EP0644981B1 - Piston machine - Google Patents

Abstract

PCT No. PCT/EP93/01422 Sec. 371 Date Jan. 12, 1995 Sec. 102(e) Date Jan. 12, 1995 PCT Filed Jun. 4, 1993 PCT Pub. No. WO93/25801 PCT Pub. Date Dec. 23, 1993A piston machine usable as an expansion engine, suction pump or compressor for a compressible medium has a housing containing a working space with two parallel walls and an interconnecting peripheral wall perpendicular to the parallel walls, the peripheral wall having a first chamber surface. A piston has a second chamber surface cooperating with the first chamber surface to form a chamber of variable volume. Two identical crankshafts rotatably mounted in the housing have crank pins rotatable in and supporting the piston, the crankshafts being coupled together for angular synchronous rotation. A first sealing element is mounted on the piston and a second sealing element is mounted on the peripheral wall. Each of the first and second chamber surfaces is formed as a sliding surface to slidingly engage the sealing element of the other of said chamber surfaces during piston rotation. The housing provides a high-pressure gas passage having an opening communicating with the chamber and a low-pressure gas passage communicating with a working space outside of the chamber.

Description

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

An older piston machine is known from US-PS 18 64 699. In this construction, the chamber surfaces delimiting the working chambers on the peripheral wall and on the piston are each composed of cylindrical-section-shaped treads which engage 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 is similar to 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 which is sought because of its use as a steam engine and which has an extremely unfavorable thermodynamic effect for use as an internal combustion engine or compressor. A disadvantage of this design is also the course of the force acting on the piston with 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 particularly 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 force application angle always runs essentially perpendicular to the larger flat surface of the piston. The situation is similar to that of a reciprocating piston in which the force can only be transmitted to the crank over a very small crank angle range with a favorable angle of attack.

A generic piston machine is known from DE-A-1 551 119.

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 on the surfaces with a small chamber volume, that is to say with maximum compression. This also results in an essentially constant force being introduced into the piston, since with increasing compression pressure the loaded piston area becomes smaller and smaller. Overloading the machine even with the highest compression is avoided. Extremely high compression internal combustion engines or compressors can be realized. 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 stands, especially 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 to the advantages of this type of construction, which runs with low vibrations due to the parallel rotation principle and, due to the possibility of compact arrangement of several chambers, enables a high power density and low construction costs. 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.

A disadvantage of this design is the mandatory mounting of the crank pins in the "heads", ie concentrically within the sealing elements of the piston. On the one hand, this extremely limits the design options. Formations other than the four-chamber training shown cannot be constructed or can only be constructed with considerable problems. Furthermore, the position of the crank pins in the "heads" forces the crankshafts to be arranged in the chambers. This leads to considerable sealing problems. The crankshafts must be provided in the parallel walls of the machine housing as eccentric disks that are overrun by the piston. In addition to the considerable sealing problems, there are also problems with the high thermal load on the seals and the bearings, since the chambers generate considerable amounts of heat both when used as expansion chambers of an internal combustion engine and when used as compression chambers, which stress these closely adjacent sealing and bearing points. One of the limitations of the design options is in particular that the sealing surface radii must always be larger than the crank pin radius. This results in significant restrictions regarding the design options for the chamber geometries. Leave certain compression ratios don't reach at all. There are also severe restrictions with regard to the construction of the sealing elements. Thus, as can be seen from the document, these can only be rigidly connected to the piston - namely as a crank pin bearing. Suspended sealing elements, such as are required for better sealing to achieve higher compression, cannot be provided here.

The object of the present invention is to provide a piston machine of the generic type which, with improved design options, in particular enables the provision of more effective sealing elements and in which thermal loads on sealing and bearing points are avoided.

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

According to the invention, the bearing journals are arranged in the piston in a central area in which there is a lot of space available and the bearings of the crank journals can be arranged in a largely free construction, also with regard to the number of crank journals - and thus the crankshafts. The adjacent part of the piston remains free of crankshafts, in particular in the region of the other end of the running surface of the piston. There, the piston can be designed regardless of the need to attach crank caps. The running surface can end here in any way and the piston part delimiting the running surface can be made very narrow here, which benefits the construction of the running space receiving the piston. Of particular advantage is the possibility of providing a sealing element at this end of the tread which can be designed independently of a crank pin bearing to be attached here according to the prior art. Very small radii of the sealing element are therefore possible, and in particular also the design as a sprung sealing strip to increase the gas tightness. Preferably, the piston is free of crankpin bearings at both ends of the tread in the manner according to the invention, so that at both The sealing elements can be freely designed at the ends of the running surface and the piston can be constructed very narrow. The displacement of the crankshafts in relation to the known construction of the generic type from the area of the chambers also results from the displacement of the crank pins in the piston in order to enable their storage in the piston at a point remote from the ends of the running surfaces. This displacement of the crankshafts out of the chambers creates the advantage that the crankshafts are no longer overrun by the piston as in the known construction. Therefore, they do not have to be designed as eccentric discs, but can be constructed in the usual way with crank arms. Sealing problems and thermal problems in the crankshaft bearing and in the crankpin bearing are effectively avoided. In addition, the arrangement of the crank pins and crankshafts according to the invention, as the exemplary embodiments show, creates the possibility of a considerable variety of designs with a largely arbitrary and largely arbitrary number of chambers compared to the generic prior art, which can only be used sensibly for a four-chamber arrangement in single and double arrangement.

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 increased surface area of the sealing elements also leads to less wear.

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 always lies on its mating surface in almost the same line. 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 correspondingly with an enlarged surface of the sealing element, a larger chamber also results. The surfaces of the sealing elements can be Claim 2 be circular in cross section, but other curved surface configurations, in particular elliptical shapes, are also possible. The counter-running 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, various 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 of different shapes. For example, a chamber can be designed such 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. The sealing elements can lie with their plane of symmetry parallel to the plane of symmetry of the adjacent tread, but can also deviate from this angular position. If, as stated in claim 4, they are arranged tilted outwards, this results in an extension of the counter surface and thus a larger maximum chamber volume or a larger compression ratio.

The features of claim 5 are also advantageously provided. The chambers can be sealed by means of sealing elements which are designed as rigid integral parts of the piston or the motor housing. Then the sealing elements can only take into account the manufacturing tolerances at a gap are led to their counter surface, which results in leaks that limit, for example, the maximum compression pressure of a compression chamber. For low pressure compressors this can be enough. 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 6 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 depends on the chamber pressure and is therefore constantly adjusted to the required level.

The features of claim 7 are advantageously provided. In this way, the jumping or rattling of the sealing strips can be better suppressed.

The features of claim 8 are advantageously provided. From the adjacent pressurized chamber, gases under high pressure can 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, which improves their tightness.

The features of claim 9 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, for example, in the possibility of creating a self-compressing internal combustion engine in which one chamber always compresses air, which is external or internal mixture formation can be burned in the other chamber. The separation between the compression chamber and the expansion chamber provided in such an internal combustion engine results in cooling advantages with better partial efficiencies 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 afterburned by this flushing process, so that the pollutant emissions are considerably lower than in other internal combustion engines. The expansion chamber can also be used as a suction pump, eg as a vacuum pump. A piston machine is created 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 10 are advantageously provided. At largely any angle, essentially determined only by the space required, several chamber arrangements can be provided distributed over the circumference of a larger motor, each with a tread provided on the piston and one or two treads provided on the circumferential wall. This can be used, for example, to provide 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. An interesting possibility highlighted only by way of 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 11 are also advantageously provided. In this way, in a manner known, for example, from Wankel engines, the power can be increased in accordance with the number of parallel disks by designing the machine with multiple disks. By angular misalignment of the cranks and / or angular misalignment of the treads in the disks, the degree of uniformity of the machine can be improved, and possibilities can be created to allow compression chambers of one disk to be acted upon directly, without intermediate pressure storage, when the training takes place as an internal combustion engine using the angular offset between the disks.

The features of claim 12 are also advantageously provided. With the generic construction 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 synchronously with the piston run, 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 specifying the desired maximum pressure.

The features of claim 14 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 15 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 16 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.

Fig. 1

im Schnitt in der Achse einer der Kurbelwellen gemäß Linie 1 - 1 in Fig. 2 eine erfindungsgemäße Zweischeibenbrennkraftmaschine,

Fig. 2

einen Schnitt nach Linie 2 - 2 in Fig. 1,

Fig. 3 - 9

Darstellungen gemäß Fig. 2 in aufeinanderfolgenden Winkelstellungen des Kolbens,

Fig. 10 - 15

in Ansicht entsprechend Fig. 2 unterschiedliche Kolbenvarianten verschiedener Ausführungsformen,

Fig. 16

eine zur Verdeutlichung vergrößerte Darstellung des oberen Teiles der Fig. 2,

Fig. 17

eine Darstellung entsprechend Fig. 16 einer Ausführungsvariante mit unterschiedlich großen Dichtleisten,

Fig. 18

eine Darstellung entsprechend Fig. 17 mit schräggestellter Dichtleiste,

Fig. 19

einen Schnitt entsprechend Fig. 2 durch eine Kolbenmaschine mit zwei kleineren und einer sehr großen Kammer,

Fig. 20

im Schnitt gemäß Fig. 2 eine weitere Variante einer Kolbenmaschine mit zylinderförmigen und elliptischen Laufflächen.

The invention is shown schematically, for example, in the drawings. Show it:

Fig. 1

in section on the axis of one of the crankshafts according to line 1 - 1 in FIG. 2 an inventive two-disc internal combustion engine,

Fig. 2

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

3 - 9

2 in successive angular positions of the piston,

Figures 10-15

2 different piston variants of different embodiments,

Fig. 16

2 shows an enlarged representation of the upper part of FIG. 2 for clarification,

Fig. 17

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

Fig. 18

17 with an inclined sealing strip,

Fig. 19

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

Fig. 20

2 shows another variant of a piston machine with cylindrical and elliptical running surfaces.

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 assembly purposes, for example in a disk-like manner. Two parallel, identical crankshafts 2, 2 'are mounted in the housing, of which the crankshaft 2' is visible in FIG. 3. The crankshafts penetrate two running spaces 3, 3 'arranged one behind the other in the manner of disks, of which the running space 3 can be seen open 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. From Fig. 1 it can be seen that the crankshaft 2 at its end lying on the gear set 7 passes through the end wall of the housing 1 and carries there, for example, a drive pulley 8 provided. 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 crankshafts 2, 2 'the piston executes an orbit which, as shown in several successive orbital phases in FIGS. 2 to 9, can be called 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 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 describes a circle during the parallel rotation of the piston 6, as shown in FIGS. 2 to 9, 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 tread 14 in the piston 6, which also has a semi-cylindrical shape with the same radius of the tread 11. Draw imaginary connecting lines through the end points of the tread 11 and through the end points of the tread 14 and compare these connecting lines in the circulation phases in FIGS 9, it can be seen that these connecting lines are always parallel to each other.

By comparing the successive phases of FIGS. 2 to 9, it can also be seen that, in the crank angle position of FIG. 8, the sealing strip 12 of the piston 6 simultaneously engages with the start of the running surface 11 of the peripheral wall on the left in the figures and the sealing strip 13 comes into engagement with the tread 14 formed on the piston. The sealing strips 12, 13 then slide (Fig. 9, 2, 3) on their mating surfaces 11, 14 to to the opposite end of the tread until they run together 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.

Between the running surfaces 11 and 14, therefore, a chamber is formed which is enclosed on all sides 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 a circular provided in the side surfaces of the piston 6 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 circumferential radius of the crank pin 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 expelled through a high-pressure outlet channel 17, the opening of which in the parallel wall is shown in FIGS. 2 to 9, at a greatly increased pressure.

As shown in FIGS. 2 to 9, a further running surface 18 is arranged laterally in addition to the previously described running surface 11 in the peripheral wall 10, which is designed mirror-symmetrically to the sealing strip 13 and identical to the running surface 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 in engagement with the tread 18 runs 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 opens (FIG. 5) and to close again at Fig. 8.

A high-pressure inlet channel 20 also opens into the chamber 18.14, which, in contrast to the high-pressure outlet channel 17, is not provided for the outlet of compressed gas, but rather 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 obviously works according to the single-stroke principle, since it requires only 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 duct 20 to the chamber 18.14 at a point in time of small chamber volume or through the high-pressure inlet duct 20 'to the chamber 18'.14' and brought to an explosion there. 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.

In a departure from the arrangement shown in FIG. 2, the expansion chamber 18.14 and the compression chamber 11'.14 '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 work on the diesel principle. An injection nozzle is then to be provided 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 ducts 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 region of the minimum chamber volume.

The low-pressure channels 16 and 22, which serve for the inlet and outlet, are 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 flushing from the low-pressure inlet channel 16 to the low-pressure outlet channel 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 synchronization 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 flap 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 crank pins 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 crank pin 5 to give the radius of the tread 11, based on its 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 surfaces, So sealing strip 12 on counter surface 11, sealing strip 13 on counter surface 14 and sealing strip 19 on counter surface 18, you can see in the sequence of FIGS. 2 to 9 that the sealing strips run with constantly changing contact line on the counter surfaces, each of which is an envelope curve of the circulation result in a sealing strip when the piston 6 rotates in parallel.

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, displaceably mounted with a slide 27 in a sliding guide and form a piston 28 at their end opposite the surface 14 that 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 by a bore, not shown, in order to be acted upon by the latter with high pressure gas, which presses the sealing strip against its running surface with additional prestress 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 a symmetrical position 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 shown in FIG. 2, reversed with respect to the treads 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 in the clockwise direction, while the chamber 18'.14 'is an expansion chamber is. Correspondingly, the length of the inlet and outlet low pressure ducts 16 ', 22' and the high pressure ducts 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 disc, 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 one disk to release high-pressure gas at a point in time when the expansion chambers of the other disk 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 a 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 that has to be driven externally and that has only one compression chamber per disc. In a corresponding embodiment, as shown in FIG. 2, such a pure compressor per disk can also have two compression chambers (but no expansion chambers).

In another embodiment, for example, only expansion chambers can be provided in one pane and only compression chambers in another pane. 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 rotational movement of the piston enables a largely arbitrary number of chamber arrangements per piston.

In FIG. 10, the center of the running surface is marked in the running surface 14 of the piston shown as an intersection of the tangent T created at this point and the perpendicular S to it. The two lines T and S form the four quadrants labeled with the Roman numerals I, II, III and IV. As shown in FIG. 10, the two crank pins 5 in the piston are located on the other side of the line T, ie in the central part of the piston, from the chamber which can be formed by the running surface 14. At both ends of the tread 14 As shown, the piston can be made very narrow and with spring-loaded sealing strips with small sealing strip radii. In the construction of the piston in the region of the two ends of the running surface 14, it is not necessary to take account of any crank pin bearings to be arranged here.

The design variants of FIGS. 11 to 15 clearly show the same geometric relationships for each of the running surfaces 14 present in these with respect to the arrangement of the bearings of the crank pins 5. These geometric relationships also apply to the other piston constructions shown in FIGS. 1 to 18.

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 act at a different angle than in the embodiment of FIGS. 2 to 9 and 11. 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 generated 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 are. The same parts are provided with the same reference numerals.

In a departure 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 of 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. This results in a shape of the new tread 110, which is widened to the left and upwards, 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 tread 14 and the tread 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 differently from the sealing strips 13 and 19, for example with other dimensions even bigger or a little 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 tread 14 of the piston would remain unchanged. Both stationary treads 11 and 18 would have to be changed accordingly.

Fig. 18 shows a variant of the construction of Fig. 17 in the same representation. Matching parts are provided with the same reference numerals. The reference numbers of modified parts have also been 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, for comparison, the enlarged sealing strip 120 of FIG. 17 is viewed again, it can be seen that, like the sealing strip 19 located at the right end of the running surface 14 of the piston, its plane of symmetry lies exactly parallel to the plane of symmetry of the adjacent running surface 14. As shown in FIG. 17, this results in a maximum circumferential angle of the counter surface 110 of the chamber of approximately 180 °. Even 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.

In Fig. 18 it is shown that there the enlarged sealing strip 120 'with its dash-dotted plane of symmetry under one Oblique angle α is arranged opposite the plane of symmetry of the tread 14 adjacent to it, 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 'into an abutment on the correspondingly lengthened counter-running surface 110' over an angular range likewise increased by a beyond 180 °. 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 chamber 110'.14 can again be greatly enlarged, while the right chamber 18.14 is kept small, since the sealing strip 19 is arranged at 90 ° and also has a much smaller surface area of its 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 enlargement of the chamber.

The tightness of a sealing strip can be improved, for example in the sealing strip 120 of FIG. 17, if the lifting is prevented. Lifting usually occurs when the sealing strip vibrates or rattles in its resilient mounting in the event of smooth running malfunctions. Such vibration movements can be prevented by shock absorption. For this purpose, shock-absorbing devices can be provided in the resilient bearing seat of a sealing strip, for. B. hydraulic damping devices in the manner of conventional hydraulic piston shock absorbers.

In piston machines with thermally highly loaded chambers, heat can be dissipated, for example by Water cooling channels in the housing near the treads provided there and by liquid cooling of the piston, which can be done, for example, with oil channels in the piston, which are connected to the two crankshafts via the bearings. However, air cooling in the housing is also possible, for example by means of external ribbing. The piston can also be adequately 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) an opening that 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 tread 14 and the bearings on the crank pin 5 passes and which is flowed through by air when the piston rotates.

1-18, the sealing elements that delimit a chamber are always shown to be significantly 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. 19.

FIG. 19 shows a cross-section of the 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 °. It connects 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 runs on the running surface 530 when the piston 506 rotates in parallel. and that with parallel rotation of the piston 506 clockwise between the start of the tread 530 at the low pressure inlet channel 516 to the end of the tread 530 at the 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 that is very 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. 19, the chamber 530.532 is combined with the two chambers 110.14 and 18.14 of the embodiment of 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. 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 construction of the gas routing 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 forms an internal combustion engine in the manner described above, which drives the compressor. The low-pressure compression chamber 530.532 can also serve as a pre-compression chamber, the gas pre-compressed in it 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.

19 shows a further construction detail, which is of great advantage when used as an internal combustion engine 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. 19, 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 on one Nose 541 of the piston 506 a sealing element 542, which runs sealingly on the tread 540 during the critical crank angle range in which the expansion chamber 18.14 opens and creates a gas seal between the low-pressure channels 522 and 516, so that exhaust gas is mixed in this critical time range and fresh gas is avoided.

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

However, other surface shapes are also possible, such as conic sections in particular, e.g. Sections of circles, ellipses and parabolas, but also spiral sections. The mating surfaces that are covered by such sealing element arms 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, with a circular surface of the sealing element, circular running surfaces result in cross section. The elliptical surface of the sealing element results in an elliptical running surface. Such an example with elliptical surfaces is explained in FIG. 20.

20 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 crank pins 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-sectional elliptical surface, which extends from the marking line 633 to the marking line 634 enough. It connects to a circular cross-sectional tread 632 at 633. A sealing strip 631 with a small circular cross section is arranged on the peripheral wall of the housing 601 and runs on the running surface 632 of the piston. 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 clockwise in parallel, the sealing element 635, which is designed as an elliptical section, runs from the angular position shown in FIG. 20, 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 and operates alternately with the chamber 630.632 described first when the piston 606 rotates.

In the embodiments of FIGS. 19 and 20, 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, respectively. These sealing elements 535, 635 are therefore surfaces which are rigidly connected to the piston trained and can only seal with a gap seal against their counter surface.

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, although technically possible.

The geometric arrangement of the crank pins in the piston is also explained in FIGS. 19 and 20, as shown in FIG. 10.

19 shows the position of the quadrants I-IV with the auxiliary lines T and S, which intersect at right angles in the middle of the tread, on the two running surfaces 14 and 532 provided on the piston 506. It can be seen that in the case of the tread 14, the crank pins 505 of the piston lie outside the quadrants I and IV. In the case of the tread 532, this only applies to the quadrant I. No crank pin is mounted on this end of the tread 532 opposite the sealing element 532. It can therefore be constructed freely; in the exemplary embodiment shown, the running surface 532 merges at a right angle into an adjacent surface. Crank pins are not provided in this area.

20 shows the quadrant arrangement for the tread 632 ', which also applies in a corresponding manner to the tread 632, which is arranged completely symmetrically. It can be seen that a crank pin 605 can be arranged in quadrant IV, but not in quadrant I, in which the tread 632 'is delimited by the piston 606 in a very narrow configuration, which benefits the closely matched design of the piston chamber surrounding the piston .

In all of the illustrated embodiments of FIGS. 1 to 20 it can be seen that crankshafts are always outside of the chambers formed when the sealing elements engage the treads. In the case of Fig. 20, the tread 632 'is not in chamber engagement. However, the tread 632 is in engagement with the maximum chamber size with the tread 630. All crankshafts, which in this illustration are in the center of the illustrated circles of the crank pins 605, are outside the chamber 630.632 and thus protected against heat attack and against attack by compressed gases.

Claims (16)

1. A piston machine as an expansion machine, suction pump or compressor for compressible mediums, having at least one piston (6, 6'; 506, 606) mounted, in a running space formed by two parallel walls (9) and a circumferential wall (10) connecting and being perpendicular to said parallel walls, on crankshaft journals (5; 505; 605) of at least two identical crankshafts (2) which are rotatably mounted perpendicular to the parallel walls and which are coupled with angular synchronisity, said piston, on the one end with respect to the running direction, of the chamber surface formed on said piston, and said circumferential wall, on the other end with respect to the running direction, of a chamber surface formed on said circumferential wall, each comprising a sealing element (12, 13, 19; 120, 13, 19; 120', 13, 19; 531, 535; 631, 635) directed perpendicularly to the parallel walls with which said piston and circumferential wall, with an engagement forming a chamber, slide on one another on running surfaces (11, 14, 18; 110, 14, 18; 110', 14, 18; 532, 530; 632, 630) which are formed as surfaces being followed by the sealing elements in their parallel rotation, and having a high-pressure gas channel (17, 20; 537; 637) provided with a valve and opening into the chamber, and having at least one low-pressure channel (16, 16', 22, 22'; 516, 522; 616) which opens into the running space outside of the chamber, said chamber surfaces being completely formed as running surfaces (11, 14, 18; 110, 14, 18; 110, 14, 18; 532, 530; 632, 630) on one end of which the sealing elements (12, 13, 19; 120, 13, 19; 120', 13, 19; 531, 535; 631, 635) are arranged, said highpressure gas channel (17, 20; 537; 637) in each case opening near the sealing elements (13; 531; 631) provided at the end of the running surface (11, 18; 110, 18; 110', 18; 530; 630) of the circumferential wall (10), characterised in that the crankshaft journals (5) lie in a central region of the piston (6, 506, 606) and, with quadrants being defined by the tangant (T) incident with the centre point of the running surface (14, 532, 632) of the piston (6, 506, 606) and the perpendicular (S) thereto in the centre point, lie outside at least of that quadrant of the quadrants (I, IV) lying outside the tangent (T) which (1) contains the other end, with respect to the direction of travel, of the running surface (14, 532, 632), and that the crankshafts (2, 2') are arranged outside of the chambers (11.14, 18.14, 11'.14, 18.14', 530.532, 630.632) formed when the sealing elements are engaged with the running surfaces.
2. A piston machine according to claim 1, characterised in that the sealing elements (12, 120) comprise surfaces provided for the sliding engagement, which are formed circular-sector-shaped in cross-section with a centre point (25; 250) of the sector which lies in the distance of the crank radius from to the cylinder axis (26; 260) of the opposite running surface (11, 110), said radius of the opposite running surface (10, 110) amounting to the crank radius plus the radius of the surface (24) of the sealing element (12, 120).
3. A piston machine according to claim 2, characterised in that the surfaces of the sealing elements (120, 13, 19; 531; 535; 631, 635) are different with a correspondingly adapted shape of their opposite running surfaces (110, 14, 18; 532, 530; 632, 630).
4. A piston machine according to one of claims 2 or 3, characterised in that at least one (12', 120') of the sealing elements (12', 120', 19) arranged on the piston (6), with the plane of symmetry of its surface, lies at an angle a which is greater than 0°, to the plane of symmetry of its neighbouring running surface (14), its opposite running surface (11', 110') being correspondingly lengthened.
5. A piston machine according to one of the preceding claims, characterised in that the sealing elements (12, 13, 19) are formed as spring mounted sealing strips (12, 13, 19; 120).
6. A piston machine according to claim 5, characterised in that the sealing strips are guided with a piston (28) in grooves serving as cylinders (29, 30) and being connected to the corresponding running surface (11, 18) via channels.
7. A piston machine according to one of claims 5 or 6, characterised in that the mountings of the sealing strips are provided with shock absorbing devices.
8. A piston machine according to one of claims 5 to 7, characterised in that the sealing strips (12, 13, 19; 120) are broadened, essentially mushroom-shaped, with respect to their part (27) which is guided in the running surface.
9. A piston maclline according to one of the preceding claims, characterised in that there are two running surfaces (11, 18) with a common middle sealing element (13) provided on the circumference (10) of the running space (3, 3'), said running surfaces together with a running surface (14), provided with a sealing element at both ends, of the piston (6), forming two neighbouring chambers (11.14, 18.14), of which, on circulation of the piston, simultaneously, one chamber functions as a compression chamber (11.14) and the other as an expansion chamber (18.14).
10. A piston machine according to one of the preceding claims, characterised in that several chamber arrangements (11.14, 18.14, 11'.14',18'.14',530.532, 630.632) are arranged spaced about the circumference.
11. A piston machine according to one of the preceding claims, characterised in that in parallel running spaces (3, 3') on the same crankshafts (2, 2') pistons (6, 6') are provided with, where appropriate, angularly displaced journals (5) and/or running surfaces.
12. A piston machine according to one of the preceding claims, characterised in that the valve (23, 23') is controlled synchronously with the circulation of the piston.
13. A piston machine according to one of the preceding claims, characterised in that the valve of a compression chamber is designed as a one-way valve.
14. A piston machine according to one of the preceding claims, characterised in that the crankshafts (2, 2') are coupled by sets (7, 7') of gears.
15. A piston machine according to one of the preceding claims, characterised in that the piston (6), on the outside of its running surface (14), is provided with a surface formation which improves the air contact on the motion of the piston.
16. A piston machine according to one of the preceding claims, characterised in that the piston is mounted on bearings on one side only.

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