EP0087746A1 - Rotary piston supercharger driven by exhaust gases - Google Patents

Abstract

An internal axis crankless rotary piston engine with an internal rotor and an external rotor mounted eccentrically to one another, with a circularly curved path of the external rotor mounting surrounding an internal rotor shaft, wherein the path of an external rotor bearing is mounted by a plurality of bearings which are spaced from one another in the direction of movement and which are fixed relative to the geometrical axis of the internal rotor shaft.

Description

The invention relates to an exhaust gas-operated rotary piston loader for the mechanically independent arrangement on the exhaust pipe of an internal combustion engine with two rotary piston machines in drive connection with one another, one of which forms the loading unit and the other the exhaust unit of the rotary piston loader.

From DE-PS 2 232 592 by the same inventor, an exhaust gas-operated rotary piston charger of this type is known, the charging unit and exhaust unit of which are each formed by an external-axis rotary piston machine with a stationary working chamber wall. Together with the widely used so-called turbochargers, this known rotary piston supercharger has the advantage over mechanically driven superchargers that the internal combustion engine is charged as a function of the exhaust gas flow, i.e. in line with the actual air requirement. In the case of mechanically driven superchargers (compressors), on the other hand, an adjustable gearbox is required to adapt to the actual air requirement, since the air requirement is not directly dependent on the speed of the internal combustion engine. In addition, a mechanically operated supercharger removes useful power from the internal combustion engine. The utilization of the energy of the exhaust gas flow, on the other hand, can contribute to reducing the noise, so that the technical outlay for the "exhaust" system of the internal combustion engine is reduced.

The current state of technical development comes for example in the technical article "Turbo or mechanical charger?" in the magazine "Automobil-Revue" No. 6 v. February 11, 1982 (from "Automotive Industries" by John McElroy). The "Roots" compressor mentioned in this article also represents an external-axis rotary lobe machine that would be suitable for very high speeds due to its fixed rotary axes and could therefore be built in a small size and with low inertia. In fact will however, these Roots compressors were driven at a comparatively low speed because, due to strong squeezing currents at a higher speed, there is an extremely high level of noise and a strongly increasing power requirement. The squeezing currents result from the surfaces moving towards each other at high speed when the two runners roll over each other.

The disadvantage of turbochargers is also well known. It is caused by the non-proportional power curve, which is characteristic of turbines, so that charging in the lower speed ranges and load conditions of the internal combustion engine is inadequate and is only optimal when the engine is at full power. In order to reduce this misunderstanding, it has started to reduce the diameter of the exhaust gas turbine in such a way that it already has the high speed required for the supercharging when the engine is a quarter or a half power.

As a result, however, at full load, part of the exhaust gas must be bypassed the turbine unused so that the charge does not become too high.

The above-mentioned DE-PS 2 232 592 was based on the task of creating an exhaust gas-operated charging device which ensures optimal charging both in the lower speed ranges and load conditions and in the upper ones. Optimal supercharging in the upper speed ranges could not be achieved with the previously known exhaust-powered rotary piston loaders because of their excessive internal power requirement due to squeezing flow losses. These squeezing flows arise when the piston of one rotor engages in the gap of the counter rotor. As the already known rotary piston superchargers were consequently not superior to the turbocharger in the upper speed range and less importance was attached to the performance improvement in the lower three-speed range, the experts have so far paid no attention to them.

In addition to the disadvantages caused by squeezing currents the previous exhaust-gas-operated rotary piston loader has the further disadvantage that its manufacture is relatively complex owing to the shape of its rotor and the required precise mutual geometrical assignment, particularly taking thermal loads into account.

The object of the present invention is to find an exhaust gas-operated rotary piston loader suitable for high speeds, which reduces energy losses and noise caused by squeezing currents to an insignificant extent, has no or only negligible harmful spaces and thus enables optimal charging in all speed ranges of an internal combustion engine. In addition, it should cope with the high thermal and mechanical loads with the least possible design effort, so that it can be inexpensively manufactured as a permanent mass product. To solve this problem, an exhaust gas-operated rotary piston charger of the type mentioned at the outset is proposed, which is characterized according to the invention in that at least one of the rotary piston machines is an internal-axis rotary or rotary piston machine.

For a better understanding of the terminology used, reference is made to the specialist book "Classification of the Rotary Piston Machines" by F. Wankel, Deutsche Verlagsanstalt, Fachverlag, Stuttgart, 1963.

Compared to an external-axis rotary piston machine, as used in a loader according to the aforementioned DE-PS 2 232 592, there is the essential advantage that the movement of surfaces towards one another during an engagement movement does not take place with the peripheral speed of one or more rotors, but with one relative sinusoidal movement, which slows down to 0, so that the squeezing currents, which have previously led to great difficulties, are practically avoided.

Although internal-axis rotary or rotary piston machines in numerous versions in addition to the numerous rotary piston machines, of which the essential examples in the above-mentioned textbook "Classification of the Rotary Piston Machines" have been known for a long time, this advantageous property for the application for an exhaust-powered rotary piston loader was not obvious to the person skilled in the art. Examples of internal-axis rotary lobe or rotary lobe machines with meshing engagement are shown in DE-PS 103 82 (1879), GB-PS 961 872, FIG. 3 of GB-PS 1 046 504, DE-OS 24 38 189 with the corresponding US PS 3 876 348, 3 800 941, 3 954 355 and others. From a formal point of view, it should be noted that although the mechanical coupling of an internal-axis rotary piston machine to an internal-axis internal combustion engine is known per se for pre-compression and post-expansion in the diesel process, this combination is not comparable with a mechanically independently operated rotary piston loader, the mode of operation of which is not dependent on the cycle with the internal combustion engine, but is independent of the engine speed. A combination of this known type, the rotary pistons of which have slip engagement, is shown in US Pat. No. 3,405,692.

If internal-axis rotary piston machines are not designed as rotary piston machines, but as rotary piston machines, i.e. with a rotor mounted on the rotating or eccentric journal of a crankshaft, special control devices must be provided on the outer housing if large control cross sections are to be achieved. According to one embodiment of the invention, the internal-axis machines are designed as rotary piston machines, i.e. the rotors rotate around fixed axes of rotation, so that the external rotor can itself control large inlet and outlet channels of the fixed housing. Apart from this, the rotary lobe machines are more suitable than rotary lobe machines due to their balancing ability and the lack of centrifugal bearings.

Further advantageous embodiments of the invention can be found in the following description of exemplary embodiments with reference to the drawings.

Since an exhaust gas-operated rotary piston loader according to the invention thus works, in contrast to turbomachines, according to the displacement principle and, in contrast to the prior art according to DE-PS 2 232 592, the displacement movement does not lead to any significant squeezing currents and, moreover, harmful spaces are avoided, the engine results in every speed range optimal charging. Charging is optimal because the amount of air supplied by the charger of the supercharger to the engine is directly related to the amount of exhaust gas flow that drives the exhaust unit of the charger, i.e. the exhaust gas flow directly controls the charger in an optimal way.

• Fig. 1 einen radialen Querschnitt durch die Ladeeinheit des Laders,
• Fig. 2 einen Axialschnitt durch die Ladeeinheit nach Fig. 1,
• Fig. 3 eine teilweise geschnittene Aufsicht auf den abgasbetriebenen Lader,
• Fig. 4a - 4h schematische Darstellungen mehrerer aufeinanderfolgender Bewegungspositionen der Ladeeinheit nach Fig. 1,
• Fig. 5 und 6 schematische Darstellungen von zwei Ausführungsbeispielen einer Antriebsverbindung zwischen Innen- und Aussenläufer,
• Fig. 7 einen axialen Teilschnitt zu der Antriebsverbindung nach Fig. 6,
• Fig. 8 einen Querschnitt entlang der Linie VIII-VIII der Fig. 9 der Lageranordnung für den Aussenläufer,
• Fig. 9 einen axialen Teilquerschnitt durch eine Rotationskolbeneinheit des Laders entlang der Linie IX-IX der Fig. 8,
• Fig. 10 einen axialen Teilquerschnitt durch eine Abgaseinheit des Laders mit Mitteln zur Kühlung der Läufer,
• Fig. 11 einen Querschnitt durch den lnnenläufer entsprechend einem weiteren Ausführungsbeispiel für seine Kühlung, und
• Fig. 12a - 12d schematische Darstellungen von aufeinanderfolgenden Bewegungspositionen einer Abgaseinheit.

In the drawings:

• 1 is a radial cross section through the loading unit of the loader,
• 2 shows an axial section through the loading unit according to FIG. 1,
• 3 is a partially sectioned top view of the exhaust gas powered loader,
• 4a-4h are schematic representations of several successive movement positions of the loading unit according to FIG. 1,
• 5 and 6 are schematic representations of two embodiments of a drive connection between the inner and outer rotor,
• 7 is a partial axial section of the drive connection of FIG. 6,
• 8 shows a cross section along the line VIII-VIII of FIG. 9 of the bearing arrangement for the external rotor,
• 9 is an axial partial cross section through a rotary piston unit of the charger along the line IX-IX of FIG. 8,
• 10 is an axial partial cross section through an exhaust unit of the charger with means for cooling the rotor,
• 11 shows a cross section through the inner rotor in accordance with a further exemplary embodiment for its cooling, and
• 12a-12d are schematic representations of successive movement positions of an exhaust gas unit.

1 and 2 show, in two sectional views, the charging unit 2 of the exhaust gas-operated charger shown overall in FIG. 3. The exhaust gas unit 3 is of essentially the same design as the loading unit, and its inner runners 4 are rigidly connected via a common rectilinear main shaft 6 of the charger. However, it is understood that the inner rotors 4, 4 'of the two machine units, which are eccentrically mounted on the shaft 6, are angularly offset from one another, e.g. are arranged at 90 °, so that the starting of the loader is facilitated or even with a dead center position of one of the units is possible without an external drive. The start-up takes place due to a vacuum in the pressure line 7 of the charging unit caused by the internal combustion engine and an excess pressure in the inflow line of the exhaust gas unit of the charger, not shown.

The inner rotor 4 has a circular cross section and rotates in an eccentric movement about the shaft 6, which only rotates about its own axis. In the example shown, the shaft 6 is rigid with the side housing covers 11, 12 by two needle or roller bearings 9, 10 connected sealing and bearing bodies 13, 14 mounted. By rotating the inner rotor 4 about the locally fixed shaft 6, it can be completely balanced in itself, so that no centrifugal forces act on the bearings and very high speeds are possible.

One each on the outer circumference of the sealing and bearing bodies 13, 14 provided second bearing 15, 16 for example, execution of like _r ung intended for storage of the external rotor 18 of this rotary piston machine about its thus locally fixed center axis. A seal 20, 21 on the part 19 of the sealing and bearing bodies 13, 14 facing the working space 22 of the loading unit 2 and projecting into the side walls of the outer rotor 18 prevents the bearings 9, 10, 15, 16 and a drive transmission 23 between the Internal and external rotor disadvantageously come into contact via the shaft 6 with the medium flowing through the unit 2 or also 3. The sealing and bearing bodies 13, 14 also allow a much larger diameter of the shaft 6 of the inner rotor, which is of particular advantage for a large axial length of the rotary piston unit and in the case of a hollow version of the shaft. In addition, they allow a greater distance between the bearing axes of both rotors, which leads to a greater throughput of the rotary piston machine.

The outer runner 18 enclosing the inner runner 4 consists of two crescent-shaped peripheral parts 24, 25 lying opposite one another and two side parts 26, 27 closing them between them. The connection to one another is made by pins 28 and screw bolts 29. The inner surfaces 30, 32 of the outer runner directed towards one another run planar and at least parallel to each other in the direction parallel to the axis of rotation, so that the inner rotor can execute a reciprocating movement in the working space 22 delimited by these surfaces 30, 32. This relatively linear movement of the inner rotor 4 relative to the outer rotor 18 despite the rotational movement of both rotors results from the kinematics of gimbals. The inner rotor 4, which can also be referred to as a rotary piston, comes into meshing engagement with two gaps in the outer rotor so that it rotates at twice the speed of the outer rotor. The transmission ratio corresponds to 1: 2, so that the pitch circle of the pinion 34 of the drive transmission 23 between the inner and outer rotor is half the diameter of the pitch circle of the hollow gear 35 points. 5 to 7 show, however, that the transmission ratio of 1: 2 can also be achieved in other ways, as will be explained in more detail below.

FIG. 4 shows the relative positions that result one after the other during the rotation of both rotors, from which the mode of operation of the charging and exhaust gas unit 2, 3 of the charger becomes clear. Due to the back and forth movement of the inner rotor 4 in the working space 22 enclosed by the outer rotor, air is sucked in via the intake duct 8 in the direction of the arrow shown in FIG. 4a in the loading unit and expelled via the pressure line 7 in the direction of the arrow shown. 4a shows a rotational position in which the center of the inner rotor coincides with the center of the outer rotor and thus also the center of the surrounding housing wall. The rotors rotate in the direction of the arrows shown in FIG. 4a, the inner rotor 4 rotating around the center A and the outer rotor around the center C, both of which are stationary. Due to the rotation of the outer rotor, its two opposing openings 38, 40 move past the openings of channels 7, 8 in the surrounding housing at half the rotational speed of the inner rotor, so that the outer rotor fulfills the task of a controlled valve. This has the advantage that the inner rotor moves during its movement, e.g. starting from the position according to FIG. 4e, does not have to move against the pressure in the pressure line 7 leading to the internal combustion engine and the connection to this pressure line does not move in the position until after the movement positions corresponding to FIGS. 4f, g, h, a have been passed through Fig. 4b shows in which the pressure in the pressure line 7, for example has just been exceeded by a small amount. This results in a significant advantage over an external-axis rotary piston machine e.g. after roots.

According to the embodiment of FIGS. 1 and 2, a pinion 34 is attached to the end of the shaft 6 of the loader, which pinion is intended to drive the external rotor 18 at half the rotational speed. To at the same Ab stood the axis of the shaft 6 from the axis of the outer rotor and with the same transmission ratio a larger diameter of the shaft 6 and a correspondingly larger diameter of the pinion 34 can be realized, so that the shaft for cooling the inner rotor 4 as shown in Fig. 11th can be made hollow and / or for a greater axial length of the rotor a higher strength, in an advantageous embodiment of the invention according to the schematic representations of FIGS. 5 and 6 between the pinion 34 ', 34 "and the outer ring gear 35', 35 "arranged at least two intermediate gears. In the example according to FIG. 5, the drive transmission between the pinion and the ring gear takes place via two intermediate gears 42, 43 of different sizes, while in the example according to FIG. 6 between the pinion 34 "and the ring gear 35" a ring gear provided with internal and external teeth 45 is arranged. This hollow gear 45 is mounted via a bearing 46 relative to the housing of the loader or a part 47 connected to the housing, as the axial partial cross section of a further embodiment of the loader according to FIG. 7 shows.

The illustrations in FIGS. 8 and 9 show an advantageous embodiment of the mounting of the outer rotor 18 on a sealing and bearing body 50 in order to reduce the speeds of the bearings. Instead of numerous small bearing bodies, three rollers 52, 53, 54 are used, two of which are supported by a journal 56 in the sealing and bearing body 50, while the third is supported on the shaft 6. These rollers 52-54, which are made of hardened material, roll on a bearing ring 58 made of hardened steel, which is inserted into a lateral hub part 60 of the outer rotor. A lateral end face of the hollow gear 62 secures the axial position of this bearing ring 58. The direct arrangement of the bearing ring 58 and the bearing rollers 52-54 rolling on it next to the gear transmission 23 'ensures good lubrication by the lubricant supplied to the transmission.

A sealing ring 66 is arranged between the outer circumference of the hub part 60 and the lateral housing part 65, which ensures the sealing of the part of the loading unit provided with lubricant.

10 and 11 show advantageous design options, in particular for the exhaust gas unit 3 of the charger which is exposed to high temperature loads. 10, in the crescent-shaped circumferential parts 24, 25 of the outer rotor 18 ', longitudinal air channels 70 are provided, the number, cross-sectional size and cross-sectional shape of which are not visible in the illustration according to FIG. 10, taking into account the data from the illustration in FIG 1 obvious possibilities for the arrangement and the strength requirements can be designed differently. When the air duct is machined in, a circular cross section is of course easier to produce. These air ducts 70 continue in the side parts 26 ', 27' and form short deflection bends 72, 73 there for connection to inlet openings 75 and outlet openings 76 in the housing side part 12 'and housing circumferential part 78, which are radially offset with respect to the duct 70, or for the arrangement of fan blades indicated by dash-dot lines. These inlet and outlet openings 75, 76 can be arranged in a grid-like manner in large numbers with webs between them or form larger, circumferential slots, as the inlet slots 80 of the exemplary embodiment according to FIG. 3 show. The rotational movement of the outer rotor 18 thus causes an air flow through the channel 70 in a fan-like manner, so that there is effective cooling of the outer rotor and adjacent parts.

FIG. 11 shows an embodiment of the inner rotor 4 ', the hollow interior 82 of which is also connected to hollow shaft parts 6' for reasons of balancing. Blower-like blades 83, which in the example shown are only arranged on the inflow side of the air flowing in the direction of the arrow 84, bring about sufficient Flow through the rotation of this inner rotor. The curvature of these blades 83 is indicated schematically by dash-dot lines, as is the case in the deflection arches 72, 73 according to FIG. 10.

The housing peripheral parts 86, 87 of the charging unit 2 and the exhaust gas unit 3 are provided with cooling fins 88, as can best be seen from the overall illustration in FIG. 3. In the exemplary embodiment according to FIG. 3, the charging unit is shown on the left-hand side and shows the intake duct 8 for fresh air and the outlet connection 7 leading to the internal combustion engine. Because of the arrangement of the exhaust gas unit 3 offset by 90 °, the inflow connection and the outflow connection for the exhaust gas are shown in FIG. 3 the internal combustion engine not visible. The exhaust gas unit effects a permanently effective closure of the exhaust gas line of the internal combustion engine, since there is no direct connection between the inlet and outlet channels 8, 7 in any of the rotary positions of the rotors 4, 18 shown in FIGS. 4a to 4h. The exhaust gas can therefore only flow into the open after it has released a substantial part of its energy to the exhaust-gas-operated rotary piston charger according to the invention. This also results in a substantial reduction in the sound waves transmitted via the exhaust gas flow and a correspondingly low outlay for the exhaust gas pipe or exhaust pipe connected to the exhaust gas unit 3 on the outflow side.

In addition to the exemplary embodiments shown, numerous further advantageous variants are possible. For example, the gearwheel-gear connection between the inner and outer rotor can be avoided by arranging a plurality of identically shaped rotary piston units next to one another on the same main shaft 6. The drive transmission then takes place directly from one rotor to the other, and corresponding rotors from axially adjacent rotary piston units are then connected to one another rigidly and at an angle. If there is nevertheless a gear connection between the inner and outer rotors, one is sufficient for several rotary piston units arranged next to one another, i.e. Charger and / or exhaust units.

Through the use of intermediate gears e.g. 5 and 6, a stronger main shaft can be used, which allows a greater axial length of a rotary piston unit.

A reduction in flow losses results from the widest possible radially directed openings 38, 40 in the outer rotor 18, which are preferably wider than the opening width of the housing openings 7, 8. For example, the mutually parallel surfaces 30, 32 of the outer rotor can extend outwards. so that the openings 38, 40 do not form a constriction.

Finally, a small part of the fresh air conveyed by the loading unit can be supplied via axially directed connection channels (not shown) for cooling the exhaust gas unit which is subjected to higher thermal loads.

12 shows for various rotational positions a to d of the rotor 4 ', 18' of an exhaust gas unit the relative positions of the inflow and outflow channel 8, 7 and the direction of the exhaust gas inflow. The rotational positions shown show that the pressure of the exhaust gas flowing in in the direction of the arrows 90 acts, especially when the cross section of the inflow channel 8 is fully open, in a direction which acts on the inner rotor due to the largest lever arms in these rotational positions, based on the rotational axis A ', leads to the greatest torque. Both channels 8, 7 run at an obtuse angle to one another, so that the flow through the exhaust gas unit takes place without a significant change in the flow direction.

Claims (11)

1.Exhaust-powered rotary piston loader for the mechanically independent arrangement on the exhaust pipe of an internal combustion engine, with two rotary piston machines in driving connection with one another, one of which forms the charging unit (2) and the other the exhaust unit (3) of the rotary piston charger, characterized in that at least one of the Rotary piston machines is an internal-axis rotary or rotary piston machine. 2. Rotary piston loader according to claim 1, characterized in that the inner-axis rotary or rotary piston machine has an outer rotor (18) mounted in fixed bearings with at least two openings (38, 40) arranged on its circumference, so that these are located on the inlet and outlet channels ( 8, 7) of a stationary housing of the rotary or rotary piston machine and move these channels (8, 7) off. 3. Rotary piston loader according to claim 2, characterized in that the rotors (4, 18) of the rotary piston machines rotate in fixed bearings. 4. Rotary piston loader according to claim 2 or 3, characterized in that the rotational speeds of the inner rotor (4) and the outer rotor (18) have a ratio of 2: 1. 5. Rotary piston loader according to one of claims 2 to 4, characterized by a gear-gear connection between the outer and inner rotor (4, 18), wherein between a on the shaft (6) of the inner rotor fixed pinion (34 ', 34 ") and at least one intermediate gear (42, 43, 45) with an external toothing (FIG. 5) or an external and internal toothing (FIG. 6) is arranged in a hollow gearwheel (35 ', 35 ") fixedly connected to the external rotor (18). 6. Rotary piston loader according to one of claims 1 to 5, characterized in that the shaft (6 ') of the inner rotor (4') is hollow and forms a flow channel together with a cavity (82) of the inner rotor. 7. Rotary piston loader according to one of claims 2 to 6, characterized in that the outer rotor is provided with at least one axially parallel cooling air duct (70) which has at least one deflection bend (72, 73) with openings (75, 76) in the housing ( 12 ', 78) of the rotary piston machine is connected. 8. Rotary piston loader according to claim 6 or 7, characterized in that in a radially directed part of the at least one in the inner rotor (4 ') or the outer rotor (18') provided flow channel (82; 72, 70, 73) for air cooling fan blades are arranged. 9. Rotary piston loader according to one of claims 2 to 8, characterized in that lateral housing covers (11, 12) carry an insert body (13, 14, 50) which projects into the adjacent side part (26, 27) of the external rotor (18), so that it seals the necessary opening in the side parts (20, 27) of the outer rotor (18). 10. Rotary piston loader according to one of claims 2 to 8, characterized in that the bearings of the outer rotor (18) are formed by the shaft of the inner rotor and two rollers (52, 53) which are on a bearing ring (58) of the outer rotor (18 ) pass on. 11. Rotary piston loader according to claim 10, characterized in that the inner diameter of the bearing ring (58) is equal to the pitch circle diameter of a hollow gear of a drive connection (23) between the inner and outer rotor (4, 18) and a roller (54) of a bearing of the outer rotor mounted on the shaft (6) of the inner rotor, the outer diameter of this roller (54) being the same as the part circular diameter of the pinion (34) fastened on the shaft (6).

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