EP0644981A1

EP0644981A1 - Piston machine.

Info

Publication number: EP0644981A1
Application number: EP93912894A
Authority: EP
Inventors: Manfred Max Rapp
Original assignee: Individual
Current assignee: Individual
Priority date: 1992-06-09
Filing date: 1993-06-04
Publication date: 1995-03-29
Anticipated expiration: 2013-06-04
Also published as: US5681156A; ES2109503T3; DE59307216D1; EP0644981B1; ATE157426T1; DE4218847A1; AU4323193A; WO1993025801A1

Abstract

A piston machine which can be used as an expansion engine, suction pump or compressor for compressible media has a piston (6) borne on crankpins (5) of at least two identical, synchronous-angle-coupled crankshafts (2) fitted so as to rotate and perpendicular to the parallel walls (9). The chamber surfaces are fully shaped as running surfaces (11, 14, 18) at the ends of which are fitted sealing components (12, 13, 19), and the high-pressure gas channel (17, 20) opens near the sealing component fitted at the end of the running surface of the peripheral wall.

Description

Piston machine

The invention relates to a piston machine of the type mentioned in the preamble of claim 1.

Such a piston machine is known from US-PS 18 64 699. In this construction, the chamber surfaces delimiting the working chambers on the circumferential wall and on the piston are each composed of cylinder-section-shaped treads which are in engagement with a sealing element in the form of a sealing strip, and of larger flat surfaces. When the sealing strips engage, chambers with very large surfaces result, which are essentially determined by the flat surface parts. The design resembles a reciprocating piston machine without a connecting rod.

A disadvantage of this known construction is the very large ratio of chamber surface to chamber volume sought because of its use as a steam engine, which has an extremely unfavorable thermodynamic effect for use as an internal combustion engine or compressor. A disadvantage of this construction is the course of the force acting on the piston when the different crank angles. Since the piston surface changes only slightly at different crank angles, the force acting on the piston remains almost constant. This is disadvantageous both when used as an internal combustion engine and, in particular, when used as a compressor, since the piston load becomes extremely high. Another disadvantage is the angle at which the resulting force acting on the piston acts on the crankshaft. Angles in the range of 90 °, which achieve high torque with low bearing loads, would be favorable. In the known construction, however, this angle of force always runs essentially perpendicular to the larger flat surface of the piston. The conditions are similar to those in the case of a reciprocating piston in which the force can be transmitted to the crank only over a very small crank angle range under a favorable angle of attack.

The object of the present invention is to provide a piston machine of the type mentioned at the outset, which is thermodynamic and can be improved with regard to the force profile and the force attack angle.

This object is achieved with the features of the characterizing part of claim 1.

In this construction, when the chamber becomes smaller, the sealing element of the piston delimiting the chamber runs over the sealing element of the peripheral wall delimiting the chamber. If the chamber runs with its volume towards zero, its surfaces also run towards zero. This results in an essentially constant ratio of chamber surface to chamber volume with very good thermodynamic properties, in particular with regard to the heat losses at the surfaces with a small chamber volume, that is to say with maximum compression. This also results in an essentially constant force which is introduced into the piston, since with increasing compression pressure the loaded piston surface becomes smaller and smaller. Overloading the machine even with the highest compression is avoided. Extremely high compression internal combustion engines or compressors can be real. The force resulting on the piston is essentially perpendicular to the respective connecting line of the sealing elements delimiting the chamber. This changes its angle in such a way that the direction of the resulting force is essentially perpendicular to the crank over a very large crank angle range, in particular with a larger chamber volume. This results in very smooth, uniform running and high torque in internal combustion engines. The more compact design of the working chambers formed also enables a more compact design of the piston machine. These advantages add up to the generic advantages of this type of construction, which runs with low vibrations due to the parallel rotation principle and, due to the possibility of the compact arrangement of several chambers, enables a high power density and inexpensive construction costs. Since a complete work cycle only requires around 180 ° crankshaft angle, i.e. the piston machine works according to the single-stroke principle, this results in an increased power density compared to other engine systems. Depending on the design, the piston machine can be used as a suction pump, for example a vacuum pump, compressor or as an expansion machine. As an expansion machine, it can be used with external combustion, for example as a steam engine, or with internal combustion as a gasoline or diesel engine. As explained in more detail below, this basic construction is furthermore distinguished by surprisingly versatile variation and combination options which enable a large number of very different, individually adapted variants of piston machines.

The features of claim 2 are advantageously provided. This design option permits particularly simple manufacture with cylindrical running surfaces and cylindrical surfaces of the sealing elements. The enlarged surface of the sealing elements also leads to less wear, among other things.

The features of claim 3 are advantageously provided. The possible variations in the surfaces of the sealing elements are surprisingly diverse. The surface of a sealing element can be very small, so that the sealing element is always almost on the same line on its counter surface. The surfaces can be enlarged. Then there is an improved barrel wear because it is distributed over the surface of the sealing element. Since the counter surface increases with a larger surface area of the sealing element, a larger chamber also results. The surfaces of the sealing elements can be circular in cross-section, but other curved surface configurations, in particular elliptical shapes, are also possible. The mating surfaces of such a sealing element that can be moved in parallel rotation then deviate from the circular shape in accordance with the eccentricity of the surface of the sealing element. If the surface of the sealing element is elliptical, the mating surface is also elliptical. In this way, a variety of chamber designs are possible, which can be adapted to the individual purpose, for example as an expansion chamber, as a pump chamber or as a chamber of a low-pressure or high-pressure compressor. The sealing elements delimiting a chamber on both sides can be of different sizes or also of different shapes. For example, a chamber can be designed in such a way that it is delimited on one side by a very small sealing strip with a circular cross-sectional surface, the counter surface of which is circular in cross section with a radius that is only slightly larger than the crank radius. The sealing strip on the other side of the chamber can be elliptical with very large dimensions and runs on an elliptical counter-running surface which is only slightly larger than the surface of the sealing element. In this way, a large number of very different chambers can be formed, which are adapted to different purposes.

The features of claim 4 are advantageously provided. With their plane of symmetry, the sealing elements can lie parallel to the plane of symmetry of the tread adjacent to them, but can also deviate from this angular position. If, as stated in claim 4, they are arranged tilted outwards, then one results Extension of the counter surface and thus a larger maximum chamber volume or a larger compression ratio.

The features of claim 5 are advantageously provided. Such a sealing element provided with a toothed surface in cross section results in the event of leaks, e.g. when the sealing element is lifted off, a gap with higher flow resistance, ie a higher gas tightness. This is based on gas-dynamic effects when flowing through the gap, which alternately has different widths on the flow path, which leads to turbulence and thus to increased flow resistance.

The features of claim 6 are also advantageous. The chambers can be sealed by means of sealing elements which are designed as rigid integral parts of the piston or of the engine housing. Then, taking into account the manufacturing tolerances, the sealing elements can only be guided to their mating surface at a gap distance, which results in leaks which, for example, limit the maximum compression pressure of a compression chamber. However, this can be sufficient for low-pressure compressors. By designing the sealing elements as spring-loaded sealing strips, as is generally known from engine construction, higher sealing values and thus higher compression pressures can be achieved.

The features of claim 7 are advantageously provided. This ensures that the sealing strips alone or in addition to spring force are acted upon by the gas pressure in the chamber to the rear. The sealing force of the sealing strips thus depends on the chamber pressure and is therefore constantly adapted to the required dimension.

The features of claim 8 are advantageously provided. In this way, the jumping or rattling of the sealing strips can be better suppressed.

The features of claim 9 are advantageously provided. From the adjacent pressurized chamber can under High-pressure gases reach under the mushroom-shaped part of the sealing strip in order to exert a force component increasing the pressure on the sealing strip. As a result, the sealing strips are pressed with increased pressure, depending on the gas pressure to be sealed, as a result of which their tightness is improved.

The features of claim 10 are advantageously provided. Surprisingly, the parallel rotation of the piston results in the possibility of simultaneously engaging a running surface provided on the piston with two adjacent running surfaces of the peripheral wall. Thus, chambers are formed simultaneously with both running surfaces of the peripheral wall, one of which, depending on the running direction, works as a compression chamber and the other as an expansion chamber. This results in e.g. the possibility of creating a self-compressing internal combustion engine in which one chamber always compresses air, which can be burned in the other chamber after external or internal mixture formation. The separation between the compression chamber and expansion chamber provided in such an internal combustion engine results in cooling advantages with better partial efficiency levels of the engine. After the expansion has ended, the combustion chamber opens and the hot gases are immediately blown into the low-pressure outlet with fresh air. The residual hydrocarbons are re-burned by this flushing process, so that the pollutant emission is considerably lower than with other combustion engines. The expansion chamber can also be used as a suction pump, e.g. can be used as a vacuum pump. The result is a piston machine in which the compression chamber supplies compressed air, while the expansion chamber works as a vacuum pump. Such a piston machine can be advantageous in certain manufacturing processes in which compressed air and vacuum are required at the same time.

The features of claim 11 are advantageously provided. At largely any angle, essentially determined only by the space required, can extend over the circumference of a larger motor distributed several chamber arrangements can be provided, each with a tread provided on the piston and one or two treads provided on the peripheral wall. Thus, for example, a compressor with several chambers working in parallel or an internal combustion engine with several expansion chambers or with, for example, the same number of expansion and compression chambers can be provided. An interesting possibility highlighted only as an example is the provision of a self-propelled compressor with, for example, an expansion chamber operating as an internal combustion engine and a plurality of compression chambers driven by the latter.

The features of claim 12 are also advantageous. In this way, in a manner known, for example, from Wankel motors, by designing the machine with multiple disks, the output can be increased in accordance with the number of parallel disks. By means of angular offset of the cranks and / or angular offset of the treads in the disks, the degree of uniformity of the machine can be improved and possibilities can be created in the case of training as an internal combustion engine using the angular misalignment between the disks compression chambers of the one disk directly, without pressure intermediate storage. tion to let expansion chambers act on the other pane.

The features of claim 13 are also advantageous. With the generic design principle, in which the chambers always work only as either a compression chamber or an expansion chamber, valves are always required in the high-pressure channels. These can advantageously be designed as valves controlled in synchronism with the piston travel, for example in the form of globe valves or in the form of rotary valves.

Advantageously, the valves of compression chambers can also be designed as one-way valves, for example as spring-loaded flap valves, their spring load stipulating the desired maximum pressure. The features of claim 15 are advantageously provided. The joint mounting of two crankshafts on the piston already results in an angle synchronization, which is very sensitive to play and can lead to jamming. The game dependency depends on the number of crankshafts, so that this synchronization alone can be sufficient with more than two crankshafts. A completely exact synchronization results when the crankshafts are externally coupled via gear sets, so that two crankshafts are then sufficient to support a piston without there being a risk of jamming.

The features of claim 16 are advantageously provided. A piston of this type is cooled as it moves by contact with the gas in the barrel. For this purpose, the piston can be provided with ribs or with openings flushed with gas.

The features of claim 17 are advantageously provided. With this design, the heat transfer from the chambers, which are particularly heat-loaded in the case of an internal combustion engine or in the case of a highly compressed compressor, or the running space to the bearings is hindered, so that the bearings remain cool and there is the possibility of storing the bearings in simple design, without cooling, for example with permanent lubrication. The parts that hinder heat transfer can be used as parts with a long path and intermediate cooling, e.g. B. air ribs, be formed or with heat-insulating intermediate layers.

The features of claim 18 are advantageously provided. Since the piston is supported on at least two cranks, one-sided mounting may be sufficient and leads to a significant simplification of construction and a more compact design of the machine.

The invention is shown, for example and schematically, in the drawings. Show it: Fig. 1 in section along the axis of one of the crankshafts

Line 1 - 1 in Fig. 2 is a two-disc internal combustion engine according to the invention,

2 shows a section along line 2 - 2 in FIG. 1,

3 - 9 representations according to FIG. 2 in successive angular positions of the piston,

10 - 15 in the view corresponding to FIG. 2 different piston variants of different embodiments,

16 is an enlarged illustration of the upper part of FIG. 2 for clarification,

17 shows a representation corresponding to FIG. 16 of an embodiment variant with sealing strips of different sizes,

18 shows a representation corresponding to FIG. 17 with an inclined sealing strip,

19 shows a schematic section through a sealing strip with a toothed surface in contact with its mating surface,

Fig. 20 shows a section in the cutting direction of Fig. 1 by a

Design variant of a piston machine with a running chamber having a chamber arrangement and thermal insulation to the crankshaft bearings,

21 shows a section corresponding to FIG. 2 through a piston machine with two smaller and one very large chamber,

22-27 representations according to FIG. 2 in successive angular positions of the piston, 28 in section according to FIG. 2 a further variant of a

Piston machine with cylindrical and elliptical running surfaces and

29 shows a representation of the lower part of FIG. 21 in a variant with several sealing strips.

1, 2 and 3 and in particular the enlarged representation of FIG. 16, the basic construction of the illustrated embodiment is first explained. It is an internal combustion engine with a housing 1, which is shown in one piece for the sake of simplification of the drawing, but which in a practical embodiment has to be made in several pieces, for example, in a disk-like manner, for assembly purposes. Two parallel, identical crankshafts 2, 2 'are mounted in the housing, of which the crankshaft 2' is visible in FIG. 3. The crankshafts pass through two disk spaces 3, 3 'arranged one behind the other in the manner of disks, of which the barrel space 3 can be seen opened in the section of FIG. 2.

The crankshafts 2, 2 'have cranks 4 in each running space, on the crank pin 5 of which a piston 6, 6' is mounted in each of the running spaces 3, 3 '.

As shown in FIG. 2, the crankshafts 2, 2 'are identical in terms of their cranks for the piston 6 shown, in particular with the same crank radius and also with an identical angular position. The crankshafts therefore rotate in synchronism with the angle. For this purpose, corresponding gear sets 7, 7 'are provided on one or both crankshaft ends. It can be seen from FIG. 1 that the crankshaft 2 passes through the end wall of the housing 1 at its end lying on the gear set 7 and. there carries a drive pulley 8 provided for example. The gear set 7 'drives an output shaft 8'. 2 to 9 show that by the bearing of the piston 6 on the crank pin 5 of the two angularly synchronized crank shafts 2, 2 ', the piston executes an orbit which, as in several successive orbital phases in FIGS. 2 to 9 shown, can be referred to as parallel rotation. The piston is parallel to its other positions in all angular positions of the crankshafts. Each point of the piston rotates with the radius of the cranks 4, but in each case around its own center. Therefore, more than two crankshafts can also be used to support a piston, as shown in FIG. 13 in an embodiment variant of a piston which runs on the crank pin of three crankshafts which are coupled in an angularly synchronous manner.

The construction is first further explained with reference to FIG. 2. The running space 3 is delimited by parallel surfaces 9, which are perpendicular to the crankshafts 2, 2 ', and by a peripheral wall 10, which is perpendicular to the parallel walls 9 everywhere.

In the peripheral wall 10, a tread 11 is provided, which is designed in the form of a half cylinder in the section of FIG. 2, that is to say semicircular. At one point of the piston 6, a sealing strip 12 is arranged as a sealing element, which, when the piston 6 rotates in parallel, as shown in FIGS. 2 to 9, describes a circle, on the upper half of which it slides in contact with the running surface 11 .

At the right end of the tread 11 in FIG. 2, a sealing strip 13 is arranged on the peripheral wall 10 as a further sealing element. Associated with this is a running surface 14 in the piston 6, which likewise has a semi-cylindrical shape with the same radius of the running surface 11. Draw imaginary connecting lines through the end points of the running surface 11 and through the end points of the running surface 14 and compare these connecting lines in the phases of the circulation 2 to 9, it can be seen that these connecting lines are always parallel to one another.

By comparing the successive phases of FIGS. 2 to 9, it can also be seen that in the crank angle position of FIG. 8 at the same time the sealing strip 12 of the piston 6 comes into engagement with the start of the running surface 11 of the peripheral wall lying on the left in the figures and the sealing strip 13 comes into engagement with the running surface 14 formed on the piston. The sealing strips 12, 13 then slide (FIGS. 9, 2, 3) on their mating surfaces 11, 14 to the opposite end of the tread until they run against one another, as shown in FIG. 4. The sealing surfaces then come out of engagement, as shown in FIGS. 5 to 7. A new intervention begins in FIG.

A chamber, which is enclosed on all sides, is formed between the running surfaces 11 and 14 and is delimited by the parallel surfaces 9 and the running surfaces 11 and 14. This chamber is sealed by the sealing strips 12 and 13 and additionally by in the side surfaces of the piston 6 provided circularly arranged side sealing strips 15 which seal against the parallel surfaces 9.

This chamber formed by engagement of the treads 11 and 14, which is referred to below as chamber 11.14, changes its volume when the crankshafts 2, 2 'rotate according to the sequence of FIGS. 2 to 9, that is to say clockwise. In Fig. 7 the chamber 11.14 is open. It closes in Fig. 8 with maximum volume, which is calculated from the distance between the parallel surfaces 9 and essentially a circular cross-section with the radius of rotation of the cranks 5. If you follow Figs. 9, 2, 3 and 4, you can see that the chamber 11.14 is reduced to substantially zero and then, as shown in FIG. 5, opens again in order to close again in FIG. 8.

In chamber 11.14, the direction of rotation of the crankshafts shown clockwise is a compression chamber. In the open position (FIGS. 5 to 7), it is connected to the running space 3 and can absorb gas of low pressure, which flows in, for example, through a low-pressure inlet channel 16 in the housing 1. 8, 9, 2 and 3, the gas in the chamber 11.14 is compressed and finally ejected through a high-pressure outlet channel 17, the opening of which is shown in the parallel wall in FIGS. 2 to 9, at a greatly increased pressure.

As shown in FIGS. 2 to 9, a further tread 18 is arranged laterally next to the previously described tread 11 in the peripheral wall 10, which is mirror-symmetrical to the sealing strip 13 and is identical to the tread 11. The left and right end points of the treads 11 and 18 and the common middle end point lie on a line. At the end of the running surface 14 of the piston 6 opposite the sealing strip 12 there is a further identical sealing strip 19.

Comparing the orbital phases of the piston 6 according to FIGS. 2 to 9, it can be seen that whenever the sealing strip 12 runs on the running surface 11 in the chamber 11.14 and at the same time the sealing strip 13 runs on the running surface 14, the sealing strip also runs 19 engages with the tread 18 on this. At the same time as chamber 11.14, a chamber is also formed with the same terminology as chamber 18.14, which, however, experiences a volume change in the opposite direction to chamber 11.14. If the chamber 11.14 reduces its volume when the piston 6 rotates, the volume is simultaneously increased in the chamber 18.14. The chamber 18.14 therefore forms an expansion chamber which is initially open (FIGS. 5 to 7), begins with the minimum volume in FIG. 8 and then increases its volume to the maximum volume until FIG. 4, and then closes (FIG. 5) open and close again at Fig. 8.

A high-pressure inlet duct 20 also opens into the chamber 18.14, but in contrast to the high-pressure outlet duct 17, it does not compress to the outlet! n gas, but is provided for the inlet of compressed gas, which is relaxed during the working cycle of the chamber 18.14.

The construction described so far can be used as an internal combustion engine, which can be seen after the single-stroke zip works because it only requires 180 ° crankshaft angle for a complete work cycle.

Air flowing in through the low-pressure inlet duct 16 is enclosed in the chamber 11.14, compressed and fed through the high-pressure outlet duct 17 to a pressure accumulator (not shown). From this, the compressed air is fed through the high-pressure inlet channel 20 to the chamber 18.14 at a point in time when the chamber volume is small or through the high-pressure inlet channel 20 'to the chamber 18'.14', where it is exploded. For this purpose, fuel, e.g. Petrol or diesel fuel supplied with injectors, not shown, e.g. in the form of an intake manifold injection into the high-pressure inlet duct or in the form of a direct injection directly into the chamber. A spark plug or injection nozzle can be arranged in the stepped bore 21 shown. After expansion and opening of the chamber 18.14, the burned gas can escape from a low-pressure outlet channel 22 opposite the low-pressure inlet channel 16.

Deviating from the arrangement shown in FIG. 2, the expansion chamber 18.14 and the compression chamber 11 M4 'can be omitted, for example in a simpler embodiment. There is then still a compression chamber 11.14 and an expansion chamber 18 '.14', which can work together in the manner described above.

The internal combustion engine can also operate on the diesel principle. An injection nozzle is then to be seen in the stepped bore 21, which injects compressed air supplied to the chamber 18.14 at the time when the chamber volume is small. Since very large volume changes can be achieved with the chambers 11.14 and 18.14 shown, the chamber 11.14 can be used to bring the air to the required pressure of 30-60 bar, for example. As shown in FIG. 2, the high-pressure channels 17 and 20 are arranged in the immediate vicinity of the sealing strip 13 provided between the chambers 11.14 and 18.14 and provided on the peripheral wall 10, that is to say in the area of the minimum chamber volume. The low-pressure channels 16 and 22, which serve for the inlet and outlet, lie opposite each other in the area in which the associated chambers 11.14 and 18.14 are to receive or release gas. The rotation of the piston 6 in the clockwise direction favors purging from the low-pressure inlet duct 16 to the low-pressure outlet duct 22, so that mixing of fresh and exhaust gases is avoided.

The high-pressure channels 17 and 20 must have valves which, in the case of the compression chamber 11.14, must be opened at maximum compression for the outlet of the high-pressure gas and, in the case of the expansion chamber 18.14, must close after the high-pressure gas has been admitted. For this purpose, valves controlled in synchronism with the rotation of the crankshafts 2, 2 'can be provided. In the illustrated embodiment, FIG. 1 shows rotary valves 23, 23 'which are driven by the respective gear set 7, 7' synchronously with the crankshafts and which control the high-pressure channels, which cannot be seen in the section of FIG. 1. In the case of a compression chamber, one-way valves which are permeable in the gas direction can also be provided for this purpose, which are designed, for example, as spring-loaded flutter valves.

The structure of the sealing strips 12, 13 and 19 will be described with reference to FIG. 3 and in particular FIG. 16. They are essentially identical and are described in detail using the example of the sealing strip 13.

The sealing strip 13 has a surface 24 which is circular in cross section, the center 25 of which is on a radius to the center 26 of the running surface 11 which corresponds to the radius of the cranks 5 of the crankshafts 2, 2 '. The radius of the surface 24 of the sealing strip 13, based on its center point 25, must be added to the circumferential radius of the cranks 5 to the radius of the tread 11, based on their center point 26. The radius of the running surface 14 of the piston is identical to that of the running surface 11. The same applies to the already described running surface 18. When the sealing strips slide on their respective counter-running surfaces, ie sealing strip 12 on counter-running surface 11, sealing strip 13 on counter-running surface 14 and sealing strip 19 on mating surface 18, one can see in the sequence of FIGS. 2 to 9 that the sealing strips run with a constantly changing contact line on the mating surfaces, which each result as an envelope curve of the rotation of a sealing strip with parallel rotation of the piston 6.

The sealing strips 12, 13 and 19, which are essentially identical to one another, are, as explained in FIGS. 3 and 16 using the example of the sealing strip 13, slidably mounted with a slide 27 in a sliding guide and form at their end opposite the surface 14 a piston 28 which slides in a cylinder with spaces 29 and 30. In both rooms 29 and 30 springs are applied to the piston 28 from above and below (indicated schematically in FIG. 16 with wavy lines) which hold the sealing strip in a defined central position. In a preferred embodiment, the space 30 located outside the piston 28 can be connected to one of the adjacent chambers with a bore, not shown, in order to be acted upon by this with high-pressure gas, which presses the sealing strip against its running surface with additional preload for the sealing contact. Such a bore 100 is shown in dashed lines in FIG. 17. It serves to pressurize the sealing strip 120.

The construction described so far can run as an internal combustion engine with compression chamber 11.14 and expansion chamber 18.14. 1 to 9, however, an embodiment is shown in which this chamber arrangement is provided twice in the symmetrical position with respect to the crankshafts 2, 2 '. Symmetrically opposite to the crankshafts 2, 2 ', two chambers are provided for the already explained chambers 11.14 and 18.14, which are identified by identical reference numerals, each with a comma. The position of the treads 11 'and 18' is, as Fig. 2 shows, exchanged with respect to the running surfaces 11 and 18, since the chamber 11 '.14' is a compression chamber corresponding to the chamber 11.14, corresponding to the direction of rotation of the piston 6, while the chamber 18 '.14' is an expansion chamber . Accordingly, the position of the supply and discharge low-pressure channels 16 ', 22' and the high-pressure channels 17 'and 20' is reversed.

The construction shown overall in FIGS. 1 to 9 thus forms an internal combustion engine which has two compression chambers and two expansion chambers per disk, that is to say a total of four compression and four expansion chambers. 1 shows, the cranks 4 of the crankshafts in the running spaces 3 and 3 'are angularly offset from one another. The pistons 6, 6 'thus run with a phase shift. This makes it possible, for example, for the compression chambers of the one disc to release high-pressure gas at a point in time in which the expansion chambers of the other disc require high-pressure gas.

In an embodiment not shown, an internal combustion engine can also have more than the two disks shown.

In an embodiment not shown, only one double-chamber arrangement with chambers 11.14 and 18.14 can also be provided in a pane, for example. A corresponding piston with only one running surface 14 is shown in FIG. 10. In a further simplified embodiment, for the piston shown in FIG. 10 with only one running surface 14, only one counter running surface, for example the running surface 11, can be provided in the peripheral wall 10. It is then a pure compressor which has to be driven externally and which has only one compression chamber per disc. In a corresponding embodiment, as shown in FIG. 2, such a pure compressor can also have two compression chambers per disc (but no expansion chambers).

In another embodiment, for example, only expansion chambers can be in one pane and only com- ponents in another pane. compression chambers may be provided. As these few examples show, the invention offers considerable scope for variation.

For example, a self-propelled compressor can be designed such that, for example, in the two disks shown in FIG. 1, only one disk has an expansion chamber which drives the compressor according to the internal combustion principle, but each disk has two compression chambers. As calculations show, one expansion chamber is sufficient to drive four compression chambers.

More than two chamber arrangements can also be provided on the circumference of a running space, each of which can consist of either an expansion chamber or a compression chamber or an expansion and a compression chamber. This is shown by the illustrations in FIGS. 10 to 15.

10 shows a piston with only one running surface 14, with which a single or double chamber arrangement can be provided. 13 shows a piston with three running surfaces for three such chamber arrangements. For comparison, FIG. 11 shows the piston described in FIGS. 1 to 9 for two such chamber arrangements.

FIG. 12 shows a piston with two running surfaces, which, however, when compared with FIG. 11, are arranged obliquely to the connecting line of the crankshafts. 14 and 15 show that larger numbers of chamber arrangements are easily possible. The geometric conditions only need to be taken into account in terms of space requirements. The parallel rotation movement of the piston makes a largely arbitrarily large number of chamber arrangements possible per piston.

In the embodiment of FIG. 12, it should also be noted that the forces exerted by the running surfaces 14, 14 'of the piston on the crank pin 5 at a different angle than in the embodiment of FIGS. 2 to 9 and 11 act. This can also be achieved in other ways. Thus, in the embodiment shown in FIG. 2, the chamber arrangement with the chambers 11.14 and 18.14 can be arranged inclined at an angle to the connecting line of the crankshafts 2, 2 '. The connecting line of the sealing strips 12 and 19 of the piston 6 is then at an angle to the connecting line of the crankshafts 2, 2 '. Correspondingly, the running surfaces 11 and 14 are to be arranged tilted at an angle such that the connecting line of their end points is parallel to the connecting line of the sealing strips 12 and 19 of the piston 6. In this way, too, the introduction of the forces arising in the chambers into the cranks can be designed at optimized angles.

The same effect of more favorable introduction of force into the cranks can also be achieved by increasing the distance between the crankshafts while maintaining the geometry of the chambers.

FIG. 17 shows an embodiment variant whose differences from the construction described above can be seen in comparison with FIG. 16. The same parts are provided with the same reference numerals.

Deviating from the embodiment shown in FIG. 16, the sealing strip 120 seated at the left end of the running surface 14 of the piston 6 is greatly enlarged, as the comparison with the sealing strip 12 of the construction according to FIG. 16 shows. In the exemplary embodiment, it is doubled in its radius, that is to say in its overall dimensions.

Correspondingly, the left stationary tread 110 is enlarged compared to the tread 11 shown in dashed lines, which corresponds to that of the construction in FIG. 16. The center point of the original tread 11 was at 26. The center point of the new tread 110 was at 260. The magnification to the left is clearly asymmetrical, as the lateral shift of the center points 26 and 260 shows. This results from the fact that the sealing strip 120 is not only doubled in its radius, but is also offset laterally with its center from 25 to 250. Out of it this results in a shape of the new tread 110, which is widened to the left and up, but merges at the right end towards the sealing strip 13 into the original tread 11.

Otherwise, the construction is completely unchanged, in particular with regard to the entire geometry of the crankshafts, the piston 6, its running surface 14 and the running surface 18 of the stationary chamber 18.14 on the right.

The newly formed enlarged chamber 110.14 is distinguished from the original construction according to FIG. 16 by a maximum volume increased by 25% in the exemplary embodiment and a correspondingly increased maximum compression ratio. Otherwise the mode of operation of the overall construction remains unchanged. The sequence in the individual phases according to FIGS. 2 to 9 is unchanged.

The enlarged sealing strip 120 can also be designed with other dimensions different from the sealing strips 13 and 19, for example even larger or somewhat smaller. The size of the new tread 110 must be adjusted accordingly.

With this construction it is possible to make the two chambers of different sizes without changing the other geometry in the double-chamber arrangement 18.14, 110.14. If one sees the arrangement mirror-symmetrically right / left interchanged, ie with an enlarged sealing strip 19, the right chamber would be larger than the left. It is of course also possible to design both sealing strips 120 and 19 of the piston 6 differently from the sealing strip 13 provided stationary on the peripheral wall 10, either both the same or different from one another. Then the sealing surface 14 of the piston would remain unchanged. Both stationary sealing surfaces 11 and 18 would have to be changed accordingly.

Fig. 18 shows a variant of the construction of Fig. 17 in the same representation. Corresponding parts are provided with the same reference symbols. The reference numbers of changed parts have been changed also retained, but with a comma.

As can be seen immediately, the change relates to the sealing strip 120 'located at the left end of the chamber 110'.14, that is to say at the left end of the running surface 14 of the piston 6.

If one looks at the enlarged sealing strip 120 of FIG. 17 again for comparison, it can be seen that this sealing strip 25, like the sealing strip 25 located at the right end of the running surface 14 of the piston, lies exactly parallel to the plane of symmetry of the neighboring running surface 14 with its plane of symmetry . As shown in FIG. 17, this results in a maximum circumferential angle of the counter surface 110 of the chamber of approximately 180 °. In the case of the smaller sealing strip 12 shown in dashed lines, the corresponding smaller counter-running surface 11 can only be traveled over about 180 °. This limits the maximum chamber size shown in FIG. 17.

18 shows that the enlarged sealing strip 120 'with its dash-dotted plane of symmetry is arranged at an oblique angle a with respect to the plane of symmetry of the adjacent tread 14, which is also shown with a dash-dotted line. In addition, the comparison with FIG. 17 shows that the circular sector surface of the sealing strip 120 'is formed over a somewhat larger angular range. This results in the possibility of guiding the sealing strip 120 'over an angular range likewise enlarged by a beyond 180 ° in contact with the correspondingly lengthened counter-running surface 110'. As a result, the maximum chamber volume, as the comparison of FIGS. 17 and 18 shows, can be increased considerably again without changing the crankshafts.

In particular, as shown in FIG. 18, the left-hand chamber 110'.14 can again be greatly enlarged, while the right-hand chamber 18.14 is kept small, since it has the sealing strip 25 arranged at 90 ° and also a substantially smaller surface area of it Circular sector. 18, the smaller sealing strip 12 'is also shown (dashed) within the enlarged sealing strip 120' at the same angle. Here, too, the other angular arrangement results in a correspondingly enlarged counter-running surface 11 'with a corresponding chamber enlargement.

19 shows in section a sealing strip 300 which corresponds in its basic construction to the design of the sealing strips 12 or 120. However, their surface is ribbed, these ribs extending in the longitudinal direction of the sealing strip 300 and being more or less fine with a corresponding number of ribs. The sealing strip 300 runs on its counter surface 301.

In the case shown, a chamber of higher pressure is located to the left of the sealing strip 300. In the event of leaks, for example when the sealing strip 300 is lifted off the counter-running surface 301, gas flows in the direction of the arrows shown through a gap between the sealing strip 300 and the counter-running surface 301. This leads to loss of compression.

The ribbed surface of the sealing strip 300, however, severely impedes the leakage gas flow flowing in the direction of the arrow, since it has to flow across valleys and mountains of the ribs due to the ribbed surface. Turbulence occurs in the valleys and thus the gas flow is slowed down, consequently increasing the flow resistance of the gap formed between the sealing strip 300 and the counter-running surface 301. This improves the tightness when the sealing strip lifts off.

Additionally or alternatively, the tightness of a sealing strip can also be improved, in particular also in the case of sealing strips with a smooth surface, that is to say, for example, the sealing strip 120 of FIG. 17 if the lifting off is prevented in another way. Lifting usually occurs when the sealing strip vibrates or rattles in its resilient mounting in the event of smooth running malfunctions. Such vibrational movements can can be prevented by shock absorption. For this purpose, shock-absorbing devices can be provided in the resilient bearing seat of a sealing strip, e.g. B. hydraulic damping devices in the manner of conventional hydraulic piston shock absorbers.

20 shows a construction variant of a piston machine, the basic construction of which is first described in comparison with the construction according to FIGS. 1 and 2.

The construction shown in FIG. 20 is shown in longitudinal section, that is, in the section corresponding to FIG. 1. The construction has only one running space with a piston 406 with running surface 414 and sealing strip 412 (see FIG. 2). The part of a running surface 418 with a stationary sealing strip 413 can be seen on average around the circumference of the running space. The chamber is open. The piston is therefore approximately in the position according to FIG. 5. In contrast to the piston of FIGS. 1 and 2, the piston 406 of the construction of FIG. 20 only forms a running surface on its upper side (like the piston of FIG. 10). But there are also all other piston shapes, such as 11-15 possible.

The piston 406 runs between the parallel walls 409 of the running space in a housing 401. This, arranged at a distance from the running space, has bearings for crankshafts 402 which have crank pins 405 connected via cranks 404 on which the piston 406 is spaced apart arranged parts ge is stored.

The construction corresponds essentially to the basic construction of FIG. 1 in a single-disc design, that is to say with only one running space and only one piston. However, the bearings of the crankshaft in the housing 401 and of the piston 406 on the crank pin 405 are each provided at a distance from the running space and from the piston.

The housing 401 has parts interposed between the running space 418, 413, 409 and the bearings on the crankshaft 402 420, through which heat generated in the running space must be conducted in order to reach the crankshaft bearings. These parts 420, as shown in FIG. 2, can be made very long, as a result of which the heat flow is impeded. Cooling devices, such as, for example, inner cooling channels or cooling fins on the surface which serve for air cooling, can be provided at this point in order to reduce the heat transfer from the running space to the bearings. Furthermore, the parts 420 can be heat-insulating, for example. With one or more of these possibilities, the heat transfer from the heat-stressed running space to the crankshaft bearings can be drastically reduced. These can therefore be provided as simple ball bearings with permanent lubrication, which do not require any special cooling, as is otherwise necessary in crankshaft bearings of internal combustion engines or compressors for reasons of thermal stress.

These considerations also apply to the bearings of the piston 406 on the crank pin 405. Here too, intermediate parts 421 are provided between the piston 406 and its bearing points, through which the heat from heat-loaded pistons 406, ie from its heat-loaded bearing surface 414, is increased the bearings on the crank pin 405 is passed. These intermediate parts 421 can, like the parts 420 of the motor housing 401, be provided with cooling devices, for example with air-cooled ribs or heat-insulating. Then ball bearings can also be provided on the crank pin 405 without the usual cooling means cooling. By saving a corresponding cooling circuit, the engine design can be considerably simplified.

In piston machines with thermally highly loaded chambers, heat dissipation can of course also be taken care of in other ways, for example by water cooling channels in the housing in the vicinity of the treads provided there, and by liquid cooling of the piston, which can take place, for example, with oil channels in the piston the two crankshafts via the bearings are locks. However, air cooling in the housing is also possible, for example by means of external ribbing. The piston can also be sufficiently cooled with gas cooling alone.

If, for example, in FIG. 2, the piston 6 shown there is seen, it can be seen that it rotates continuously in the running space and is in intensive gas contact with the constantly flowing cool fresh gas. If the piston is heavily ribbed outside its running surfaces 14 and 14 ', for example in its surface, sufficient gas cooling of the piston can be brought about. The piston can also be provided with openings, for example (see FIG. 2) of an opening which runs approximately in the imaginary line between the openings 16 and 22 of the housing 1 in the upper part of the piston 6 between its running surface 14 and the bearings passes through the cranks 5 and which is flowed through by air as the piston rotates.

Another possibility for simplifying the construction is to be illustrated with the aid of FIG. 20. If the crankshaft 402 shown on the right in FIG. 20 is completely omitted, the piston 406 is only supported on the crankshaft 402 shown on the left relative to the housing 401. It has to be taken into account that according to the construction principle of this piston machine, as the illustrations in FIGS. 10 to 15 show, the piston on each side is always on two or more cranks, only one of which is shown in the section of FIG. 20 is. A one-sided mounting on two or more cranks can, under certain circumstances, be sufficient for the exact mounting of the piston, as a result of which the construction can be considerably simplified.

To do this, it is better to keep the distance between the piston and its bearing on the crank pin as short as possible or to store the piston directly on the crank pin, and the piston should have sufficient lateral guidance on the parallel surfaces of its running space, as shown in FIG. 1 shows. If the construction shown in FIG. 1 is separated in the section line 2-2, the part of the con- Piston 6 'located structurally run even with one-sided bearing on two crankshafts with sufficiently precise guidance.

1 to 19, the sealing elements which delimit a chamber are always shown to be substantially smaller than the crank radius of the piston. In the embodiment of FIG. 17, for example, the sealing strips 13 and 19 have a surface radius which is approximately a quarter of the crank radius. The larger sealing strip 120 has a surface radius that is approximately half the size of the crank radius. In the previously described embodiments, the sealing strips are always spring-loaded, as is also shown in FIG. 17.

In a departure from these embodiments, the sealing elements can be designed with significantly larger surfaces compared to the crank radius, and they can also be designed as rigid parts of the piston or the housing wall without suspension. This is explained in an example in FIG. 21.

FIG. 21 shows a cross section to the crankshafts of a housing 501 in which a piston 506 for parallel rotation is mounted on crank pins 505 of three crankshafts. A very large running surface 530 is formed on the peripheral wall of the running space shown, which has the shape of a circular section in cross-section and on one end of which a sealing strip 531 of small cross-section is arranged on the housing side. A running surface 532 serving as a counter surface for the sealing strip 531 is provided on the piston 506. This extends from the corner at the location of the sealing strip 531 to the circumferential point marked with a line 533, that is to say over almost 180 °. This is followed by a part of the piston 506 which is circular in cross section, namely between the marking line 533 and the corner 534. This surface part of the piston which is circular in cross section forms the sealing element 535 which, when the piston 506 rotates in parallel, on the running surface 530 expires, with parallel rotation of the piston 506 clockwise between the beginning of the Tread 530 at low pressure inlet duct 516 to the end of tread 530 at sealing strip 531.

A working chamber 530.532 is hereby formed, which is delimited by the running surfaces 530 and 532 and the sealing elements 531 and 535. During the parallel rotation, the sealing element 535 runs on the running surface 530 while forming a chamber, while the sealing strip 531 runs on the running surface 532. The same chamber formation conditions are present as are described in the previous embodiments. In contrast to this, only the ratio of the surfaces of the sealing elements is chosen to be very large, and the sealing element 535 has a surface radius which is much larger than the crank radius. In addition, the sealing element 535 is not cushioned in this embodiment. It can only seal against its counter surface 530 with a gap required due to play. The chamber 530.532 can therefore essentially only be used as a low-pressure compression chamber, but has a very large chamber volume and can therefore be used to compress large amounts of air to low pressures.

When using this chamber 530.532 as a compression chamber, the compressed gas can be obtained through an outlet channel 536 with valve 537.

In the embodiment of FIG. 21, the chamber 530.532 is combined with the two chambers 110.14 and 18.14 of the embodiment in FIG. 18. For this purpose, the tread 14 is provided on the piston 506, on both ends of which the sealing lines 19 and 120 'are seated Zen. The housing 501 forms the running surfaces 110 ′ and 18 here. Details of these two chambers have been omitted for the sake of simplifying the drawing.

With a corresponding design of the gas guide channels, which is not explained in detail, the chamber 530.532 can be used as a low-pressure compression chamber, while the chamber pair 110'.14, 18.14 in the manner described above Forms internal combustion engine that drives the compressor. The low-pressure compression chamber 530.532 can, however, also serve as a pre-compression chamber, the gas pre-compressed therein being suitably supplied to the compression chamber 110'.14 for post-compression. The result would be a two-stage compressor that can reach very high outlet pressures. When used as an externally driven compressor, the expansion chamber 18.14 could be omitted.

21 shows a further construction detail, which is of great advantage when used as an internal combustion engine in order to avoid flushing losses. In the construction shown, the chamber 18.14 forms the expansion chamber in the clockwise direction of movement of the piston. After opening this chamber, in the position of the piston 506 shown in FIG. 21, the burned exhaust gas is to leave the machine through the low-pressure outlet channel 522, if possible without mixing with the fresh gas of the low-pressure inlet channel 516. For this purpose, a running surface 540 is formed on the separating web between the low-pressure channels 516 and 522, and a sealing element 542, which seals during the critical crank angle range at which the expansion chamber 18.14 opens, is formed on a nose 541 of the piston 506 runs on the tread 540 and creates a gas seal between the low-pressure channels 522 and 516, so that mixing of exhaust gas and fresh gas is avoided in this critical time range.

For a better understanding of the mode of operation of the construction shown in FIG. 21, several phases of a work cycle are shown in FIGS. 22-27. To simplify the drawing, the reference numerals are omitted. These result from FIG. 21.

In the position of FIG. 25, the large sealing element 535 of the piston 506 engages with the tread 530 and begins to pump fresh gas clockwise. In the angular position according to FIG. 26, the large chamber 530.532 closes and begins to compress through the position of FIG. 27 to the position 22 with gas outlet from the outlet channel 536. The gas is fed to the high-pressure compression chamber 110 '.14, which has just closed at this time, and is compressed again by this. The high-pressure compression chamber 110'.14 also receives fresh gas without the help of the low-pressure compression chamber, so that the low-pressure compression chamber can also be used for other purposes.

22 begins with the expansion according to FIG. 22 and expands to the position shown in FIG. 24. Now the nose 541 with its sealing element 542 comes into sealing engagement with the tread 540 and blocks the low-pressure channels 522 and 516 against one another. When the expansion chamber 18.14 now opens, exhaust gas flows out of the low-pressure outlet channel 522 as without mixing with the fresh gas. The flushing of the exhaust gas is supported in the positions according to FIGS. 25 and 26 in that the piston with its large sealing element pump 535 is already pumping fresh gas clockwise.

In the previously described embodiment variants, the sealing elements or sealing strips are always designed with surfaces which are circular in cross section. Correspondingly, the running surfaces result as surfaces which are circular in cross-section and are covered by these sealing elements during the parallel rotation.

However, other surface shapes are also possible, such as in particular conic sections, for example sections of circles, ellipses and parabolas, but also spiral sections. The mating surfaces that are coated by such sealing element shapes are similar to the surface shape of the sealing element. They result in a simple construction by extending the rays emanating from a common point by the crank radius beyond the surface of the sealing element. As has already become clear from the exemplary embodiments described above, this results in the case of a circular surface of the sealing element treads circular in cross section. An elliptical surface of the sealing element results in an elliptical running surface. Such an example with elliptical surfaces is explained in FIG. 28.

Fig. 28 shows a simple low pressure compressor with two symmetrically arranged identical chambers.

The compressor has a housing 601 in which a piston 606 rotates clockwise on three cranks 605. The revolving curves of the crank centers are shown with circles.

On the underside of the piston shown, it forms a sealing element 635 with a cross-sectionally elliptical surface, which extends from the marking line 633 to the marking line 634. It connects to a circular cross-sectional tread 632 at 633. A sealing strip is on the peripheral wall of the housing 601

631 small circular cross section arranged on the tread

632 of the piston expires. On the peripheral wall, a running surface 630 is then formed on the sealing strip 631, which extends from the sealing strip 631 to a low-pressure inlet channel 616.

When the piston 606 rotates in a clockwise direction, the sealing element 635, which is designed as an elliptical section, runs from the angular position shown in FIG. 28, in which it comes into first contact with the running surface, to the contact with the sealing strip 631 and forms the sealing boundary of the chamber 630,632. At its other end, this chamber is sealed by the sealing strip 631 in contact with the tread 632.

High pressure gas from this chamber is discharged in the direction of the arrow through an outlet duct 636 with valve 637.

A second chamber 630'.632 'is provided symmetrically on the upper side of the piston 606, which works alternately with the chamber 630.632 described first when the piston 606 rotates. In the embodiments of FIGS. 21 and 28, the sealing elements 535 and 635, which have very large dimensions in relation to the crank radius, form a considerable peripheral part of the piston 506 and 606. These sealing elements 535, 635 are therefore rigid with Surfaces connected to the piston are formed and can only seal against their mating surface with a gap seal.

It would of course be desirable to also design these sealing elements as spring-loaded sealing strips which achieve a better sealing effect. With such large surfaces of a sealing element, however, this is difficult, even if it is technically possible.

An advantageous solution to this problem is shown in FIG. 29, which shows a section of the lower part of FIG. 21 in a variant. The same reference numerals are used as in FIG. 21. .

In order to achieve a better sealing of the circular-shaped sealing element 535 on the running surface 530, sealing strips 745 are resiliently mounted in the surface of the sealing element 535, namely three sealing strips in the example shown.

The sealing strips 745 are resiliently mounted in such a way that they protrude slightly above the surface of the sealing element 535 and come into good sealing contact with the running surface 530, while the intermediate surface areas of the sealing element 535 remain in the gap distance. The sealing strips 745 come into engagement one after the other during rotation, namely when the piston 506 rotates clockwise, first the sealing strip adjacent to the low-pressure inlet duct 516, and lastly the sealing strip adjacent to the stationary sealing strip 531 in FIG. 29. In this way, a compression chamber can be created that has a very large volume, but can still compress very high. A corresponding design with the arrangement of a plurality of sealing strips in the sealing element 635 is also possible in the embodiment of FIG. 28.

Claims

PATENT CLAIMS:

1. Piston machine as an expansion machine, suction pump or compressor for compressible media, with at least one piston (6, 6 ';406;506; 606), which is in one of two parallel walls (9; 409) and one connecting them, perpendicular to the parallel wall standing peripheral wall (10) formed running space on crank pin (5; 405; 505; 605) at least two identical angularly synchronized crankshafts (2; 402), rotatably mounted perpendicular to the parallel walls, is mounted, whereby the piston on the one end of a chamber surface formed on it in relation to the running direction and the peripheral wall on the other end of a chamber surface formed on it in relation to the running direction each have a sealing element aligned perpendicular to the parallel walls ( 12, 13, 19; 120, 13, 19; 120 ', 13, 19; 300; 412, 413; 531, 535; 631, 635) with which they act on chamber surfaces (11, 14, 18; 110, 14, 18; 110 ', 14, 18; 414, 4 18; 532, 530; 632, 630), which are designed as surfaces which can be moved away from them during the parallel rotation of the sealing elements, and with a high-pressure gas channel (17, 20; 537; 637), which opens into the chamber, and with at least one low-pressure channel (16, 16 ", 22, 22 '; 516, 522; 616), which is outside the chamber opens into the barrel, characterized gekenn¬ characterized in that the chamber surfaces completely as treads (11, 14, 18; 110, 14, 18; 110 ', 14, 18; 414, 418; 532, 530; 632, 630 ) are formed, at the ends of which the sealing elements (12, 13, 19; 120, 13, 19; 120 ', 13, 19; 300; 412, 413;

531, 535; 631, 635) are arranged, the high-pressure gas channel (17, 20; 537; 637) in each case near the end of the tread (11, 18; 110, 18; 110 ', 18; 418; 530; 630) the peripheral wall (10) provided sealing elements (13; 413; 531; 631) opens.

2. Piston machine according to claim 1, characterized gekennzeich¬ net that the sealing elements (12, 120) for sliding engagement have certain surfaces that are circular sector-shaped in cross-section with a center (25; 250) of the circular sector, which is at a distance of the crank radius to the cylinder axis (26; 260) of the counter surface (11, 110), the radius of the counter surface (10, 110) being equal to the crank radius plus the radius of the surface (24) of the sealing element (12, 120).

3. Piston machine according to claim 2, characterized gekenn¬ characterized in that the surfaces of the sealing elements (120, 13, 19; 531, 535; 631, 635) are different with a correspondingly adapted shape of their counter surfaces (110, 14, 18;

532, 530; 632, 630).

4. Piston machine according to one of claims 2 or 3, da¬ characterized in that at least one (12 ', 120') of the piston (6) arranged sealing elements (12 ', 120', 19) with the plane of symmetry of its surface in is at an angle α greater than 0 ° to the plane of symmetry of the running surface (14) adjacent to it, its counter-running surface (11 ', 110') being correspondingly extended.

5. Piston machine according to one of the preceding claims, characterized in that the surfaces of the density elements (300) are provided with an axially parallel toothing.

6. Piston machine according to one of the preceding claims, characterized in that the sealing elements (12, 13, 19) are designed as spring-mounted sealing strips (12, 13, 19; 120).

7. Piston machine according to claim 6, characterized gekennzeich¬ net that the sealing strips with a piston (28) in Zylin¬ der (29, 30) serving grooves are guided, which via channels with the associated tread (11, 18) in connection stand.

8. Piston machine according to one of claims 6 or 7. da¬ characterized in that the bearings of the sealing strips are provided with shock-absorbing devices.

9. Piston machine according to one of claims 6 to 8, characterized in that the sealing strips (12, 13, 19; 120) are widened substantially mushroom-shaped relative to their part (27) guided in the tread.

10. Piston machine according to one of the preceding claims, characterized in that on the circumference (10) of the running space (3, 3 ') two running surfaces (11, 18) with a common central sealing element (13) are provided, which with one on both Ends of the piston (6) provided with sealing elements (12, 19) on the running surface (14) form two adjacent chambers (11.14, 18.14), one of which simultaneously functions as a compression chamber (11.14) and the other as an expansion chamber (11.14) when the piston rotates. 18.14) works.

11. Piston machine according to one of the preceding claims, characterized in that a plurality of chamber arrangements (11.14, 18.14 and 11 '.14', 18 '.14') are provided spaced over the circumference.

12. Piston machine according to one of the preceding claims, characterized in that in parallel running spaces (3, 3 ') on the same crankshafts (2, 2') with cranks (5) and / or running surfaces pistons (6, 6 ') are provided.

13. Piston machine according to one of the preceding claims, characterized in that the valve (23, 23 ') is controlled synchronously with the piston run.

14. Piston machine according to one of the preceding claims, characterized in that the valve of a compression chamber is designed as a one-way valve.

15. Piston machine according to one of the preceding claims, characterized in that the crankshafts (2, 2 ') are coupled by gear sets (7, 7').

16. Piston machine according to one of the preceding claims, characterized in that the piston (6) outside of its tread (14) is provided with a surface design which improves the air contact during its movement.

17. Piston machine according to one of the preceding claims, characterized in that the bearings of the crankshafts (402) with respect to the piston (406) and the stationary Tei¬ len (418) of the running space over the heat transfer obstructing parts (421, 420) connected are.

18. Piston machine according to one of the preceding claims, characterized in that the piston is only supported on one side.