cylinder group 57 are controlled by the common rotary valve
61, in comparison with a case in which separate rotary valves are
used for each the dimensions of the expander M can be reduced.
Moreover, since a valve for supplying the high-temperature,
high-pressure steam to the first axial piston cylinder group 49 is
formed on the flat sliding surface 68 on the front end of the
stationary valve plate 63, which is integral with the rotary valve
main body 62, it is possible to prevent effectively the
high-temperature, high-pressure steam from leaking. This is because
the flat sliding surface 68 can be machined easily with high
precision, and control of clearance is easier than for a
cylindrical sliding surface.
[0047] In particular, since the plurality of preload springs 75
apply a preset load to the rotary valve main body 62 and bias it
forward in the axis L direction, a surface pressure is generated on
the sliding surfaces 68 between the stationary valve plate 63 and
the movable valve plate 64, and it is thus possible to prevent
effectively the steam from leaking past the sliding surfaces 68.
Furthermore, even when the steam supply pipe 77 is moved in the
axis L direction due to vibration, etc., the movement is absorbed
by the sealing means 97, which is a gland packing, and will not be
transmitted to the rotary valve main body 62. It is therefore
possible to ensure the intimacy of contact between the sliding
surfaces 68 of the stationary valve plate 63 and the movable valve
plate 64, thus enabling the supply and discharge of steam to be
carried out reliably.
[0048] Because of the properties of the sealing means 97, which is
a gland packing, there is inevitably a small amount of leakage of
steam, and the steam passing through the sealing means 97 is
discharged into the steam discharge chamber 90 via the internal
space of the spring case 94 and the steam recovery passage 18e. In
this way, recovering the steam that has leaked past the sealing
means 97 enables loss of the working medium from the closed circuit
of the Rankine cycle system to be prevented, and the necessity for
replenishing the working medium can be minimized. Moreover, since
low temperature, low pressure steam that has leaked past the
sealing means 97 is recovered on the downstream side of the first
axial piston cylinder group 49 and the second axial piston cylinder
group 57, it is possible to prevent the output of the expander M
from being decreased by this steam.
[0049] Although a valve for supplying the medium-temperature,
medium-pressure steam to the second axial piston cylinder group 57
is formed on the cylindrical sliding surface 71 on the outer
periphery of the rotary valve main body 62, since the pressure of
the medium-temperature, medium-pressure steam passing through the
valve is lower than the pressure of the high-temperature,
high-pressure steam, the leakage of the steam can be suppressed to
a practically acceptable level by maintaining a predetermined
clearance without generating a surface pressure on the sliding
surface 71.
[0050] Furthermore, since the first steam passage P1 through which
the high-temperature, high-pressure steam passes, the seventh steam
passage P7 and the eighth steam passage P8 through which the
medium-temperature, medium-pressure steam passes, and the
seventeenth steam passage P17 to the twentieth steam passage P20
through which the low-temperature, low-pressure steam passes are
collectively formed within the rotary valve main body 62, not only
can the steam temperature be prevented from dropping, but also the
parts (for example, the seal 81) sealing the high-temperature,
high-pressure steam can be cooled by the low-temperature,
low-pressure steam, thus improving the durability.
[0051] Moreover, since the rotary valve 61 can be attached to and
detached from the casing main body 12 merely by removing the rear
cover 18 from the casing main body 12, the ease of maintenance
operations such as repair, cleaning, and replacement can be greatly
improved. Furthermore, although the temperature of the rotary valve
61 through which the high-temperature, high-pressure steam passes
becomes high, since the swash plate 39 and the output shaft 28,
where lubrication by oil is required, are disposed on the opposite
side to the rotary valve 61 relative to the rotor 27, the oil is
prevented from being heated by the heat of the rotary valve 61 when
it is at high temperature, which would degrade the performance in
lubricating the swash plate 39 and the output shaft 28. Moreover,
the oil can exhibit a function of cooling the rotary valve 61, thus
preventing overheating.
[0052] The operation of the expander M of the present embodiment
having the above-mentioned arrangement is now explained.
[0053] As shown in FIG. 11, high-temperature, high-pressure steam
generated by heating water in an evaporator is supplied to the
pressure chamber 76 of the expander M via the steam supply pipe 77,
and reaches the sliding surface 68 with the movable valve plate 64
via the first steam passage P1 formed in the rotary valve main body
62 of the rotary valve 61 and the second steam passage P2 formed in
the stationary valve plate 63 integral with the rotary valve main
body 62. The second steam passage P2 opening on the sliding surface
68 communicates momentarily with the third steam passage P3 formed
in the movable valve plate 64 rotating integrally with the rotor
27, and the high-temperature, high-pressure steam is supplied, via
the fourth steam passage P4 formed in the rotor 27, from the third
steam passage P3 to, among the seven high-pressure operating
chambers 82 of the first axial piston cylinder group 49, the
high-pressure operating chamber 82 that is present at the top dead
center.
[0054] Even after the communication between the second steam
passage P2 and the third steam passage P3 has been blocked due to
rotation of the rotor 27, the high-temperature, high-pressure steam
expands within the high-pressure operating chamber 82 and causes
the high-pressure piston 43 fitted in the high-pressure cylinder 42
of the sleeve 41 to be pushed forward from top dead center toward
bottom dead center, and the front end of the high-pressure piston
43 presses against the dimple 39a of the swash plate 39. As a
result, the reaction force that the high-pressure piston 43
receives from the swash plate 39 gives a rotational torque to the
rotor 27. For each one seventh of a revolution of the rotor 27, the
high-temperature, high-pressure steam is supplied into a fresh
high-pressure operating chamber 82, thus continuously rotating the
rotor 27.
[0055] As shown in FIG. 12, while the high-pressure piston 43,
having reached bottom dead center accompanying rotation of the
rotor 27, retreats toward top dead center, the medium-temperature,
medium-pressure steam pushed out of the high-pressure operating
chamber 82 is supplied to the eleventh steam passage P11
communicating with the low-pressure operating chamber 84 that,
among the low-pressure operating chambers 84 of the second axial
piston cylinder group 57, has reached top dead center accompanying
rotation of the rotor 27, via the fourth steam passage P4 of the
rotor 27, the third steam passage P3 of the movable valve plate 64,
the sliding surface 68, the fifth steam passage P5 and the sixth
steam passage P6 of the stationary valve plate 63, the seventh
steam passage P7 to the tenth steam passage P10 of the rotary valve
main body 62, and the sliding surface 71. Since the
medium-temperature, medium-pressure steam supplied to the
low-pressure operating chamber 84 expands within the low-pressure
operating chamber 84 even after the communication between the tenth
steam passage P10 and the eleventh steam passage P11 is blocked,
the low-pressure piston 51 fitted in the low-pressure cylinder 50
is pushed forward from top dead center toward bottom dead center,
and the link 52 connected to the low-pressure piston 51 presses
against the swash plate 39. As a result, the pressure force of the
low-pressure piston 51 is converted into a rotational force of the
swash plate 39 via the link 52, and this rotational force transmits
a rotational torque from the high-pressure piston 43 to the rotor
27 via the dimple 39a of the swash plate 39. That is, the
rotational torque is transmitted to the rotor 27, which rotates
synchronously with the swash plate 39. In order to prevent the
low-pressure piston 51 from becoming detached from the swash plate
39 when a negative pressure is generated during the expansion
stroke, the link 52 carries out the function of maintaining a
connection between the low-pressure piston 51 and the swash plate
39, and it is arranged that the rotational torque due to the
expansion is transmitted from the high-pressure piston 43 to the
rotor 27 rotating synchronously with the swash plate 39 via the
dimples 39a of the swash plate 39 as described above. For each one
seventh of a revolution of the rotor 27, the medium-temperature,
medium-pressure steam is supplied into a fresh low-pressure
operating chamber 84, thus continuously rotating the rotor 27.
[0056] During this process, as described above, the pressure of the
medium-temperature, medium-pressure steam discharged from the
high-pressure operating chambers 82 of the first axial piston
cylinder group 49 pulsates seven times for each revolution of the
rotor 27, but by damping these pulsations by the pressure
regulating chamber 89 steam at a constant pressure can be supplied
to the second axial piston cylinder group 57, thereby enhancing the
efficiency with which the low-pressure operating chambers 84 are
charged with the steam.
[0057] As shown in FIG. 13, while the low-pressure piston 51,
having reached bottom dead center accompanying rotation of the
rotor 27, retreats toward top dead center, the low-temperature,
low-pressure steam pushed out of the low-pressure operating chamber
84 is discharged into the steam discharge chamber 90 via the
eleventh steam passage P11 of the rotor 27, the sliding surface 71,
the sixteenth steam passage P16 of the sliding member 70, and the
seventeenth steam passage P17 to the twentieth steam passage P20 of
the rotary valve main body 62, and supplied therefrom into a
condenser via the steam discharge hole 18c.
[0058] When the expander M operates as described above, since the
seven high-pressure pistons 43 of the first axial piston cylinder
group 49 and the seven low-pressure pistons 51 of the second axial
piston cylinder group 57 are connected to the common swash plate
39, the outputs of the first and second axial piston cylinder
groups 49 and 57 can be combined to drive the output shaft 28,
thereby achieving a high output while reducing the dimensions of
the expander M. During this process, since the seven high-pressure
pistons 43 of the first axial piston cylinder group 49 and the
seven high-pressure pistons 51 of the second axial piston cylinder
group 57 are displaced by half a pitch in the circumferential
direction, as shown in FIG. 10, pulsations in the output torque of
the first axial piston cylinder group 49 and pulsations in the
output torque of the second axial piston cylinder group 57 are
counterbalanced, thus making the output torque of the output shaft
28 flat.
[0059] Furthermore, although axial type rotary fluid machines
characteristically have a high space efficiency compared with
radial type rotary fluid machines, by arranging two stages in the
radial direction the space efficiency can be further enhanced. In
particular, since the cylinders of the first axial piston cylinder
group 49, which are required to have only a small diameter because
they are operated by high-pressure steam having a small volume, are
arranged on the radially inner side, and the cylinders of the
second axial piston cylinder group 57, which are required to have a
large diameter because they are operated by low-pressure steam
having a large volume, are arranged on the radially outer side, the
space can be utilized effectively, thus making the expander M still
smaller. Moreover, since the cylinders 42 and 50 and the pistons 43
and 51 that are used have circular cross sections, which enables
machining to be carried out with high precision, the amount of
steam leakage can be reduced in comparison with a case in which
vanes are used, and a yet higher output can thus be
anticipated.
[0060] Furthermore, since the first axial piston cylinder group 49,
which is operated by high-temperature steam, is arranged on the
radially inner side, and the second axial piston cylinder group 57,
which is operated by low-temperature steam, is arranged on the
radially outer side, the difference in temperature between the
second axial piston cylinder group 57 and the outside of the casing
11 can be minimized, the amount of heat released outside the casing
11 can be minimized, and the efficiency of the expander M can be
enhanced. Moreover, since the heat escaping from the
high-temperature first axial piston cylinder group 49 on the
radially inner side can be recovered by the low-temperature second
axial piston cylinder group 57 on the radially outer side, the
efficiency of the expander M can be further enhanced.
[0061] Moreover, when viewed from an angle perpendicular to the
axis L, since the rear end of the first axial piston cylinder group
49 is positioned forward relative to the rear end of the second
axial piston cylinder group 57, the heat escaping rearward in the
axis L direction from the first axial piston cylinder group 49 can
be recovered by the second axial piston cylinder group 57, and the
efficiency of the expander M can be yet further enhanced.
Furthermore, since the sliding surfaces 68 on the high-pressure
side is present deeper within the recess 27b of the rotor 27 than
the sliding surfaces 71 on the low-pressure side, the difference in
pressure between the outside of the casing 11 and the sliding
surfaces 71 on the low-pressure side can be minimized, the amount
of leakage of steam past the sliding surfaces 71 on the
low-pressure side can be reduced and, moreover, the pressure of
steam leaking past the sliding surfaces 68 on the high-pressure
side can be recovered by the sliding surfaces 71 on the
low-pressure side and utilized effectively.
[0062] Although an embodiment of the present invention is explained
above, the present invention can be modified in a variety of ways
without departing from the spirit and scope thereof.
[0063] For example, the operating part of the present invention is
not limited to the axial piston cylinder groups of the embodiment,
and a radial piston cylinder type or a vane type may be
employed.
INDUSTRIAL APPLICABILITY
[0064] The rotary fluid machine of the present invention can be
applied suitably to an expander that employs steam as a working
medium, but it can also be applied to a compressor that compresses
a compressible fluid such as air or a pump that feeds by pressure
an incompressible fluid such as oil or water.
* * * * *