JP4854068B2 - High pressure fluid engine and closed cycle engine - Google Patents

Description

  The present invention relates to a cat-and-mouse type non-uniform speed rotation conversion mechanism that gives a slow relative motion to a piston in an annular cylinder to generate energy, and a high-pressure fluid engine and a closed cycle engine using the same.

  2. Description of the Related Art Conventionally, a so-called cat-and-mouse non-uniform speed rotation conversion mechanism that generates energy by giving a slow relative movement to a piston in an annular cylinder is known. As a representative example, a pseudo ellipse / elliptical gear, a three-side link system, and the like are known as known transmissions (see Patent Document 1 and Patent Document 2).

  Although it is applied to various other devices, the disadvantage of the transmission mechanism described above is that high load transmission (near the maximum load acting on the piston) is possible if it is a pseudo-elliptical oval gear (oval gear). In this case, the load applied to the teeth is one type of force because the load direction is slanted with respect to the meshing direction on the pitch circle, especially in the vicinity of the constant speed of meshing with each other compared to the normal helical gear and straight gear. There is a problem of strength remaining due to the large force of the wedge.

The farther the focal point of the ellipse of the gear becomes, that is, the harder the ellipse becomes, the more the strength acting on the tooth becomes a problem, and there is a problem such as abnormal wear of the tooth or breakage when it is severe. Therefore, in order to make the gear suitable for the transmission load, the size of the engine, that is, the size of the piston body, will eventually become larger, and the problem of moment of inertia will be highlighted due to the inconstant speed transmission. Dynamic problems arise. In practice, a compact device as expressed in the drawing of Patent Document 2 is not conceivable. In addition, there is a disadvantage that the manufacturing cost increases due to an increase in the size of the apparatus .

  Furthermore, when the slow relative motion is a three-sided link system, both ends of the connecting rod connecting the crank pin and the crank pin are rotated in reverse rotation, so that there is a problem that it is difficult to achieve dynamic balance. Further, it is necessary to add another transmission element to eliminate the dead point. However, if the machining accuracy is poor, the dead point cannot be eliminated and a machining accuracy equivalent to that of the gear is required. Also, as a common problem, a lubrication device is essential, and the complexity of the device cannot be avoided.

  On the other hand, as a conventional cat-and-mouse engine, for example, the one described in Patent Document 3 is known. In general, a gear is used at a constant speed rotation location. Therefore, the strength of a gear is determined by statically determining a force acting from a transmitted torque. And finally, the specifications are determined in consideration of wear and accuracy. However, for example, in the cat-and-mouse engine described in Patent Document 1, the impact load due to combustion that the rotary piston receives, that is, the force applied to the gear due to a sudden shift becomes very large. It is extremely difficult to construct.

  Regarding the relationship between the curve and the pin for controlling the slow speed relationship of the rotary piston, it has been very difficult to cope with the speed. The current car cam shape is a beautiful foliage shape, but to increase engine performance, the cam acceleration increases drastically just by raising the cam lift by just 1 mm. As a result, the rod-like push rod in the valve system changes to a spring system that contracts and expands, causing a problem of valve jumping in which the valve is separated from the cam profile. For this reason, in order to increase the speed of current car cams, a so-called DOHC (direct over head camshaft) system in which push rods are eliminated is currently the mainstream.

  In particular, it is considered that a cat-and-mouse engine is established from a so-called mechanistic viewpoint unrelated to the acting force. However, in order to force the pin to move according to the groove in consideration of the force received by the piston, the load acting on the contact point between the pin and the groove becomes a magnitude that cannot be measured. The contact between the curve and the pin should theoretically be a point contact or a line contact, and therefore, it will result in abnormal wear due to the load acting on it, and in severe conditions it will lead to breakage due to insufficient strength. For example, in Patent Document 3 described above, even if an attempt is made to make the pin and groove corresponding to the load of the piston larger than the piston cylinder, it is actually impossible in structure.

  Furthermore, if the fuel energy of a car engine is assumed to be 100%, the effective utilization energy is 30%, the exhaust gas energy is 32%, the energy for cooling the engine is 28%, and the friction and radiation loss is 10% (internal combustion) Institutional Lecture (first volume), 12th edition, written by Fujio Nagao, Yokendo Co., Ltd. (published in 1963), 60% of energy is actually wasted in the air from the muffler and radiator. If such loss energy can be used effectively, energy can be used efficiently and energy can be saved.

The inventors of the present invention have reached the present invention as a result of intensive studies to develop a novel inconstant speed rotation conversion mechanism that solves the above problems, and a high-pressure fluid engine and a closed cycle engine using the same. is there.
JP 53-59907 A JP 54-72310 A JP 2001-241303

  That is, an object of the present invention is to solve the problem of balance of dynamic balance, abnormal wear due to a load applied to a gear and damage due to insufficient strength, and a compact and inexpensive non-constant speed rotation conversion mechanism and a high pressure using the same. It is to provide a fluid engine and a closed cycle engine.

In order to solve the above-described problems, a high-pressure fluid engine according to the present invention is a crankpin guide that passes through the rotation center of a constant-speed rotation output shaft that is eccentric with respect to the first crankshaft and the second crankshaft having the same rotation center the means, in the high-pressure fluid engine using the first crankshaft and the first crank pin and a second, respectively a crank pin slidably guiding and non-uniform rotary converting mechanism comprising mounted respectively to the second crankshaft A cylinder for installing therein a first disk rotor having a first piston and a second disk rotor having a second piston, and dead center positions of the first piston and the second piston in the cylinder In addition, a fluid inflow hole and a fluid exhaust hole respectively provided at the location of the cylinder corresponding to When the first piston and the second piston are rotated by the crank pin guide means serving as a flywheel, the first crank pin and the second crank pin are reciprocated while sliding on the crank pin guide means. As a result, the first piston and the second piston have a slow relationship, and the slower of the first piston and the second piston shuts off the communication with the fluid inlet hole and the fluid outlet hole. The faster piston, which is in communication with the fluid inflow hole, allows high-pressure fluid to flow in. The higher piston receives the expansion pressure of the high-pressure fluid to obtain an output, and at the same time, it flows in and expands. by discharging extruding the finished fluid, especially that the constant speed output shaft is caused to continuously rotate To.

In addition, this inconstant speed rotation conversion mechanism is a first crankshaft that transmits power and performs a rotational motion, a first disk rotor attached to the first crankshaft, and a first disk rotor, The first piston for rotating the first crankshaft, the first crankpin attached to the first crankshaft, and the same center of rotation as the first crankshaft, transmitting power and rotating A second crankshaft that moves, a second disk rotor that is provided opposite to the first disc rotor and is attached to the second crankshaft, and is attached to the second disc rotor, and A second piston for rotating the shaft, and a second crank mounted on the second crankshaft. A crank pin guide means that slidably guides the crank pin, the first crank pin and the second crank pin, and is rotatably provided, the rotation center of the first crank shaft and the second crank shaft A crankpin guide means whose center of rotation is eccentric by a predetermined distance, and a constant speed rotation output shaft attached to the crankpin guide means for outputting rotational motion from the crankpin guide means. When the first piston and the second piston are rotated, the first crank pin and the second piston are reciprocated while sliding on the crank pin guide means. Of the first piston and the second piston. There Repeat slow at a predetermined angle delay may be the constant rotation output shaft is configured to be allowed to rotate continuously.

Further, high-pressure fluid engine of the present invention, Ru high pressure fluid engine der using non-uniform rotary conversion mechanism described above.

  According to the configuration described above, the problem of wear resistance is solved because the crank pin is brought into rolling contact and surface contact by a lubrication method such as ball bearings and a slider. In addition, the problem of strength can be considered in the same way as a conventional reciprocating engine, so that the experience cultivated in the reciprocating engine is utilized as it is.

  Examples of the high-pressure fluid include high-pressure gas, water vapor, and air, but any high-pressure fluid may be used, and a high-pressure liquid or the like can also be used.

  Furthermore, the closed cycle engine of the present invention is a closed cycle engine using the high-pressure fluid engine described above, and an exhaust gas release engine that discharges exhaust gas, and heat that generates water vapor with the exhaust gas of the exhaust gas release engine. An exchanger, and a condenser for cooling the exhaust fluid from the high-pressure fluid engine to return it to water, connecting the heat exchanger to the fluid inlet of the high-pressure fluid engine, and the condensate A closed cycle structure is formed by connecting a vessel to the fluid discharge hole of the high-pressure fluid engine.

  The closed cycle engine of the present invention means an engine having a closed cycle structure, but by configuring as described above, it is possible to recycle energy that has been conventionally lost by cycling exhaust gas → water vapor → water. It becomes possible.

  Further, as the closed cycle engine, a temperature sensor is attached to a cylinder head and a cylinder of the cylinder of the exhaust gas emission engine, and water is generated at the same time as water is sprayed and cooled to a high temperature portion of the cylinder according to information of the temperature sensor. It is preferable to be configured to operate using

According to this inconstant speed conversion mechanism, and a high-pressure fluid engine and a closed cycle engine using the same, the balance of dynamic balance, the problem of abnormal wear due to the load on the gears, and the damage due to insufficient strength are eliminated, and the compact There is a great effect that it does not cost.

  In addition, the inconstant speed rotation conversion mechanism, and the high-pressure fluid engine and the closed cycle engine using the same can be particularly preferably used for a positive displacement engine and a compressor.

  Furthermore, the high-pressure fluid engine of the present invention can be miniaturized and rotated at a high speed, can have high thermal efficiency, and can be used as a power source as an external combustion engine.

  Moreover, according to the closed cycle engine of this invention, there exists an effect that it becomes possible to use effectively the loss energy of 60% which was conventionally made into a loss.

  Embodiments of the present invention will be described below with reference to the accompanying drawings. However, the illustrated examples are merely illustrative, and various modifications other than the illustrated examples are possible without departing from the technical idea of the present invention. Needless to say.

Figure 1 is a perspective view showing one embodiment of the present non-uniform rotary conversion mechanism. FIG. 2 is an exploded view of FIG. FIG. 3 is a schematic explanatory view showing the amount of change in rotation of the crankpin in the inconstant speed rotation conversion mechanism of FIG. FIG. 4 shows one embodiment of the high-pressure fluid engine of the present invention, in which (a) is a cross-sectional view and (b) is an enlarged schematic view of IV (b) portion of (a). FIG. 5 is a cross-sectional view taken along line VV in FIG. FIG. 6 shows one embodiment of the high-pressure fluid engine of the present invention, where (a) is a view taken in the direction of arrow Y in FIG. 5, and (b) is a sectional view taken along line VI (b) -VI (b) in (a). It is.

  FIG. 7 shows one embodiment of a high-pressure fluid engine, (a) is a magnified view of one embodiment of a counterweight used in the high-pressure fluid engine of the present invention, and (b) is VII of (a). (B) It is a sectional view taken along the line -VII (b). FIG. 8 shows one embodiment of the high-pressure fluid engine of the present invention, in which (a) is a sectional view, (b) is a sectional view taken along line VIII (b) -VIII (b) of (a), and (c). FIG. 2 is a partially cutaway schematic perspective view of (a). FIG. 9 shows one embodiment of the seal of the high-pressure fluid engine of the present invention, where (a) is a sectional view and (b) is a sectional view taken along line IX (b) -IX (b) in (a). .

  FIG. 10 is an explanatory diagram of a mechanism that controls the slow-speed movement of the piston in the high-pressure fluid engine of the present invention. FIG. 11 is an operational state diagram of the piston in the high-pressure fluid engine of the present invention. FIG. 12 is an operation diagram in the case where a fluid inflow hole of a type giving priority to thermal efficiency is provided in the high-pressure fluid engine of the present invention. FIG. 13 is an explanatory view of a mechanism for changing the position of the opening edge in the high-pressure fluid engine of the present invention. 14A and 14B show another embodiment of the high-pressure fluid engine of the present invention, where FIG. 14A is a cross-sectional view, FIG. 14B is a cross-sectional view of FIG. 14A, and FIG. is there.

  FIG. 15 shows another embodiment of the high-pressure fluid engine of the present invention, in which (a) is a side view of (b), (b) is a sectional view, and (c) is a side view of (b). FIG. 16 is a schematic view showing one embodiment of a closed cycle engine of the present invention. FIG. 17 is a schematic view showing another embodiment of the closed cycle engine of the present invention. FIG. 18 is a graph showing the relative angular velocities of the first crankpin and the second crankpin. FIG. 19 is a graph showing the relationship between pressure and volume.

One embodiment of this inconstant speed rotation conversion mechanism is shown in FIGS. 1-3, the code | symbol 10 shows this inconstant speed rotation conversion mechanism. As shown well in FIGS. 1 and 2, the non-uniform speed rotation conversion mechanism 10 includes a first crankshaft 12 that transmits power and performs a rotational motion, and a first disk rotor attached to the first crankshaft 12. 14 and a first piston 16 attached to the first disk rotor 14 for rotating the first crankshaft 12 and a first piston 16 attached to the first crankshaft 12 via a first crank arm 18. One crank pin 20, a second crankshaft 22 having the same rotation center O1 as the first crankshaft 12, and transmitting rotational power to rotate, and provided opposite to the first disk rotor 14. And a second disk rotor 24 attached to the second crankshaft 22 and a second disk rotor 24. A second piston 26 for rotating the second crankshaft 22, a second crankpin 30 attached to the second crankshaft 22 via a second crank arm 28, and the first crank Crank pin guide means that slidably guides the pin 20 and the second crank pin 30 and is rotatably provided, from the rotation center O 1 of the first crank shaft 12 and the second crank shaft 22. A crankpin guide means 32 provided with a center of rotation O2 decentered by a predetermined distance; a constant speed rotation output shaft 36 attached to the crankpin guide means 32 for outputting a rotational motion from the crankpin guide means 32; , Is composed of.

  Since the crankpin guide means 32 is only required to guide the first crankpin 20 and the second crankpin 30 slidably, there is no particular limitation on the configuration thereof. In the illustrated example, the crank pin guide means 32 is provided with a guide groove 34, and the first crank pin 20 and the second crank pin 30 are inserted into the guide groove 34 in a slip-fit state so that the inside of the guide groove 34. It is slidable.

  Further, the rotation center O1 of the first crankshaft 12 and the second crankshaft 22 and the rotation center O2 of the crankpin guide means 32 are eccentric by an eccentricity e. The amount of eccentricity is not particularly limited because it can be appropriately set according to the purpose and application.

  With the configuration described above, when the first piston 16 and the second piston 26 are rotated, the first crank pin 20 and the second crank pin 30 are reciprocated while sliding on the crank pin guide means 32. The first piston 16 and the second piston 26 have a slow speed relationship, and the first piston 16 and the second piston 26, whichever can be turned faster, turns faster and outputs the constant speed rotation output. The shaft 36 is continuously rotated.

Then one of the embodiments of the present non-uniform speed rotation conversion mechanism 10 will be described with reference to FIG.

  In FIG. 3A, the guide groove 34 of the crankpin guide means 32 is vertical, and the first crankpin 20 and the second crankpin 30 are in relation to a horizontal line passing through the rotation center O2 of the crankpin guide means 32. Each is in a symmetrical position.

  When the guide groove 34 rotates 45 degrees clockwise from the state of FIG. 3A, the second crankpin 30 rotates 58 degrees and the first crankpin 20 rotates 28 degrees (FIG. 3B). When the guide groove 34 further rotates 45 degrees to the right from the state of FIG. 3B, the second crankpin 30 rotates 73 degrees and the first crankpin 20 rotates 17 degrees (FIG. 3C).

  When the guide groove 34 further rotates 45 degrees to the right from the state of FIG. 3C, the first crank pin 20 rotates 17 degrees and the second crank pin 30 rotates 73 degrees (FIG. 3D). When the guide groove 34 further rotates 45 degrees to the right from the state of FIG. 3D, the second crankpin 30 rotates 58 degrees, the first crankpin 20 rotates 28 degrees, and the guide groove 34 becomes vertical again (FIG. 3). (E)).

  3 (a) to 3 (e), it can be seen that the crankpin guide means 32 rotates 180 degrees, and the positions of the first crankpin 20 and the second crankpin 30 are interchanged during that time. As is apparent from FIG. 3, the first crank pin 20 rotates 98 degrees (see FIG. 3A) while the second crank pin 30 rotates 262 degrees. That is, the second crankpin 30 turns fast and the first crankpin 20 turns late. From FIG. 3 (e), the crank pin guide means 32 further rotates 180 degrees to complete one cycle, and the second crank pin 30 and the first crank pin 20 each make one rotation.

  Regarding the strength problem, since the first crankpin 20 and the second crankpin 30 are in contact with the guide groove 34 that is a flat surface at a right angle, the force acting on the first crankpin 20 and the second crankpin 30 is considered. Thus, a simple load design is possible.

  For dynamic balance, by providing a counterweight or the like on the opposite side of the first crankpin 20 and the second crankpin 30, the balance weight of the first crank arm 18 and the second crank arm 28 is completely balanced. Is possible.

  As will be described later, the problem of lubrication can be solved by fitting ball bearings, shield ball bearings, sliders, and the like to the first crankpin 20 and the second crankpin 30.

Thus, above according to the present non-uniform rotary conversion mechanism, there is an advantage that can be resolved at a stroke the problem engine and compressor Cat and Mouse scheme faced.

Next, an embodiment of the high pressure fluid engine using the present non-uniform rotary converting mechanism shown in FIGS. 4 to 15 and FIGS. 18 to 19.

  4-15, the code | symbol 40a, 40b, 40c shows embodiment of the high pressure fluid engine of this invention, respectively. The high-pressure fluid engine 40 a uses the above-described inconstant speed rotation conversion mechanism 10 and has the first disk rotor 14 having the first piston 16 and the second disk rotor 24 having the second piston 26 installed therein. And a fluid inlet hole 44 and a fluid outlet hole for high-pressure fluid respectively provided in the cylinder 42 corresponding to the dead center positions of the first piston 16 and the second piston 26 in the cylinder 42. 46 is further provided.

  With the above configuration, when the crank pin guide means 32 serves as a flywheel and the first piston 16 and the second piston 26 are rotated, the first crank pin 20 and the second crank pin 30 are moved to the crank pin. The first piston 16 and the second piston 26 are caused to reciprocate while being slid on the guide means 32, and a slow relationship is generated in the circulation of the first piston 16 and the second piston 26, and one of the first piston 16 and the second piston 26 has a predetermined angle. The slow piston repeats the slow speed with a delay (180 degrees in the illustrated example), and the slower piston shuts off the communication with the fluid inlet hole 44 and the fluid outlet hole 46, and the faster piston communicates with the fluid inlet hole 44. The high pressure fluid is allowed to flow in, and the higher piston receives the expansion pressure of the high pressure fluid to obtain the output. At the same time, by discharging extruding the inlet to the expansion of the finished fluid, the constant speed output shaft 36 is caused to continuously rotate.

  As the high-pressure fluid, a high-pressure fluid such as a high-pressure liquid can be used in addition to a high-pressure gas body such as water vapor. In this embodiment, water vapor is used as the high-pressure fluid.

  The high-pressure fluid engine of the present invention will be described more specifically below.

  As well shown in FIGS. 4 to 15, the cylinder 42 includes an annular cylinder 48 and side plates 50 and 52. The cylinder 42 is provided with a fluid inflow hole 44 for allowing a high-pressure fluid to flow in and a fluid discharge hole 46 for discharging the fluid after the high-pressure fluid has expanded and finished work.

  In the cylinder 42, the first piston 16 and the second piston 26 are fixed to the first disk rotor 14 and the second disk rotor 24 having the labyrinth seal 54, respectively. The disk rotor 24 is configured to be able to freely rotate in opposition to each other, and is sealed with a labyrinth seal 54 (see FIG. 4B).

  A first crankpin 20 and a second crankpin 30 are respectively connected to the first disc rotor 14 and the second disc rotor 24 via the first crankshaft 12 and the second crankshaft 22. The first crankshaft 12 and the second crankshaft 22 are supported by bearings 56 and 58. In the above description, the angular phases of the first piston 16 and the second piston 26 and the first crankpin 20 and the second crankpin 30 are 180 degrees, respectively, but the angular phase of the piston and the crankpin is 0 degree to 360 degrees. If it is held at the same angle, it is established in terms of configuration.

  The first crankpin 20 and the second crankpin 30 are formed by the crankpin guide means 32 in which the rotation center O1 of the first disk rotor 14 and the second disk rotor 24 and the rotation center O2 thereof are eccentric by an eccentric amount e. The guide groove 34 is slidably guided. The crank pin guide means 32 is supported by a bearing 60, and the first crank pin 20 and the second crank pin 30 are slid in the guide groove 34 only in the longitudinal direction of the guide groove 34 via the slider 62. It is set to be movable. This crank pin guide means 32 plays the same role as a flywheel. In FIG. 4B, reference numeral 64 indicates a fluid drain hole for draining fluid leaking from the labyrinth seal 54.

  The counterweight 66 provided on the opposite side of the first crankpin 20 and the second crankpin 30 with respect to the balance weight of the first crankpin 20 and the second crankpin 30 and the first crank arm 18 and the second crank arm 28 is as follows. 68, the dynamic balance can be perfectly balanced.

  The shape and structure of the counterweights 66 and 68 are not particularly limited. For example, the counterweights have a fan-shaped structure and a thinned structure having punched holes 101 as shown in FIGS. 7 (a) and 7 (b). Can be preferably used.

  Further, a structure for hermetically sealing (gas sealing) high-pressure fluid will be described with reference to FIGS. 9 (a) and 9 (b). The second piston 26 fixed to the second disk rotor 24 with bolts 70 is provided with airtight pieces 72, 74, 76, and the current piston ring 78 (that is, the circular one currently used in the reciprocating engine) is first. Mounted on the outside of the two pistons 26. Thereafter, the piston crown 80 is attached with a bolt 82 (partially omitted). By doing so, the airtight pieces 72, 74, and 76 are projected so as to come into contact with the inner surface of the cylinder 42 due to the tension of the working piston ring 78, and the airtightness of the high-pressure fluid such as gas is produced even in the low speed state. . Reference numeral 84 denotes a punching hole for weight reduction. The same effect can be obtained by using a coil spring instead of the piston ring.

  Further, with respect to leakage on the outer circumferences of the first disk rotor 14 and the second disk rotor 24, a disk piston ring 86 is mounted, and a soft rubber-like elastic body 88 is provided on the inner side, whereby a disk piston is provided. A force is generated that the ring 86 projects to the piston side and the outside of the disk rotor, thereby enabling airtightness. Of course, the effect is the same with a metal leaf spring instead of soft rubber.

  Further, by providing a disk piston ring 90 on the mating surface of the first disk rotor 14 and the second disk rotor 24 in the same manner, it is possible to achieve airtightness. Note that the above-described airtight piece, piston ring, etc. can be used without lubrication and can be expected to be airtight if a BEAREE (registered trademark) product of NTN Corporation is used.

  Next, the mechanism governing the slow movement of the first piston 16 and the second piston 26 in the high-pressure fluid engine 40a of the present invention will be described with reference to FIGS. 10 (a) to 10 (c). FIG. 10A shows the state where the crankpin guide means 32, the guide groove 34, the fluid inlet hole 44 and the fluid outlet hole 46 of the annular cylinder 48, the first crankpin 20 and the second crankpin 30 are left from the state of FIG. It is.

  In FIG. 10A, the relationship between the first piston 16 and the second piston 26 and the first crank pin 20 and the second crank pin 30 is that the position of the guide groove 34 is at a right angle, while the first piston 16 Each of the second pistons 26 is in a state of 45.4 degrees.

  In order to make it easier to understand, the fluid inflow hole 44 and the fluid discharge hole 46 are omitted, and the crank pin guide means 32 is rotated 30 degrees to the right from the state of FIG. The second piston 26 rotates 37 degrees and the first piston 16 rotates 23 degrees in the same direction.

  In order to explain the force relationship, only the guide groove 34, the first crankpin 20 and the second crankpin 30 are shown in FIG. The reason why the first piston 16 and the second piston 26 rotate slowly and the reason why they rotate in the same direction even if the same force acts on the first piston 16 and the second piston 26 will be described below.

  The forces received by the two pistons (first piston 16 and second piston 26) that rotate in the same direction (rotation center O1) are applied to the center CF and CS of the respective crankpins of the first crankpin 20 and the second crankpin 30. The acting force is defined as force F.

  The rotational torque when the force F is viewed from the rotation center O2 of the crankpin guide means 32 is defined as a point A with a foot perpendicular to the force F from the rotation center O2 and a distance a between the A and the rotation center O2. The rotational torque is a × F. In the same way, assuming that a point perpendicular to the force F acting on the first crankpin 20 is a line B and the distance from O2 is a distance b, the rotational torque is b × F, and the total rotational torque is F × (a -B).

  From the above, it can be seen that the rotational force is generated and the slow speed relationship between the first piston 16 and the second piston 26 is generated.

  FIG. 18 shows a continuous shift state with respect to the relative angular velocities of the first crankpin 20 and the second crankpin 30. In FIG. 18, the angular velocity of the first crank pin 20 is indicated by a solid line, and the angular velocity of the second crank pin 30 is indicated by a one-dot chain line. As can be seen from FIG. 18, each of the first piston 16 and the second piston 26 repeats the slow speed once while the guide groove 34 of the crankpin guide means 32 makes one rotation. This is preferable from the viewpoint of torque fluctuation.

  Next, the operating state of the high-pressure fluid engine 40a of the present invention will be described with reference to FIGS. 11 (a) to 11 (e). In order to facilitate understanding, the first crankpin 20 and the second crankpin 30 paired with the first piston 16 and the second piston 26 are expressed in the same pattern. In FIG. 11, the first piston 16 and the first crankpin 20 and the second piston 26 and the second crankpin 30 are in a paired relationship.

  FIG. 11A shows a dead center state of a normal reciprocating engine, and when this state is rotated to the right in the direction of the arrow, a narrow space between the first piston 16 and the second piston 26 (FIG. 11A). 11b), the high pressure fluid continues to flow, the speed of the second piston 26 is higher than that of the first piston 16, and the state shown in FIG. The flow of high-pressure fluid stops due to blockage. After that, the second piston 26 adiabatically expands without being supplied with the high-pressure fluid, and enters the state of FIG. 11E through FIG.

  On the other hand, in FIG. 11 (a) and FIG. 11 (b), the wide space on the opposite side formed by the first piston 16 and the second piston 26 allows the fluid that has flowed in and finished to expand in the previous stroke to the second piston 26. Is pushed out to the fluid discharge hole 46, while the first piston 16 continues to block the fluid inflow hole 44 of the high-pressure fluid. This completes one cycle, and the crank pin guide means 32 rotates 180 degrees between FIGS. 11 (a) to 11 (e). Comparing FIG. 11 (e) and FIG. 11 (a), only the positions of the first piston 16 and the second piston 26 are interchanged, and the configuration is exactly the same. Then, the above-described operation is repeated.

  In order to control the amount of high-pressure fluid flowing into the first piston 16 and the second piston 26 (without changing the fluid pressure), the first piston 16 in FIG. Control is possible by changing the position of the opening edge 94 of the guide opening 92, and the energy held by the gas can be optimally utilized.

  That is, by moving the opening edge 94 of the guide opening 92 of the fluid inflow hole 44 toward the fluid discharge hole 46, the width of the guide opening 92 is consequently reduced, and the opening edge 94 is moved to the fluid discharge hole 46. The first piston 16 or the second piston 26 narrows the fluid inflow hole 44 as compared with the direction before moving toward the first side. As a result, the amount of high-pressure fluid that flows in decreases, but the amount of expansion increases, so the output per one stroke decreases but the thermal efficiency increases.

  11 (a) to 11 (e) are operation diagrams in the case where a fluid inflow hole 44 of a type giving priority to output per one stroke is provided, whereas FIGS. 12 (a) to 12 (f) show thermal efficiency. It is an operation | movement figure at the time of providing the fluid inflow hole 44 of the priority type. 12 (a) to 12 (f), the first piston 16 or the second piston 26 can be obtained by moving the opening edge 94 of the guide opening 92 of the fluid inflow hole 44 toward the fluid discharge hole 46. The position where the fluid is cut off becomes earlier, and the amount of high-pressure fluid flowing in decreases, but the amount of expansion increases.

  FIGS. 13A to 13D are explanatory views of a mechanism for changing the position of the opening edge 94. In FIG. 13A, reference numeral 96 denotes an inflow amount adjusting piece. In a state where the inflow amount adjusting piece 96 is inserted into the guide opening 92 of the fluid inflow hole 44, the position of the opening edge 94 is the position shown in FIG. FIG. 13B is a sectional view taken along line XII (b) -XII (b) of FIG. 13A in a state where the inflow amount adjusting piece 96 is inserted. FIG. 13C shows a state where the inflow adjustment piece 96 is slid and removed. In a state where the inflow amount adjusting piece 96 is removed, the position of the opening edge 94 is the position shown in FIG. FIG. 13B is a cross-sectional view taken along the line XII (d) -XII (d) in FIG. 13C with the inflow amount adjusting piece 96 removed.

  In this way, it can be seen that the position of the opening edge 94 shown in FIG. 13A becomes the position shown in FIG. 13C by inserting the inflow amount adjusting piece 96. Since various means for sliding the inflow rate adjusting piece 96 are conceivable, illustration is omitted. Needless to say, various mechanisms other than the slide mechanism can be employed as long as the amount of inflow of the high-pressure fluid can be adjusted.

  Further, even if the fluid is changed from the high pressure gas to the high pressure liquid, it can be used as it is as long as the liquid compression does not occur in the way of taking the position of the opening edge 94 of the fluid inflow hole 44.

  FIG. 19 is a graph showing the relationship between pressure and volume when the piston blocks the fluid inflow hole 44. As shown in FIG. 19, the ideal energy usage is an IJEMI diagram. That is, the fluid enters from the fluid inflow hole 44 between IJ, the first piston 16 blocks the fluid inflow hole 44, and then the adiabatic expansion between JE and the second piston 46 communicates with the fluid discharge hole 46 to work. Since the fluid that has finished is discharged, there is no waste. That is, the thermal efficiency is in the best state.

  On the other hand, the output that is output in this state but the worst thermal efficiency is the IREMI diagram shown in FIG. 19. Since the fluid exits the fluid discharge hole 46 with a high pressure, it is between the hatched LHEL lines. The energy that still adiabatically expands is wasted. In the middle is the IKFEMI diagram, and in reality it is necessary to experiment variously to determine the best blocking position. Nevertheless, the shaded portion of FGEF is lost, and the position of the so-called opening edge 94 occupies an important factor.

  The above-described configuration is a configuration in which the guide groove 34 is provided in the crankpin guide means 32. However, in addition to the above-described configuration, for example, the configurations shown in FIGS. 14A to 14C may be adopted. it can.

  In FIG. 14, the code | symbol 40b shows another aspect of the high pressure fluid engine of this invention. In the crankpin guide means 98, a polished polishing shaft 100 is screwed to the center thereof, and sliders 102 and 104 are loosely inserted in the polishing shaft 100 so as to be slidable in the longitudinal direction. The first crankpin 20 and the second crankpin 30 are respectively fitted to the sliders 102 and 104 so as to be freely rotatable as shown in FIG.

  If comprised in this way, the function similar to the structure which provided the guide groove 34 in the crankpin guide means 32 can be acquired, and it will become possible to control the slow speed of the 1st piston 16 and the 2nd piston 26. FIG.

  The above description is based on the configuration in which the crankpin guide means is integrated in one direction. Next, a configuration in which the crankpin guide means is provided in opposite directions, that is, on both sides, will be described with reference to FIG.

  15A to 15C, reference numeral 40c represents another embodiment of the high-pressure fluid engine of the present invention. In the high-pressure fluid engine 40c of the present invention, the first crankpin 20 and the second crankpin 30 are provided on both sides, respectively, and corresponding to the respective crankpins, FIGS. 15 (a) and 15 (c). As shown well, crank pin guide means 32 having a guide groove 34 is provided. Reference numerals 106 and 108 denote the same number of gears.

  As shown in FIGS. 15A to 15C, the crankpin guide means 32 is connected to the gear 108, and the constant speed rotation output shaft 36 is connected to the gear 106. Then, when the first disk rotor 14 to which the first piston 16 is attached rotates, the first crankshaft 12 rotates, and the first crankpin 20 reciprocates while sliding in the guide groove 34, whereby the crankpin When the guide means 32 rotates, the gear 108 rotates, and the constant speed rotation output shaft 36 is rotated by the gear 106 meshing with the gear 108.

  Similarly, when the second disk rotor 24 to which the second piston 26 is attached rotates, the second crankshaft 22 rotates, and the second crankpin 30 reciprocates while sliding in the guide groove 34, thereby When the crankpin guide means 32 rotates, the gear 108 rotates, and the constant speed rotation output shaft 36 is rotated by the gear 106 meshing with the gear 108. At this time, as shown in the figure, the first piston 16 and the second piston 26 are rotated by the positions of the first crankpin 20 and the second crankpin 30 guided by the crankpin guide means 32 existing on the left and right. As a result, the first crank pin 20 and the second crank pin 30 are reciprocated while sliding on the left and right crank pin guide means 32, and the rotation of the first piston 16 and the second piston 26 has a slow relationship. One of the piston and the second piston repeats the slow speed with a delay of a predetermined angle (180 degrees in the illustrated example), and the constant speed rotation output shaft 36 is continuously rotated. The first crankpin 20 and the second crankpin 30 are provided with counterweights 66 and 68, respectively.

  Furthermore, the closed cycle engine of the present invention using the above-described high-pressure fluid engine will be described with reference to FIGS.

  In FIG. 16, reference numeral 110a denotes a closed cycle engine of the present invention. The closed cycle engine 110a includes the above-described high-pressure fluid engine 40a, an exhaust gas discharge engine 112 that discharges exhaust gas, a heat exchanger 114 that generates water vapor from the exhaust gas of the exhaust gas discharge engine 112, and the high-pressure fluid engine. A condenser 116 that cools the discharged fluid from 40a and returns it to water, and connects the heat exchanger 114 to the fluid inlet hole 44 of the high-pressure fluid engine 40a, and the condenser. A closed cycle structure is formed by connecting 116 to the fluid discharge hole 46 of the high-pressure fluid engine 40a.

  A conventional reciprocating engine can be used as the exhaust gas discharge engine 112, and a known heat exchanger can be used as the condenser as the condenser 116.

  Further, a temperature sensor 122 is attached to the cylinder head 120 and the cylinder 118 of the cylinder 118 of the exhaust gas emission engine 112, and water is injected from the nozzle 126 into the high temperature portion 124 of the cylinder 118 according to the information of the temperature sensor 122 to cool. At the same time, it is configured to operate using steam generated.

  The closed cycle engine 110a sets a temperature sensor 122 in a high temperature portion 124 of a component (cylinder head 120 and cylinder 118 in the figure) of an exhaust gas emission engine 112 (reciprocating engine in the figure), and sends the information to the controller 128 to supply water. As shown in FIG. 16, the pressure water from the pump 130 is sprayed from the nozzle 126 to the high temperature portion 124 of the cylinder 118 shown at the representative location, thereby causing the water to be generated. In addition, the high temperature part 124 means the location which becomes high temperature during driving | operation, and the representative location was shown in the example of illustration.

  The temperature sensor 122 is provided in the cylinder 118 so as not to impair the function of the oil, that is, to avoid the burning of the piston, and the temperature sensor 122 is provided in the cylinder head 120 for troubles in the valve stem and the valve seat. This is to avoid it.

In setting the temperature of the temperature sensor 122, it is considered that the temperature of the cylinder is infinitely low, that is, the saturated steam temperature is lowered. Generally, in order to reduce heat loss from the cylinder or cylinder head, A wall temperature in the range of around 180 degrees is appropriate. Accordingly, the saturation pressure is 10.2 kg / cm ² . Furthermore, if high pressure is aimed, the wall temperature of the cylinder is allowed to rise from 260 to 270 degrees, so it can be further increased.

  Exhaust gas that exits from the exhaust port 138 of the exhaust gas emission engine 112 enters the exhaust gas intake port 140 of the heat exchanger 114, heats the water vapor, and then is discharged into the air from the exhaust gas outlet 142 of the heat exchanger 114.

  On the other hand, the water vapor accumulated in the jacket 132 enters the water vapor inlet 136 of the heat exchanger 114 from the vapor outlet 134 at the top of the cylinder 118. This heated superheated steam exits from the steam outlet 144 of the heat exchanger 114 by connecting the fluid inlet hole 44 and the steam outlet 144 of the high-pressure fluid engine 40a, and rotates a generator or the like (not shown). Energy conversion is performed by the high-pressure fluid engine 40a. Reference numeral 156 denotes a drain outlet of the heat exchanger 114.

  The water vapor that has finished its work in the high-pressure fluid engine 40a enters the low-pressure steam intake port 146 of the condenser 116 through the fluid discharge hole 46, is cooled by the cooling air of the cooling fan 148, condenses the water vapor, and the water droplet outlet After exiting 150, it enters the reservoir 152 and is reused. Reference numeral 154 denotes a distributor for distributing water.

  Next, another embodiment of the closed cycle engine of the present invention is shown in FIG. In FIG. 17, the code | symbol 110b shows another aspect of the closed cycle engine of this invention. Exhaust gas from the exhaust gas emission engine 112 enters the exhaust gas intake port 140 of the heat exchanger 114, heats the water vapor, and is then released into the air from the exhaust gas outlet 142 of the heat exchanger 114.

  On the other hand, the water that has entered the steam inlet 136 of the heat exchanger 114 from the water pump is heated to become steam, and this steam is connected to the fluid inlet hole 44 and the steam outlet 144 of the high-pressure fluid engine 40 a, whereby the heat exchanger 114. The steam is discharged from the steam outlet 144, and a generator (not shown) is rotated to convert the energy by the high-pressure fluid engine 40a. Then, the fluid discharge hole 46 is connected to the condenser 116 to make a closed cycle, and the energy can be effectively used.

It is a perspective view which shows one embodiment of this inconstant speed rotation conversion mechanism. FIG. 2 is an exploded view of FIG. 1. It is a schematic explanatory drawing which shows the amount of rotation changes of the crankpin in the inconstant speed rotation conversion mechanism of FIG. 1 shows an embodiment of a high-pressure fluid engine according to the present invention, in which (a) is a cross-sectional view, and (b) is an enlarged schematic view of IV (b) portion of (a). It is the VV sectional view taken on the line of Fig.4 (a). 1 shows an embodiment of a high-pressure fluid engine according to the present invention, in which (a) is a view taken in the direction of the arrow Y in FIG. 5, and (b) is a cross-sectional view taken along line VI (b) -VI (b) in (a). 1 shows an embodiment of a high-pressure fluid engine, (a) is a magnified view of one embodiment of a counterweight used in the high-pressure fluid engine of the present invention, and (b) is a VII (b) of (a). It is -VII (b) sectional view taken on the line. 1 shows an embodiment of the high-pressure fluid engine of the present invention, where (a) is a cross-sectional view, (b) is a cross-sectional view taken along line VIII (b) -VIII (b) of (a), and (c) is (a) FIG. 1 shows an embodiment of a seal of a high-pressure fluid engine according to the present invention, in which (a) is a sectional view and (b) is a sectional view taken along line IX (b) -IX (b) of (a). It is explanatory drawing of the mechanism which governs the slow motion of the piston in the high pressure fluid engine of this invention. It is an operation state figure of a piston in a high pressure fluid engine of the present invention. FIG. 5 is an operation diagram in the case where a fluid inflow hole of a type giving priority to thermal efficiency is provided in the high pressure fluid engine of the present invention. It is explanatory drawing of the mechanism of a position change of an opening part edge in the high pressure fluid engine of this invention. 4 shows another embodiment of the high-pressure fluid engine of the present invention, in which (a) is a cross-sectional view, (b) is a cross-sectional view of (a), and (c) is a partially cutaway schematic perspective view of (a). The other embodiment of the high-pressure fluid engine of this invention is shown, (a) is a side view of (b), (b) is sectional drawing, (c) is a side view of (b). It is the schematic which shows one embodiment of the closed cycle engine of this invention. It is the schematic which shows another embodiment of the closed cycle engine of this invention. It is a graph which shows the relative angular velocity of the 1st crankpin and the 2nd crankpin in the high pressure fluid engine of this invention. It is a graph which shows the relationship between the pressure and the volume in the high pressure fluid engine of this invention.

10: This inconstant speed rotation conversion mechanism, 12: 1st crankshaft, 14: 1st disc rotor, 16: 1st piston, 18: 1st crank arm, 20: 1st crankpin, 22: 2nd crankshaft 24: second disk rotor, 26: second piston, 28: second crank arm, 30: second crank pin, 32: crank pin guide means, 34: guide groove, 36: constant speed rotation output shaft, 40a, 40b, 40c: high pressure fluid engine, 42: cylinder, 44: fluid inflow hole, 46: fluid discharge hole, 48: annular cylinder, 50, 52: side plate, 54: labyrinth seal, 56, 58, 60: bearing 62: Slider, 64: fluid vent hole, 66, 68: counterweight, 70: bolt, 72, 74, 76: airtight piece, 78: working piss Ring: 80: piston crown, 82: bolt, 84: punching hole, 86, 90: piston ring for disc, 88: elastic body, 92: guide opening, 94: edge of opening, 96: inflow adjustment piece, 98 : Crank pin guide means, 100: polished shaft, 101: punched hole, 102, 104: slider, 106, 108: gear, 110a, 110b: closed cycle engine, 112: exhaust gas emission engine, 114: heat exchanger, 116: condenser, 118: cylinder, 120: cylinder head, 122: temperature sensor, 124: high temperature section, 126: nozzle, 128: controller, 130: water pump, 132: jacket, 134: steam outlet, 136: steam Inlet, 138: Exhaust port, 140: Exhaust gas inlet, 142: Exhaust gas out 144: Steam outlet, 146: Low pressure steam inlet, 148: Cooling fan, 150: Water droplet outlet, 152: Reservoir, 154: Distributor, 156: Drain outlet, a, b: Distance, e: Eccentricity A, B: point, CF, CS: center of rotation, F: force, O1: center of rotation, O2: center of rotation.

Claims (3)

Attached to the first crankshaft and the second crankshaft, respectively, to the crankpin guide means passing through the rotation center of the constant speed rotation output shaft eccentrically set with respect to the first crankshaft and the second crankshaft having the same rotation center a first crank pin and a second crank pin that is a high-pressure fluid engine with non-uniform rotary converting mechanism respectively slidably formed by guide,
A cylinder for installing therein a first disk rotor having a first piston and a second disk rotor having a second piston;
A fluid inlet hole and a fluid outlet hole respectively provided at the location of the cylinder corresponding to the dead center position of the first piston and the second piston in the cylinder;
Further provided,
When the first piston and the second piston are rotated by the crank pin guide means serving as a flywheel, the first crank pin and the second crank pin are reciprocated while sliding on the crank pin guide means. As a result, the first piston and the second piston have a slow relationship, and the slower of the first piston and the second piston shuts off the communication with the fluid inlet hole and the fluid outlet hole. The faster piston, which is in communication with the fluid inflow hole, allows high-pressure fluid to flow in. The higher piston receives the expansion pressure of the high-pressure fluid to obtain an output, and at the same time, it flows in and expands. The constant speed rotation output shaft is continuously rotated by extruding and discharging the finished fluid. High pressure fluid engine. A closed cycle engine using the high-pressure fluid engine according to claim 1 ,
An exhaust emission engine that emits exhaust gas;
A heat exchanger that generates water vapor with the exhaust gas of the exhaust gas emission engine;
A condenser that cools and returns the exhaust fluid from the high-pressure fluid engine to water;
Further provided,
A closed cycle structure in which the heat exchanger is connected to the fluid inlet hole of the high-pressure fluid engine and the condenser is connected to the fluid outlet hole of the high-pressure fluid engine. engine. A temperature sensor is attached to a cylinder head and a cylinder of the cylinder of the exhaust gas emission engine, and water is jetted into the high temperature part of the cylinder according to the information of the temperature sensor to cool and operate using steam generated at the same time. The closed cycle engine according to claim 2 .

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