JP2002155707A - Rankine cycle apparatus - Google Patents

Abstract

PROBLEM TO BE SOLVED: To make full use of the performance of an expansion device and a condenser of a Rankine cycle apparatus. SOLUTION: Relative to an arbitrary relation between the pressure Pevp and temperature Tevp of steam sucked by the expansion device 4 having a first stage of cylinder chamber and a second stage of vane chamber disposed in series, an expansion ratio of steam sucked/discharged by the expansion device 4 is set at a prescribed expansion ratio εcorresponding to the arbitrary relation, so that the pressure Rexp2 and the temperature Texp2 of steam discharged by the expansion device 4 are made to meet a target value, thereby making full use of performance of the expansion device 4 and the condenser 5. Steam in a cylinder chamber at a first stage contains no water because it is within an overheated steam area, therefore no water-hammering phenomenon occurs in the cylinder chamber. Steam at an outlet of the vane chamber contains water because it is within a saturated steam area, therefore lubrication and sealing of the vane chamber can be attained.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an evaporator for heating a liquid to generate vapor, an expander for expanding a vapor supplied from the evaporator to output a constant axial torque, and an expander for discharging the vapor. And a condenser for cooling the recovered steam into a liquid.

[0002]

2. Description of the Related Art Japanese Patent Application Laid-Open No. 4-47104 discloses a Rankine cycle device in which an expander is operated with steam generated by an evaporator, and the steam discharged from the expander is liquefied by a condenser and returned to the evaporator. Open and close the valve provided at the inlet of the expander according to the magnitude of the energy of the steam generated by the evaporator,
It describes that the maximum output torque is ensured by controlling the timing of supplying steam to the expander.

[0003] Also, JP-A-58-48706 discloses that
In a Rankine cycle device that operates the expander with the steam generated by the evaporator and liquefies the steam discharged by the expander with the condenser and returns the vapor to the evaporator, the pressure at which the steam is introduced into the condenser is the vapor discharge pressure from the expander. When the pressure is higher than that, a bypass passage connecting the inlet side of the condenser and a position immediately before the completion of the expansion of the expander is opened to reduce excessive expansion loss of the expander.

Japanese Patent Application Laid-Open No. 61-87990 discloses that
A vane type compressor is described in which a rotary valve for controlling intake and exhaust to and from a vane chamber is provided on a rotary shaft of a rotor supporting a vane, and the intake timing and exhaust timing of the rotary valve are made variable.

[0005]

For example, water is heated by an evaporator that performs heat exchange with exhaust gas of an internal combustion engine to generate steam, and the steam is used to operate a positive displacement expander to drive a shaft. In a Rankine cycle device that takes out the output, returns the steam discharged from the expander to water with a condenser, and supplies it to the evaporator again, the pressure and temperature of the steam supplied from the evaporator to the expander depend on the performance of the expander. The temperature of steam discharged from the expander to the condenser is preset as a rated value corresponding to the performance of the condenser. However, the pressure and temperature of the steam generated in the evaporator fluctuate according to the transient state of the evaporator, the operating state of the internal combustion engine, the amount of water supplied to the evaporator, and the like, and the condenser exhibits the maximum performance. The pressure and temperature of the obtained steam fluctuate according to the transient state of the condenser, the cooling state of the condenser (outside air temperature, the rotation speed of the cooling fan, the intensity of the traveling wind), and the like.

In FIG. 21 (A), the vertical axis and the horizontal axis represent the pressure p and the specific volume v of the steam, respectively. The steam having the rated pressure p1 at the inlet of the expander is generated inside the expander. When the expander expands by the preset expansion ratio ε and the outlet pressure of the expander changes from the aforementioned p1 to the rated value p2, the expander and the condenser can exhibit the maximum performance. However, as described above, the inlet pressure of the expander varies due to various factors, and the outlet pressure of the expander at which the expander and the condenser can exhibit the maximum performance also varies according to various factors. Therefore, the outlet pressure of the expander may not match the pressure at which the expander and the condenser can exhibit the maximum performance at that time, and the expander and the condenser may not be able to exhibit sufficient performance.

That is, as shown in FIG. 21B, even if the expansion ratio matches the set expansion ratio ε, if the inlet pressure of the expander is p1 ′ which is larger than the rated value p1, the expansion The outlet pressure of the expander becomes higher than the rated value p2, the steam that still has the energy for driving the expander is wasted, and the performance of the expander cannot be fully exhibited. However, there is a problem that the load on the device increases and the condensation performance decreases. On the other hand, as shown in FIG. 21 (C), even if the expansion ratio matches the set expansion ratio ε, if the inlet pressure of the expander is p1 ′ smaller than the rated value p1, the outlet of the expander Since the pressure is lower than the rated value p2, there is a problem that the steam performs a negative work inside the expander and the output is reduced.

[0008] Such a problem occurs when the inlet temperature of the expander is higher or lower than the rated value, when the amount of steam leakage inside the expander is large or small, or when the expander and the condenser have the maximum performance. This also occurs when the outlet pressure of the expander that can be exerted fluctuates from the rated value p2 due to various factors.

The present invention has been made in view of the above-mentioned circumstances, and has as its object to maximize the performance of an expander and a condenser of a Rankine cycle device.

[0010]

In order to achieve the above object, according to the first aspect of the present invention, an evaporator for heating a liquid to generate a vapor and a vapor supplied from the evaporator are provided. In a Rankine cycle device including an expander that expands to output an axial torque, and a condenser that cools steam discharged by the expander and returns it to a liquid, an arbitrary relationship between the pressure and temperature of steam sucked by the expander On the other hand, by setting the expansion ratio of the steam sucked / discharged by the expander to a predetermined expansion ratio according to the above-mentioned arbitrary relationship, the pressure and temperature of the steam discharged by the expander are made to match the target values. A featured Rankine cycle device is proposed.

According to the above configuration, even if the pressure and the temperature of the steam sucked by the expander have an arbitrary relationship, the expansion ratio of the steam sucked / discharged by the expander is set to a predetermined expansion ratio according to the arbitrary relationship. By setting the ratio, the pressure and temperature of the steam discharged from the expander can be controlled. Therefore, if the expansion ratio is set with the target value of the pressure and temperature at which the expander and the condenser can exhibit the maximum performance, the pressure and temperature of the steam discharged by the expander are made to match the target values, and The performance of the condenser can be maximized.

According to the second aspect of the present invention,
2. The Rankine cycle device according to claim 1, wherein the pressure and the temperature of the steam sucked by the expander are in a superheated steam region, and the pressure and the temperature of the steam discharged by the expander are in a saturated steam region. Is proposed.

According to the above configuration, the steam sucked by the expander is in the superheated steam region containing no liquid, and the steam discharged by the expander is in the saturated steam region containing the liquid. The load on the condenser that returns the vapor to the liquid can be reduced while minimizing the effect.

According to the third aspect of the present invention,
2. The Rankine cycle apparatus according to claim 1, wherein the expander includes a plurality of expansion chambers connected in series, and a product of a steam expansion ratio in each expansion chamber is the set expansion ratio. Is proposed.

According to the above configuration, a plurality of expansion chambers are connected in series to integrate and output the axial torque generated by each expander, and the product of the expansion ratio of steam in each expansion chamber is set to the set expansion ratio. As a result, the condensation efficiency of the condenser can be maximized.

According to the fourth aspect of the present invention,
In addition to the configuration of claim 3, at least the steam in the most upstream expansion chamber of the plurality of expansion chambers of the expander is in the superheated steam area, and at least the steam in the most downstream expansion chamber is in the saturated steam area. A Rankine cycle device characterized by this is proposed.

According to the above configuration, the vapor of at least the most upstream expansion chamber of the plurality of expansion chambers is in the superheated steam region containing no liquid, and at least the most downstream expansion chamber of the plurality of expansion chambers. Is in the saturated vapor region including the liquid, so that the load on the condenser that returns the vapor to the liquid can be reduced while minimizing the effect of the liquid on the operation of the expander.

According to the invention described in claim 5,
In addition to the configuration of claim 4, a Rankine cycle device is proposed in which the expansion chamber in which the steam at the discharge position is in the superheated steam region is constituted by a cylinder chamber.

According to the above configuration, since the steam is in the superheated steam region at the discharge position of the expansion chamber formed of the cylinder chamber, it is possible to prevent the liquid from being mixed with the steam, and to cause the liquid to stay in the cylinder chamber. The troubles that occur can be avoided beforehand.

According to the invention described in claim 6,
In addition to the configuration of claim 4, a Rankine cycle device is proposed in which the expansion chamber in which the steam at the discharge position is in the saturated steam region is constituted by a vane chamber.

According to the above configuration, since the steam is in the saturated steam region at the discharge position of the expansion chamber formed by the vane chamber, the liquid is mixed with the steam to improve the lubricity and sealing property of the vane with the liquid. be able to.

According to the seventh aspect of the present invention,
In addition to the configuration of claim 3, a Rankine cycle device is proposed in which at least the suction position of the most upstream expansion chamber of the plurality of expansion chambers of the expander is variable.

According to the above arrangement, the suction position of at least the most upstream expansion chamber of the plurality of expansion chambers is made variable, so that the pressure of the steam sucked by the expander is changed to expand the entire expander. The ratio can be varied from the set expansion ratio. Accordingly, even if the pressure and temperature of the steam sucked by the expander deviate from the predetermined relationship, the expander discharges by changing the expansion ratio of the steam sucked and discharged by the expander from the set expansion ratio. The pressure and temperature of the steam can be matched to the target values.

According to the invention described in claim 8,
In addition to the configuration of claim 3, a Rankine cycle device is proposed, wherein at least the discharge position of the most downstream expansion chamber among the plurality of expansion chambers of the expander is variable.

According to the above arrangement, the discharge position of at least the most downstream expansion chamber of the plurality of expansion chambers is made variable, so that the pressure of the steam discharged from the expander is changed and the expansion of the entire expander is performed. The ratio can be varied from the set expansion ratio. Accordingly, even if the pressure and temperature of the steam sucked by the expander deviate from the predetermined relationship, the expander discharges by changing the expansion ratio of the steam sucked and discharged by the expander from the set expansion ratio. The pressure and temperature of the steam can be matched to the target values.

The cylinder member 39 of the embodiment forms the expansion chamber and the cylinder chamber of the present invention, and the vane chamber 5 of the embodiment.
4 constitutes the expansion chamber of the present invention.

[0027]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS In FIG. 1, a waste heat recovery apparatus 2 for an internal combustion engine 1 uses a waste heat of the internal combustion engine 1, for example, exhaust gas as a heat source to raise the temperature from a high-pressure liquid, for example, water. An evaporator 3 that generates high-pressure steam, that is, high-temperature and high-pressure steam, an expander 4 that generates an output by expansion of the high-temperature and high-pressure steam, and the expanded temperature and pressure discharged from the expander 4 The evaporator 3 includes a condenser 5 for liquefying the steam that has dropped, that is, the temperature-reduced and reduced-pressure steam, and a supply pump 6 that pressurizes and supplies the liquid, for example, water from the condenser 5 to the evaporator 3.

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

2 to 7, the casing 7 is composed of first and second metal halves 8 and 9. Both halves 8, 9
Consists of a main body 11 having a substantially elliptical concave portion 10 and a circular flange 12 integral with the main body 11. A substantially elliptical rotor chamber 14 is formed by overlapping the two circular flanges 12 via a metal gasket 13. Is done.
The outer surface of the main body 11 of the first half 8 is covered by a deep bowl-shaped main body 16 of a shell-shaped member 15, and a circular flange 17 integral with the main body 16 is attached to the circular flange 12 of the first half 8 by a gasket. The three circular flanges 12, 12, 17 are superimposed via 18, and are fastened by bolts 19 at a plurality of positions in the circumferential direction. Thereby, both the main body 1 of the shell-shaped member 15 and the first half 8 are
A relay chamber 20 is formed between 1 and 16.

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

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

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

In each slot-like space 34, first to twelfth vane piston units U1 to U12 having the same structure are mounted so as to be reciprocally movable in the radial direction as follows. In the substantially U-shaped space 34, a stepped hole 38 is formed in a portion 37 defining the inner peripheral side thereof, and a stepped cylinder member 39 made of ceramic (or carbon) is fitted into the stepped hole 38. . The end surface of the small diameter portion a of the cylinder member 39 abuts on the outer peripheral surface of the large diameter portion 24 of the output shaft 23, and the small diameter hole b communicates with the through hole c opened on the outer peripheral surface of the large diameter portion 24. Further, a guide cylinder 40 is arranged outside the cylinder member 39 so as to be located coaxially with the member 39. The outer end of the guide cylinder 40 is engaged with the opening of the space 34 existing on the outer peripheral surface 36 of the rotor 31, and the inner end is formed with a large-diameter hole d of a stepped hole 38.
And comes into contact with the cylinder member 39. The guide cylinder 40 has a pair of long grooves e extending opposite to each other from the outer end to the vicinity of the inner end, and both long grooves e face the space 34. A piston 41 made of ceramic is slidably fitted in the large-diameter cylinder hole f of the cylinder member 39, and the distal end side of the piston 41 is always located in the guide cylinder 40.

As shown in FIG. 2 and FIG.
The cross section B of the rotor chamber 14 in the imaginary plane A including the rotation axis L is a pair of semicircular cross sections B1 having diameters g facing each other, and one of the two diameters g of the two semicircular cross sections B1 facing each other. It has a square cross section B2 formed by connecting the ends and the other opposing ends to each other, and has a substantially track shape for competition. In FIG. 8, the portion indicated by the solid line indicates the maximum cross section including the major axis, while the portion partially indicated by the two-dot chain line indicates the minimum cross section including the minor axis. The rotor 31 has a cross section D slightly smaller than the minimum cross section including the minor diameter of the rotor chamber 14, as shown by the dotted line in FIG.

As shown in FIGS. 2 and 9 to 12, the vane 42 has a substantially U-shaped (horshoe-shaped) vane body 43 and a substantially U-shaped plate attached to the vane body 43. It comprises a seal member 44 and a vane spring 58.

The vane body 43 includes a semicircular portion 46 corresponding to the inner circumferential surface 45 of the semicircular cross section B1 of the rotor chamber 14, and a pair of parallel portions 48 corresponding to the opposing inner end surface 47 of the rectangular cross section B2. Having. A U-shaped notch 49 at the end side of each parallel portion 48, a rectangular blind hole 50 opening at the bottom surface thereof, and a further protruding outward at the end side than each notch 49. A short axis 51 is provided. Further, U-shaped grooves 52 that open outward are formed in a series on the outer peripheral portions of the semicircular portion 46 and the two parallel portions 48, and both ends of the U-shaped grooves 52 communicate with the two notches 49, respectively. . Further, a pair of ridges 53 each having a partially circular cross section are provided on both flat portions of the semicircular portion 46. Double ridge 53
Indicates that the axis L1 of the virtual cylinder formed by the
And divide the semi-circular portion 46 into two in the circumferential direction.
It is arranged so as to coincide with a straight line that divides equally. The inner ends of the projections 53 slightly project into the space between the parallel portions 48.

The seal member 44 is made of, for example, PTFE, and has a semicircular cross section B of the rotor chamber 14.
1 and a pair of parallel portions 56 that slide on the opposing inner end surface 47 by the square cross section B2. Further, a pair of elastic claws 57 are provided on the inner peripheral surface side of the semicircular portion 55 so as to bend inward.

A seal member 44 is mounted in the U-shaped groove 52 of the vane body 43, and a vane spring 5 is provided in each blind hole 50.
8 are fitted, and a roller 59 having a ball bearing structure is attached to each short shaft 51. Each vane 42 is slidably housed in each slot-shaped space 34 of the rotor 31, and at this time, both ridges 53 of the vane main body 43 are in the guide cylinder 40, and both side portions of the two ridges 53 are guides. It is located in each of the two long grooves e of the cylinder 40, so that the inner end surfaces of the two ridges 53 can contact the outer end surface of the piston 41. Both rollers 59 are rotatably engaged with non-circular annular grooves 60 formed on the opposed inner end faces 47 of the first and second halves 8, 9, respectively. The distance between the annular groove 60 and the rotor chamber 14 is constant over their entire circumference. The forward movement of the piston 41 is controlled by the roller 5 via the vane 42.
The rotation of the rotor 31 is converted by the engagement between the groove 9 and the annular groove 60.

With the cooperation of the roller 59 and the annular groove 60,
As clearly shown in FIG. 5, the semi-circular portion 4 of the vane body 43
6, the semicircular tip surface 61 is formed from the inner peripheral surface 45 of the rotor chamber 14, and both parallel portions 48 are formed in the rotor chamber 1.
4 are always separated from the opposing inner end surfaces 47, thereby reducing friction loss. And 2
Since the trajectory is regulated by the annular groove 60 composed of a pair of strips, the vane 42 is caused to rotate by a small displacement angle in the axial direction via the roller 59 due to a right and left trajectory error, and the inner circumferential surface 45 of the rotor chamber 14 is rotated. Increase the contact pressure with At this time, in the vane body 43 having a substantially U-shaped plate shape (horse-shoe shape), the displacement in the radial direction of the contact portion with the casing 7 can be greatly reduced as compared with the rectangular (rectangular) vane. Further, as clearly shown in FIG.
, The two parallel portions 56 are connected to the respective vane springs 58.
Due to the resiliency of the rotor chamber 14, it closely adheres to the opposing inner end surface 47 of the rotor chamber 14, and in particular, seals the annular groove 60 passing between the ends of the two parallel portions 56 and the vane 42. Further, the semicircular arc-shaped portion 55 is brought into close contact with the inner peripheral surface 45 when the two elastic claws 57 are pressed between the vane body 43 and the inner peripheral surface 45 in the rotor chamber 14. That is, square (rectangular)
Since the substantially U-shaped plate-shaped vane 42 has no inflection point with respect to the vane, the close contact is good. The square vanes have corners, and it is difficult to maintain the sealing performance. Thereby, the vane 42
In addition, the sealing property between the rotor chambers 14 is improved.
Further, with thermal expansion, the vane 42 and the rotor chamber 14 are deformed. At this time, since the substantially U-shaped vane 42 is more uniformly deformed with a similar shape to the rectangular vane, the variation in the clearance between the vane 42 and the rotor chamber 14 is small, and the sealing property can be maintained well.

The sealing action between the vane body 43 and the inner peripheral surface 45 of the rotor chamber 14 is effected by the spring force of the seal member 44 itself, the centrifugal force acting on the seal member 44 itself, and the high pressure side rotor chamber 14. The steam that has entered the U-shaped groove 52 of the vane body 43 is generated by the steam pressure that pushes up the seal member 44. Thus, the sealing action is
Since there is no influence of excessive centrifugal force acting on the vane body 43 according to the rotation speed of the rotor 31, the sealing surface pressure does not depend on the centrifugal force applied to the vane body 43, and always has good sealing performance and low friction property. And can be compatible.

As described above, the twelve vanes 42 radially supported by the rotor 31, the inner peripheral surface 45 of the rotor chamber 14, and the outer peripheral surface 36 of the rotor 31 increase the volume as the rotor 31 rotates. 12 changing vane chambers 5
4 (see FIG. 4).

2 and 3, the large-diameter portion 24 of the output shaft 23 has a thick portion 62 supported by the bearing metal 25 of the second half 9, and extends from the thick portion 62 to the first half. 8
And a thin portion 63 supported by the bearing metal 25. A hollow shaft 64 made of ceramic (or metal) is fitted into the thin portion 63 so as to rotate integrally with the output shaft 23. A fixed shaft 65 is disposed inside the hollow shaft 64, and the fixed shaft 65 includes a large-diameter solid portion 66 fitted to the hollow shaft 64 so as to fit within the axial thickness of the rotor 31, and an output shaft. A small-diameter solid portion 69 fitted through two seal rings 68 into a hole 67 existing in the thick portion 62 of the 23, and a thin-walled portion extending from the large-diameter solid portion 66 and fitted into the hollow shaft 64. And a hollow portion 70. A seal ring 71 is interposed between the outer peripheral surface of the end of the hollow portion 70 and the inner peripheral surface of the hollow bearing cylinder 21 of the first half 8.

A hollow cylindrical body 72 is screwed to the left end of the fixed shaft 65 via a seal ring 73. A shaft 75 protruding to the left of the hollow cylindrical body 72 has a bearing provided at the center of the shell-shaped member 15. The sliding portion between the hollow cylindrical body 72 and the shell-shaped member 15 is sealed by a seal ring 74. The distal end of the inner tube portion 77 extending rightward from the hollow cylinder 72 is fitted into a stepped hole h in the large-diameter solid portion 66 of the fixed shaft 65 together with a short hollow connection tube 78 protruding therefrom. The high-temperature and high-pressure steam introduction pipe 80 which is press-fitted into the shaft part 75 projecting to the left side of the hollow cylindrical body 72 and extends rightward in the inner pipe part 77 has its right end fitted into the hollow connection pipe 78. A driven gear 79 is formed on the outer periphery of the shaft portion 75 of the hollow cylindrical body 72, and a driving gear 83 provided on the rotating shaft of the motor 82 meshes with the driven gear 79. Accordingly, when the motor 82 is driven, the fixed shaft 65 rotates via the drive gear 83, the driven gear 79, and the hollow cylinder 72, and the output shaft 2
3 can be generated.

As shown in FIGS. 2 to 4 and FIG. 13, the hollow shaft 64 and the output shaft are mounted on the large-diameter solid portion 66 of the fixed shaft 65, the cylinder member 39 of the first to twelfth vane piston units U1 to U12. The high-temperature and high-pressure steam is supplied through a plurality of, in this embodiment, 12 through-holes c formed in series on the shaft 23, and the first temperature-reduced and reduced-pressure steam after expansion from the cylinder member 39 is passed through the through-hole c. A rotary valve V for discharging the air is provided as follows.

FIG. 13 shows each cylinder member 39 of the expander 4.
2 shows a structure of a rotary valve V for supplying and discharging steam at a predetermined timing. In the large-diameter solid portion 66, first and second holes 86 and 87 extending in opposite directions from a space 85 communicating with the hollow connection pipe 78 are formed. The first and second holes 86 and 87 are formed. The first and second concave portions 88 and 89 open on the outer peripheral surface of the large-diameter solid portion 66 and open on the bottom surfaces. 1st, 2nd
First and second carbon seal blocks 92 and 93 having supply ports 90 and 91 are mounted in the recesses 88 and 89, and their outer peripheral surfaces rub against the inner peripheral surface of the hollow shaft 64. Short first and second supply pipes 94 and 95, which are coaxial, are loosely inserted into the first and second holes 86 and 87, and the first and second supply pipes 94 and 9 are inserted.
5, the first and second seal cylinders 96 fitted to the outer peripheral surface on the tip side,
The tapered outer peripheral surface i, j of 97 is located inside the supply ports 90, 91 of the first and second seal blocks 92, 93 and fits into the inner peripheral surface of the tapered holes k, m connected thereto. The large-diameter solid portion 66 includes first and second annular concave portions n and o surrounding the first and second supply pipes 94 and 95, and first and second adjacent concave portions n and o.
The first and second seal blocks 92,
93, the first and second annular concave portions n, o
The first and second bellows-like elastic bodies 98 and 99 having one end sides fitted to the outer peripheral surfaces of the first and second seal cylinders 96 and 97, respectively. First and second coil springs 100 and 101 are respectively housed therein, and first and second coil springs 100 and 101 are accommodated therein.
The first and second seal blocks 92 and 93 are pressed against the inner peripheral surface of the hollow shaft 64 by the elastic force of the bellows-like elastic bodies 98 and 99 and the first and second coil springs 100 and 101.

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

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

The inside of the hollow portion r of the fixed shaft 65 and the hollow cylindrical body 7
2 is a passage s for the first temperature-reduced pressure reducing steam, and the passage s
Communicates with the relay chamber 20 through a plurality of through holes t penetrating the peripheral wall of the hollow cylindrical body 72.

As shown in FIG. 2, FIG. 5, FIG. 6 and FIG. 7, a plurality of radially-introduced introduction parts are provided in the outer periphery of the main body 11 of the first half 8 near both ends of the minor axis of the rotor chamber 14. First and second rows of introduction holes 107 and 10 comprising holes 106
8 is formed, and the first temperature-reduced and reduced-pressure steam in the relay chamber 20 is introduced into the rotor chamber 14 through the introduction hole rows 107 and 108. In the outer peripheral portion of the main body 11 of the second half 9, the second introduction hole row 108 of the rotor chamber 14 is formed.
On the upstream side, there are formed nine rows of first outlet holes 110a to 110i each including a plurality of outlet holes 109 arranged in the radial direction, and on the upstream side of the first inlet hole row 107, a plurality of first outlet holes 110a to 110i. Nine second outgoing hole arrays 111a to 111i are formed. Nine first outlet hole arrays 110a to 110i and second outlet hole arrays 111a to 111a
11i are arranged in the circumferential direction with a predetermined phase difference, and the five lead-out holes 109 in each row communicate with each other through the communication path 116.

The first five rows of outlet holes 110e to 110e on the downstream side
Four first outlet hole rows 110a to 110a on the upstream side except for 10i
The four communication holes 116 of 10d have the communication holes 116
First solenoid valves 117a to 117 that can be individually opened and closed
d are respectively provided, and the five communication holes 116 of the four second outlet holes 111a to 111d on the upstream side excluding the five second outlet holes 111e to 111i on the downstream side are connected to the communication holes 116, respectively. Four second solenoid valves 1 capable of individually opening and closing 116
18a to 118d are provided respectively. Then, the first lead hole rows 110a to 110d and the second lead hole row 111a
A pressure sensor 119 is provided in each of a total of eight communication passages 116 to 111d.

The output shaft 23 and the like are lubricated by water, and the lubricating channel is configured as follows. That is, as shown in FIGS. 2 and 3, a water supply pipe 113 is connected to a water supply hole 112 formed in the hollow bearing cylinder 22 of the second half 9. The water supply hole 112 is provided in the bearing metal 2 on the second half 9 side.
5 facing the housing 114 and the housing 11
Reference numeral 4 denotes a water passage hole u formed in the thick portion 62 of the output shaft 23.
Further, the water passage holes u are formed in a plurality of water passage grooves v (see also FIG. 13) extending in the direction of the generatrix of the outer peripheral surface of the hollow shaft 64. 25 communicates with the housings 115 facing each other. An annular recess w is provided on the inner end face of the thick portion 62 of the output shaft 23 to communicate the water passage u with the sliding portion between the hollow shaft 64 and the large-diameter solid portion 66 of the fixed shaft 65. .

Thus, the space between each bearing metal 25 and the output shaft 23 and the space between the hollow shaft 64 and the fixed shaft 65 are lubricated with water.
The casing 7, the sealing member 44 and each roller 5
9 is performed.

In FIG. 4, the first and seventh vane piston units U1 and U7, which are in point symmetry with respect to the rotation axis L of the rotor 31, perform the same operation. This is because the second and eighth vane piston units U having a point symmetrical relationship
The same applies to 2, U8 and the like.

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

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

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

As shown in FIGS. 14 and 15, the expander 4 having a positive displacement and a constant axial torque comprises a first expansion chamber comprising a cylinder chamber of a cylinder member 39 and a second expansion chamber comprising a vane chamber 54. And The pressure and temperature of the steam supplied to the first expansion chamber are Pevp and Tevp, respectively, and the pressure and temperature of the steam supplied to the second expansion chamber are Pexp1 and Texp1, respectively. Pressure and temperature of steam
2, Texp2, Pevp and Pexp1
Expansion ratio ε1 of the first expansion chamber determined by
And expansion ratio ε of the second expansion chamber determined by Pexp2 and Pexp2
The total expansion ratio of the expander 4 given by the product ε1 × ε2 of 2 matches the preset expansion ratio ε (132 in this embodiment). The vertical axis of the graph in FIG. 15 is the steam pressure P, and the horizontal axis is the phase θ of the rotor 31. Pressure P
When the pressure adjusted to Pevp is supplied to the first expansion chamber, where the steam expands and the pressure P decreases to Pexp1, the expansion ratio determined by Pevp and Pexp1 is ε1. When the steam having a pressure P of Pexp1 is supplied to the second expansion chamber, where the steam expands and the pressure P decreases to Pexp2, the expansion ratio determined by Pxvp and Pexp2 is ε2.

The steam generated by the evaporator 3, that is, the steam supplied to the first expansion chamber, is controlled such that the pressure Pevp and the temperature Tevp maintain a predetermined relationship shown by a solid line in FIG. That is, the pressure Pevp and the temperature Tevp of the steam supplied to the first expansion chamber are in the transient state of the evaporator 3,
The pressure Pevp of the steam varies depending on the operation state of the internal combustion engine 1, the amount of water supplied to the evaporator 3, and the like, and can be controlled by the rotation speed (shaft torque) of the expander 4, and the steam temperature Tv
evp can be controlled by the amount of water supplied to the evaporator 3, and the rated value in this embodiment is point a on the solid line in FIG. 16 (pressure Pevp = 16 MPa, temperature Tevp = 620).
° C). If the pressure Pevp and the temperature Tevp of the steam supplied to the first expansion chamber are determined in this way, the shaft torque of the expander 4 is also determined accordingly. As shown by the broken line in FIG. 16, the higher the pressure Pevp of the steam supplied to the first expansion chamber and the higher the temperature Tevp, the higher the thermal efficiency. However, if the temperature Tevp is high, the durability and the like are affected. In the example, the rated value is set to 620 ° C. On the other hand, the pressure value Pexp2 and the temperature Texp2 of the steam discharged from the second expansion chamber are also set to the rated values at which the expander 4 and the condenser 5 can exhibit the maximum performance. Pressure Pexp2 is 0.
05MPa, temperature Tevp is 80 ° C. However, the optimal pressure Pexp2 and the optimal temperature Texp2 change according to the transient state of the condenser 5, the cooling state of the condenser 5 (outside air temperature, the number of rotations of the cooling fan, the intensity of the running wind), and the like. Not necessarily.

The pressure Pe of the steam supplied to the first expansion chamber
vp and temperature Tevp are rated values (Pevp = 16MP
a, Tevp = 620 ° C.), and the expansion ratio of the expander 4 is set to the set expansion ratio ε, the pressure Pexp2 and the temperature Texp2 of the steam discharged from the second expansion chamber are rated values (this embodiment). Then Pexp2 = 0.05MPa, Tex
p2 = 80 ° C.), and the expander 4 and the condenser 5 can exhibit the maximum performance. Further, even if the pressure Pevp and the temperature Tevp of the steam supplied to the first expansion chamber deviate from the rated values, the steam is located at any position on the solid line in FIG. 16 and the expansion ratio becomes the set expansion ratio ε = 132. If they match, the pressure Pe of the steam discharged from the second expander chamber is Pe.
xp2 and temperature Texp2 correspond to the rated values. Accordingly, when the internal combustion engine 1 is in the warm-up operation and the pressure Pevp and the temperature Tevp of the steam supplied to the first expansion chamber are lower than the rated values (for example, point b on the solid line in FIG. 16). However, the pressure Pe of the steam discharged from the second expansion chamber
xp2 and temperature Texp2 correspond to the rated values. As a result, the startup time from the start of the internal combustion engine 1 until the Rankine cycle device becomes operable can be reduced.

As described above, the pressure Pevp and the temperature Tevp of the steam supplied to the first expansion chamber are set so as to have a predetermined relationship (the relationship indicated by the solid line in FIG. 16), and the expansion ratio of the expander 4 Is set to the set expansion ratio ε, the pressure Pevp of the steam discharged from the second expansion chamber and the temperature Tev
p is always a rated value (Pexp2 = 0.05M in this embodiment)
Pa, Texp2 = 80 ° C.).
And the condenser 5 can exhibit the maximum performance.

By the way, when the pressure Pevp and the temperature Tevp of the steam supplied to the first expansion chamber deviate from the relationship shown by the solid line in FIG. Remains at the set expander ratio ε,
The pressure Pexp2 and the temperature Texp2 of the steam discharged from the second expansion chamber deviate from the rated values, and
In addition, there is a possibility that the condenser 5 cannot exhibit sufficient performance. Also, the pressure P of the steam discharged from the second expansion chamber
When the optimum values of the exp2 and the temperature Texp2 deviate from the rated values due to various fluctuation factors, the pressure Pexp2 of the steam discharged from the second expansion chamber if the expander ratio of the expander 4 remains at the set expansion ratio ε. And the temperature Texp2 becomes the rated value and deviates from the optimum value, and the expander 4 and the condenser 5 may not be able to exhibit sufficient performance.

In such a case, by changing the expansion ratio of the expander 4 from the set expansion ratio ε, the pressure Pexp2 and the temperature Tex of the steam discharged from the second expansion chamber are changed.
p2 can be matched to the optimal value. The expansion ratio of the expander 4 can be changed by changing the suction timing into the first expansion chamber or by changing the discharge timing from the second expansion chamber.

Specifically, when the pressure Pevp of the steam supplied to the first expansion chamber is excessive (FIG. 21B)
), The timing of supplying steam to the first expansion chamber may be delayed to reduce the expansion ratio ε1, and if the pressure Pevp of the steam supplied to the first expansion chamber is too small (see FIG. 21 (C)), the timing of discharging steam from the second expansion chamber may be advanced to reduce the expansion ratio ε2.

The expansion ratio ε1 of the steam in the first expansion chamber can be changed by changing the suction timing of the steam by the rotary valve V. That is, the fixed shaft 65 is rotated by the motor 82, and the phases of the supply ports 90 and 91 are changed to the retard side in FIG. 13 to adjust the timing at which steam is supplied from the evaporator 3 to the cylinder member 39 of the expander 4. If it advances earlier, since the piston 41 is located inward in the radial direction at the moment when steam is introduced, and the volume of the cylinder member 39 is reduced, the cylinder member 3
The amount of steam supplied to 9 decreases, and the expansion ratio ε1 of the first expansion chamber (cylinder member 39) of expander 4 increases.
Conversely, when the fixed shaft 65 is rotated by the motor 82, the phases of the supply ports 90 and 91 are changed to the advanced side in FIG. 13 to supply steam from the evaporator 3 to the cylinder member 39 of the expander 4. Is delayed, the moment the steam is introduced, the piston 41 is located radially outward and the volume of the cylinder member 39 increases, the amount of steam supplied to the cylinder member 39 increases, and the first The expansion ratio ε1 of the expansion chamber (cylinder member 39) decreases. Thus, by changing the timing at which steam is introduced into the first expansion chamber, the expansion ratio ε1 can be changed.

Since the steam discharged from the first expansion chamber is supplied to the second expansion chamber (vane chamber 54) via the relay chamber 20, the amount of steam discharged from the first expansion chamber is equal to the second expansion chamber. It matches the steam supply to the chamber. The timing at which steam is discharged from the second expansion chamber to the condenser 5 is determined by eight solenoid valves 11.
It is controlled by selectively opening and closing 7a to 117d and 118a to 118d. For example, in FIG. 7, at a position slightly before the position where the vane chamber 54 has the maximum volume, the vanes 42 on the rotation direction leading side of the pair of vanes 42 forming the vane chamber 54 are arranged in the first discharge hole row in the third row. 11
0c, and this position is the reference timing.
That is, at the time of rating, the two first discharge hole rows 11 on the upstream side
0a, 110b solenoid valves 117a, 117b are closed,
The solenoid valves 117c and 117d of the two first discharge hole rows 110c and 110d on the downstream side are opened, and therefore, the moment when the vane 42 on the rotational direction advancing side exceeds the third first discharge hole row 110c. Then, the discharge of steam from the first discharge hole row 110c is started.

In order to advance the discharge timing with respect to the reference timing, the first discharge hole row 110 in the second row on the upstream side is used.
In order to further advance the discharge timing, the first discharge hole row 1 of the first row on the upstream side in addition to the solenoid valve 117b of the second discharge hole row 110b of the second row may be opened.
What is necessary is just to open the solenoid valve 117a of 10a. Conversely, to delay the discharge timing with respect to the reference timing, the solenoid valve 117c of the first discharge hole row 110c in the third row only needs to be closed. In addition to the solenoid valve 117c of the first discharge hole array 110c, the solenoid valve 117d of the fourth discharge hole array 110d on the downstream side may be closed.

The solenoid valves 117a to 117c closed in this manner
By sequentially increasing the number of 117d from the upstream side, the timing at which steam is discharged from the second expansion chamber to the condenser 5 can be delayed stepwise, thereby increasing the expansion ratio ε2 by the second expansion chamber. Can be increased. Conversely, by sequentially increasing the number of solenoid valves 117a to 117e to be opened from the downstream side, the timing at which steam is discharged from the second expansion chamber to the condenser 5 can be advanced stepwise, Thereby, the expansion ratio ε2 by the second expansion chamber can be reduced.

The control of the second solenoid valves 118a to 118d of the second lead hole rows 111a to 111d is performed by controlling the solenoid valves 117a to 117d of the first lead hole rows 110a to 110d.
Is the same as the control of Further, the solenoid valves 117a to 117
d, 118a to 118d are controlled by eight rows of outlet holes 11
The pressure Pexp2 of the steam discharged from the second expansion chamber based on the outputs of the eight pressure sensors 119 provided in correspondence with 0a to 110d and 111a to 111d, respectively.
Is performed so as to match the optimum value at which the expander 4 and the condenser 5 can exhibit the maximum performance.

The state of steam determined by pressure, volume, and temperature includes a saturated steam region in which water and steam are mixed, and a superheated steam region in which only steam exists without water. The region from the inlet to the outlet of the first expansion chamber is a superheated steam region, and water is not mixed in the steam.
Accordingly, it is possible to surely prevent the water staying inside the cylinder member 39 constituting the first expansion chamber from being compressed by the piston 41 and causing a water hammer phenomenon. At least the most downstream portion of the region from the inlet to the outlet of the second expansion chamber is a saturated steam region, and water is mixed in the steam. Therefore, some water stays inside the vane chamber 54 constituting the second expansion chamber, and the vane 42 and the rotor chamber 14
The lubrication performance and sealing performance during the operation are improved.

17 to 19, the temperature Tevp of the steam supplied to the first expansion chamber is changed from 450 ° C. to 650 ° C.
When the temperature Tevp is higher, the superheated steam region inside the expander 4 becomes wider and the timing of transition from the superheated steam region to the saturated steam region is delayed (see FIG. 17), and the amount of decrease in enthalpy is increased. As a result, the output of the expander 4 increases (see FIG. 18), and the dryness at the outlet of the second expansion chamber increases, so that the amount of generated water decreases (see FIG. 19).
Conversely, the lower the temperature Tevp, the narrower the superheated steam region inside the expander 4 becomes, and the earlier the transition from the superheated steam region to the saturated steam region is made, so that the amount of decrease in enthalpy decreases and the output of the expander 4 decreases. In addition, the dryness at the outlet of the second expansion chamber decreases, and the amount of generated water increases. The boundary between the first expansion chamber and the second expansion chamber is in the superheated steam area, so that water is reliably prevented from staying in the first expansion chamber including the cylinder member 39, and the vane chamber 54 is included. It can be ensured that water stays in the second expansion chamber.

The detected temperature T at the entrance of the first expansion chamber
When evp is higher than the rated value, the pressure Pexp2 at the outlet of the second expansion chamber becomes higher than the rated value.
Delays the suction timing of the inlet of the expansion chamber at the expansion ratio ε1
May be reduced, or the expansion ratio ε2 may be increased by delaying the discharge timing of the outlet of the second expansion chamber. Conversely, if the detected temperature Tevp at the inlet of the first expansion chamber is lower than the rated value, the pressure Pe at the outlet of the second expansion chamber
Since xp2 is lower than the rated value, the suction ratio at the inlet of the first expansion chamber is advanced to increase the expansion ratio ε1, or the discharge timing at the outlet of the second expansion chamber is delayed to decrease the expansion ratio ε2. You can do it.

When the amount of leakage inside the expander 4 is large (during low-speed rotation), the same variable expansion ratio control as when the temperature Tevp at the inlet of the first expansion chamber is higher than the rated value is performed. On the contrary, when the leak amount inside the expander 4 is small (during high-speed rotation), the same variable expansion as in the case where the temperature Tevp at the inlet of the first expansion chamber is lower than the rated value is performed. What is necessary is just to perform ratio control.

Next, a second embodiment of the present invention will be described with reference to FIG.

In the expander 4 of the first embodiment, first, high-temperature and high-pressure steam is supplied to the cylinder member 39 serving as the first expansion chamber, and then the first temperature-reduced pressure-reduced steam whose temperature has been lowered is reduced to the second expansion chamber. Is supplied to the vane chamber 54. On the other hand, the second embodiment shown in FIG. 20 has a through hole t for discharging the first reduced-temperature and reduced-pressure steam from the first expansion chamber to the relay chamber 20.
Can be closed by the solenoid valve 122, and the switching valve 12
At 0, the supply of the high-temperature and high-pressure steam to the first expansion chamber is cut off so that the high-temperature and high-pressure steam can be directly supplied to the steam inlet 121 of the relay chamber 20, thereby disabling the first expansion chamber. Only the second expansion chamber can be operated independently. In this case, the expansion ratio of steam in the expansion chamber constituted by the vane chamber 54 depends on the exhaust timing from the vane chamber 54 and the electromagnetic valves 117a to 117d and 118a to 118.
It is controlled by changing d.

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

For example, in the embodiment, the first expansion chamber and the second expansion chamber are connected in series, but three or more stages of expansion chambers can be connected in series. In this case, the steam supplied to the most upstream expansion chamber needs to be in the superheated steam area, and the steam discharged from the most downstream expansion chamber needs to be in the saturated steam area.

[0077]

As described above, according to the first aspect of the present invention, even if the pressure and the temperature of the steam sucked by the expander have an arbitrary relationship, the expansion of the steam sucked and discharged by the expander is performed. By setting the ratio to a predetermined expansion ratio according to the arbitrary relationship, the pressure and temperature of the steam discharged from the expander can be controlled. Therefore, if the expansion ratio is set with the target value of the pressure and temperature at which the expander and the condenser can exhibit the maximum performance, the pressure and temperature of the steam discharged by the expander are made to match the target values, and The performance of the condenser can be maximized.

According to the second aspect of the present invention,
Since the steam inhaled by the expander is in the superheated steam region that does not contain liquid, and the steam that is discharged by the expander is in the saturated steam region that contains liquid, the influence of the liquid on the operation of the expander is minimized. The load on the condenser that returns the vapor to the liquid can be reduced.

According to the third aspect of the present invention,
While connecting multiple expansion chambers in series and integrating and outputting the shaft torque generated by each expander, the product of the expansion ratio of steam in each expansion chamber is set as the expansion ratio to maximize the condensation efficiency of the condenser. Can be increased.

According to the fourth aspect of the present invention,
At least the vapor of the most upstream expansion chamber of the plurality of expansion chambers is in a superheated steam region not containing a liquid, and the vapor of at least the most downstream expansion chamber of the plurality of expansion chambers is a saturated steam region containing a liquid. Therefore, the load on the condenser for returning the vapor to the liquid can be reduced while the influence of the liquid on the operation of the expander is minimized.

According to the invention described in claim 5,
Since the steam is in the superheated steam area at the discharge position of the expansion chamber constituted by the cylinder chamber, it is possible to prevent the liquid from being mixed with the steam and to avoid a problem caused by the liquid remaining in the cylinder chamber. Can be.

According to the invention described in claim 6,
Since the steam is in the saturated steam region at the discharge position of the expansion chamber constituted by the vane chamber, it is possible to mix the liquid with the steam to improve the lubricity and sealing property of the vane with the liquid.

According to the invention described in claim 7,
By changing the suction position of at least the most upstream expansion chamber of the plurality of expansion chambers, the pressure of steam sucked by the expander is changed to change the expansion ratio of the entire expander from the set expansion ratio. Can be. Accordingly, even if the pressure and temperature of the steam sucked by the expander deviate from the predetermined relationship, the expander discharges by changing the expansion ratio of the steam sucked and discharged by the expander from the set expansion ratio. The pressure and temperature of the steam can be matched to the target values.

According to the invention described in claim 8,
By changing the discharge position of at least the most downstream expansion chamber of the plurality of expansion chambers, the pressure of steam discharged by the expander is changed to change the expansion ratio of the entire expander from the set expansion ratio. Can be. Accordingly, even if the pressure and temperature of the steam sucked by the expander deviate from the predetermined relationship, the expander discharges by changing the expansion ratio of the steam sucked and discharged by the expander from the set expansion ratio. The pressure and temperature of the steam can be matched to the target values.

[Brief description of the drawings]

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

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

FIG. 3 is an enlarged sectional view around a rotation axis of FIG. 2;

FIG. 4 is a sectional view taken along line 4-4 in FIG. 2;

FIG. 5 is an enlarged sectional view taken along line 5-5 of FIG.

FIG. 6 is an enlarged sectional view taken along line 6-6 of FIG. 5;

FIG. 7 is an enlarged view of a main part of FIG. 5;

FIG. 8 is an enlarged view around the rotation axis of FIG. 4;

FIG. 9 is a front view of the vane body.

FIG. 10 is a side view of the vane body.

FIG. 11 is a sectional view taken along line 11-11 of FIG. 9;

FIG. 12 is a front view of a seal member.

FIG. 13 is an enlarged view around the rotation axis of FIG. 4;

FIG. 14 is a diagram showing a configuration of a control system of the expander.

FIG. 15 is a graph showing pressure changes and expansion ratios of first and second expansion chambers.

FIG. 16 is a graph showing an optimal relationship between the temperature and the pressure at the inlet of the expander.

FIG. 17 is a TS diagram of a Rankine cycle device.

FIG. 18 is an HS diagram of a Rankine cycle device.

FIG. 19 is a graph showing the relationship between the temperature at the outlet of the expander and the dryness.

FIG. 20 is a diagram showing a second embodiment of the present invention.

FIG. 21 is a graph showing changes in steam pressure and specific volume in an expander.

[Explanation of symbols]

 Reference Signs List 3 evaporator 4 expander 5 condenser Pevp pressure of steam sucked by expander Tevp temperature of steam sucked by expander Pexp2 pressure of steam sucked by expander Texp2 temperature of steam sucked by expander ε1 steam expansion ratio ε2 Steam expansion ratio ε Set expansion ratio

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Tsuneo Endo 1-4-1 Chuo, Wako-shi, Saitama Prefecture Inside Honda R & D Co., Ltd. (72) Tsutomu Takahashi 1-4-1 Chuo, Wako-shi, Saitama Inside Honda R & D Co., Ltd. (72) Inventor Kensuke Honma 1-4-1 Chuo, Wako-shi, Saitama Pref.

Claims (8)

[Claims] An evaporator (3) for heating a liquid to generate steam, an expander (4) for expanding steam supplied from the evaporator (3) and outputting an axial torque, and an expander (4). In a Rankine cycle device provided with a condenser (5) for cooling the steam discharged by 4) and returning it to a liquid, an arbitrary relationship between the pressure (Pevp) and the temperature (Tevp) of the steam sucked by the expander (4) In contrast, the expansion ratio (ε1, ε) of the steam sucked and discharged by the expander (4)
By setting 2) to a predetermined expansion ratio (ε) corresponding to the arbitrary relationship, it is possible to make the pressure (Pexp2) and the temperature (Texp2) of the steam discharged from the expander (4) coincide with the target values. Characterized Rankine cycle device. 2. The pressure (P) of steam inhaled by the expander (4).
evp) and temperature (Tevp) are in the superheated steam region, and the pressure (Pexp2) of the steam discharged from the expander (4).
The Rankine cycle device according to claim 1, wherein the temperature and the temperature (Texp2) are in a saturated steam region. 3. The expander (4) includes a plurality of expansion chambers (39, 54) connected in series, each expansion chamber (39, 5).
2. The Rankine cycle apparatus according to claim 1, wherein the product of the steam expansion ratios (ε1, ε2) in 4) is set to the set expansion ratio (ε). 3. 4. A plurality of expansion chambers (39, 5) of an expander (4).
In 4), at least steam in the most upstream expansion chamber (39) is in a superheated steam area, and at least steam in the most downstream expansion chamber (54) is in a saturated steam area. 4. The Rankine cycle device according to 3. 5. The Rankine cycle device according to claim 4, wherein the expansion chamber (39) in which the steam at the discharge position is in the superheated steam region is constituted by a cylinder chamber. 6. The Rankine cycle device according to claim 4, wherein the expansion chamber in which the steam at the discharge position is in a saturated steam region is constituted by a vane chamber. 7. A plurality of expansion chambers (39, 5) of an expander (4).
The Rankine cycle device according to claim 3, wherein at least the suction position of the expansion chamber (39) on the most upstream side is variable. 8. A plurality of expansion chambers (39, 5) of an expander (4).
4. The Rankine cycle device according to claim 3, wherein in 4), at least the discharge position of the most downstream expansion chamber (54) is variable.