EP0204136A1 - High pressure radial piston maschine, comprising means for securing the piston shoe - Google Patents

Abstract

The invention uses piston shoes (502, 119) or conical ring elements (1, 2, 11, 12, 307) in pump or motor engines (7, 48, 340). Formed within the conical part (1, 2, 11, 12, 307) of the element is a working chamber (14, 61, 311) through which fluid flows and to which are assigned inlet means (50) and outlet means (49) which, under the control of a driving arrangement (52, 55, 96, 97) acting via reciprocating pistons (94), periodically enlarges and reduces its volume. It is thereby possible to achieve the change in the volume of a working chamber without sealing pistons moved in cylinders, to avoid any sealing by sealing parts sliding against one another without lubrication and thereby to make it possible for the working chamber to be filled and operate under high pressure with a non-lubricating fluid, for example water. The driving arrangement is separated from the working chamber by means of the reciprocating piston and can be operated using a fluid having a lubricating effect. The details can be designed in such a way that pressures of over 1000 bar can be used efficiently with a non-lubricating fluid in the working chamber. Corresponding piston shoes (52, 21, 502) capable of operating under high pressure can be used in the driving arrangement. <IMAGE>

Description

The invention relates to hydrostatic or pneumatic units in which fluid flows through an increase periodically = en and contracting working chamber, such as pumps, Kompre ₌ ssoren, fluid motors, internal combustion engines or transmissions. In such units, fluid is pumped or a rotor is driven by fluid in the motor.

Pumps and motors are used as hydraulic pumps and hydraulic motors in many industries. The simplest pumps are plastically deformable diaphragm pumps for low pressures, where the diaphragm forms a conical ring element in places. Pumps for very high pressures are used, for example, in water jet cutters and often have to deliver high water pressure of a thousand or several thousand bars.

Pumps with membranes, which can also be flexible metals, are known, for example, from US Pat. No. 3,861,277. Pumps for very high pressures, e.g. for water jet (water jet) cutting systems are known from the catalogs of the specialist companies and consist of a hydraulic piston with a large cross-section for driving the water pump piston with a small cross-section or they are mechanically driven piston pumps with several Pistons side by side and the axes of the pistons mostly in a common plane.

The membrane pumps mentioned correspond to the preamble of patent claim 1. However, the diaphragm pumps have so far not been used as high-pressure water pumps for water-jet cutting systems, because they only let low pressures of a few, or in the best pressure-resistant versions up to a maximum of 100 bar. Until now there has been a lack of possibilities to make the membranes usable for pressures of more than 100 bar in the cavity within the conical ring element. Obviously, it was considered that too for so _u n ₌ possible that no efforts to diaphragm pumps for 1,000 or more bar were known.

With the piston pumps described, your high pressures of 1000 or menr boron pressure, it is necessary to seal the piston periodically enlarging and reducing the piston against leakage. The high pressure of the non-lubricating water acts on the plastic seal. If there is no lubrication, the plastic seal quickly wears out under the high water pressure that aggregates of this type decrease in tightness and delivery rate after just a few hundred hours of operation. In addition, at such high pressures, the plastic seals create a high level of friction that reduces the efficiency of the system. After all, these systems are expensive. Even small ones of 45 Kw cost around a hundred thousand marks, so that despite the clean cutting of the water jet systems, only a few hundred such systems have been used in the world so far. For the streaming of the cutting points of stone drills, these systems turn out to be too expensive and, moreover, the non-uniformity of the funding causes disturbing fluctuations at the cutting points of the stone drills, milling cutters and saws.

The invention effectively remedies this. This is because, according to patent claim 1, it creates a pressure fluid chamber, which periodically reduces and enlarges its volume through which fluid flows, for high pressures in the fluid of more than 500 bar, which, in certain embodiments according to the invention, can also operate reliably for several thousand bar. This is achieved according to the characterizing part of patent claim 1 in that the pressure fluid chamber is assigned pistons and piston shoes and / or conical ring elements with centering and supports.

According to claims 2 to 10, the units are designed so that sealants can be used that seal perfectly against high pressure from the pressure fluid chamber without being exposed to movement with wear on a part. The previous abrasion of the plastic sealing rings is thereby eliminated, as far as the seal according to the invention is free at all times with freedom of movement and sliding under the pressure of a sealing lip against a surface.

The invention, as characterized in in the claims, also solves the task of replacing the seals which seal under sliding against quick-lubricating fluid and quickly rub through non-wearing seals which do not seal under sliding and which are known from the preamble of claim 1 To make membranes or elements with conical parts for very high pressures reliable.

This increases the service life of high-pressure units, in particular those with non-lubricating fluid in the working chambers, and at the same time increases the efficiency of the units, because the invention eliminates the ramifications of the previous seals, with the unit being equal at the same time Receive funding in the fluid flow and the price of the unit can be cheaper.

The task of the invention for high pressure k can only be achieved by following the precise rules aer invention. Otherwise the elements will break under too little internal tension. In addition, a corresponding control device suitable for the high pressures must be arranged for controlling the stroke of the elements. This control is stitched in the drive application patent application, which word is to include that when used in the engine it is driven by the pressure effect of the fluid in the working chamber, ie does not drive itself. The drive arrangement must be able to overcome or absorb the high forces acting on the elements and it is therefore expedient to use particularly high-pressure piston shoes in the unit in order to move the support periodically, or that Reciprocating piston, which controls the movement of the conical parts of the elements, must be periodically driven accordingly, or its driving forces must be absorbed and utilized in the engine.

The thus for the elements of _du = OF INVENTION ng developed piston shoes, piston pressing piston and blowing agents or auxiliary parts turn out to be also useful in general hydraulic pumps, hydraulic motors, compressors or gears, respectively. also in internal combustion engines, without the conical elements also having to be arranged in them. Further parts of the task and solutions of the invention, such as the training required to implement them, are explained in more detail on the basis of the description of the preferred exemplary embodiments, and a geometrical-mathematical examination is also provided, without which the elements break quickly under high internal pressures in the working chamber would.

• Fig. 1 einen Laengsschnitt durch einen Elementenpaarsatz;
• Fig. 2 einen Laengsschnitt durch einen Teil eines Aggregates
• Fig. 3 einen Laengs schnitt durch einen Teil eines Aggregates
• Fig. 4 einen Laengs schnitt durch einen Teil eines Aggregates
• Fig. 5 einen Laengsschnitt durch einen Teil eines Elementes
• Fig. 6 einen Laengsschnitt durch einen Teil eines Elementes.
• Fig. 7 einen Laengsschnitt durch einen Teil eines Aggregates;
• Fig. 8 einen Querschnitt durch Teile eines Elementes ;
• Fig. 9 einen Laengsschnitt durch einen Teil eines Elementes
• Fig. 10 einen Laengsschnitt durch einen Teil eines Elementes;
• Fig. 11 einen Querschnitt durch einen Teil eines Aggregates;
• Fig. 12 einen Laengsschnitt durch einen Teil eines Aggregates;
• Fig. 13 eine schematische Darstellung mit Mathematik;
• Fig. 14 eine geometrische und mathematische Daten und Formeln;
• Fig. 15 die inneren Spannungen in einem Element
• Fig. 16 eine Schematik einer Festbrennstoff Anordnung;
• Fig. 17 einen Daengsschnitt durch eine Haelfte eines Elementes;
• Fig. 18 einen Laengsschnitt durch einen Teil eines Elementes;
• Fig. 19 einen Querschnitt durch einen Teil eines Aggregates;
• Fig. 19-A die Figur 19, jedoch mit den Formeln fuer die Berechnung;
• Fig.20 die Spannungen in einem Element
• Fig. 21 einen Querschnitt durch eine Anordnung;und;
• Fig. 22 einen entsprechende Laengsschnitt dazu;
• Fig.23 einen Laengsschnitt durch eine Anordnung;
• Fig. 24 einen Laengsschnitt durch ein Aggregat;
• Fig. 25 eine Abwicklung eines Teiles der Figur 24;
• Fig.26 einen Radialschnitt durch einen Teil der Figur ₂4 und eine auf einen Steuerflaechenteil der Figur 24;
• Fig. 27 einen Laengsschnitt durch Teile eines Aggregates;
• Fig. 28 einen Querschnitt durch Figur 27 entlang dem Pfeil II-II;
• Fig. 29 einen Querschnitt durch Figur 27 entlang dem Pfeil IV-IV;
• Fig. 30 eine Draufsicht auf Figur 27 von oben ;
• Fig. 31 einen Laengsschnitt durch Teile eines Aggregates;
• Fig. 32 einen Schnitt durch Figur 31 entlang der gepfeilten Linie;
• Fig.33 einen Schnitt durch einen Teil der Figur 31 ;
• Fig.34 einen Querschnitt durch Figur 33 entlang dessen Pfeillinie;
• Fig. 35 einen Laenegsschnitt durch einen Kolbenschuh der Fig. 31 :
• Fig.36 einen Querschnitt durch die Mitte der Figur 35;
• Fig. 37 Laengsschnitte durch Teile der Figur 31;
• Fig.38 einen Querschnitt durch Figur 37 entlang der gepfeilten Linie in Fig. 32.
• Fig.39 einen Laengssconitt durch einen Teil eines Aggregates;
• Fig.40 Querschnitte durch die Figur 39 entlang der in der Figur 39 gezeicnneten gepfeilten Linien;
• Fig.41 einen Laengsscnnitt durch einen Kolbenschuhe mit einem Teile des Kolbens;
• Fig. 42 einen Querschnitt durch di e Mitte der Figur 41 ;
• Fig. 43 eine Draufsicht auf Figur 41 von oben mit einer Defi = nierung der Abmessungen der Teile darin;
• Fig. 44 einen Schnitt durch den Laufteil eines Kolbenschuhes parallel zur Koloenhubfuehrungsflaeche geschnitten;
• Fig. 45 eine Draufsicht cuf einen Kolbenschuh von oben;
• Fig. 46 einen Querschnict durch Figur 45 entlang der in der Figur 45 gezeichneten gepfeilten Linie;
• Fig.47 einen Laengsscnnitt durch ein Aggregat;
• Fig. 48 einen Laengsscrnitt durch einen Teil eines Aggregates;
• Fig. 49 einen Querschnitt durch Figur 48 entlang dem Pfeil A-A;
• FIg.50-A einen Schnitt du-ch Figur 48 entlang der Pfeillinie B-B;
• Fig.50-B eine Ansicht des Kolbenschuhes der Figur 48 von rechts;
• Fig.50-C einen Schnitt durch den Kolbenschuh der Figur 48 entlangA-A;
• Fig.51-A einen Laengsschnitt durch einen Kolben und Kolbenschuh;
• Fig.51-B einen Querschnitt durch Figur 51-A entlang der pfeitinie darin;
• Fig.51-C eine Alternative zur Figur 51-B;
• Fig.52 eine Alternative zur Figur 51-B;
• Fig.53-A teilweise Lnengsschnitte durch einen Kolben und Kolbenschuh; und
• Fig.53-B einen Querschnitt durch Figur 53-A entlang der Pfeillinie darin;

The figures show:

• 1 shows a longitudinal section through a pair of elements.
• Fig. 2 shows a longitudinal section through part of an assembly
• Fig. 3 shows a Laengs section through part of an aggregate
• Fig. 4 shows a Laengs section through part of an aggregate
• Fig. 5 shows a longitudinal section through part of an element
• Fig. 6 shows a longitudinal section through part of an element.
• 7 shows a longitudinal section through part of an assembly;
• 8 shows a cross section through parts of an element;
• Fig. 9 shows a longitudinal section through part of an element
• 10 shows a longitudinal section through part of an element;
• 11 shows a cross section through part of an assembly;
• 12 shows a longitudinal section through part of an assembly;
• 13 shows a schematic illustration with mathematics;
• 14 shows geometric and mathematical data and formulas;
• 15 shows the internal stresses in an element
• 16 shows a schematic of a solid fuel arrangement;
• 17 shows a narrow section through half of an element;
• 18 shows a longitudinal section through part of an element;
• 19 shows a cross section through part of an assembly;
• 19-A shows FIG. 19, but with the formulas for the calculation;
• Fig.20 the stresses in an element
• 21 shows a cross section through an arrangement;
• 22 shows a corresponding longitudinal section;
• 23 shows a longitudinal section through an arrangement;
• 24 shows a longitudinal section through an aggregate;
• 25 shows a development of a part of FIG. 24;
• 26 shows a radial section through a part of FIG. ₂ and a part of a control surface of FIG. 24;
• 27 shows a longitudinal section through parts of an assembly;
• 28 shows a cross section through FIG. 27 along the arrow II-II;
• 29 shows a cross section through FIG. 27 along the arrow IV-IV;
• 30 shows a top view of FIG. 27 from above;
• 31 shows a longitudinal section through parts of an assembly;
• 32 shows a section through FIG. 31 along the arrowed line;
• 33 shows a section through part of FIG. 31;
• 34 shows a cross section through FIG. 33 along the arrow line;
• 35 shows a Laeneg section through a piston shoe of FIG. 31:
• 36 shows a cross section through the center of FIG. 35;
• 37 Longitudinal sections through parts of FIG. 31;
• 38 shows a cross section through FIG. 37 along the arrowed line in FIG. 32.
• 39 shows a longitudinal section through part of an aggregate;
• 40 cross sections through FIG. 39 along the arrowed lines drawn in FIG. 39;
• 41 shows a longitudinal section through a piston shoes with a part of the piston;
• 42 shows a cross section through the center of FIG. 41;
• 43 shows a top view of FIG. 41 from above with a definition of the dimensions of the parts therein;
• 44 shows a section through the running part of a piston shoe cut parallel to the Koloenhubfuehrungsflaeche;
• 45 is a top view of a piston shoe from above;
• 46 shows a cross section through FIG. 45 along the arrowed line drawn in FIG. 45;
• 47 shows a longitudinal section through an aggregate;
• 48 shows a longitudinal section through part of an aggregate;
• 49 shows a cross section through FIG. 48 along the arrow AA;
• FIg.50-A a section du-ch Figure 48 along the arrow line BB;
• 50-B a view of the piston shoe of FIG. 48 from the right;
• Fig. 50-C shows a section through the piston shoe of Fig. 48 along A-A;
• Fig. 51-A a longitudinal section through a piston and piston shoe;
• Fig. 51-B shows a cross section through FIG. 51-A along the line of arrow therein;
• Fig. 51-C an alternative to Figure 51-B;
• Figure 52 shows an alternative to Figure 51-B;
• Fig. 53-A partial longitudinal sections through a piston and piston shoe; and
• 53-B show a cross section through FIG. 53-A along the arrow line therein;

As far as elements or aggregates are mentioned in the previous brief description of the figures, this means that they are elements or aggregates of the invention.

In FIG. 1, the ring element pairs consist of fiber materials, for example carbon fiber, the pairs being glued to one another, for example by means of epoxy resin. The elements 1 and 2 are flattened at their outer ends for the purpose of connection with epoxy resin, that is to say expanded radially so that they are glued together in the connection area, adhesive area 23 and form the pump chamber 61 between. The radially inner ends of the elements 1 and 2 are also flattened, that is to say provided with planar parts which still extend radially inwards, the upper pair 1, 2 being connected to the lower pair 1, 2 by the connecting surface, adhesive = surface 24. This execution of a set of the conical elements of the invention is particularly handy, easy to manufacture and cheap. The handiness arises particularly from the fact that one can glue a number of such element pairs directly to an element column by means of the inner bores or recesses 35, so that the column is then a finished installation piece with large pump or motor chambers 61 and large pump or Motor - stroke allowed. The fiber is very durable. The durability of the adhesive seams 23, 24 is determined by the radial expansion of the flat ring parts, because the adhesive is parallel to the dimensions of the adhesive surface, for example with about 8 kg per square meter, while the fibers of the elements are up to 300 kg per have square millimeter strength. There are still such fibers in development in Japan that will still achieve twice the durability. The conical shape results from shapes in conical shapes in which the fibrous materials are inserted and coated with epoxy resin or other connecting materials and then dried.

Figure 2 shows part of a high pressure pump for non-lubricating or rust-causing materials, for example for water. In the pump body 48, the pump cylinder for the non-forging medium, which is called water below, is located in each of the figures. Pump head 48 contains the inlet - valve 50 and the exhaust valve 49. The valves are often utilized to federbe = Bei iel with spring 51. In the cylinder 61, the piston reciprocates 58. Since these non-dope _h the medium contacts and corrode the running surface in this water and would not lubricate, the piston 58 is provided with the seal 62 near the cylinder 61. Piston 58 is held by means of a connecting bolt 70 with a head 64 and a base 71 at a piston piston 59, on the rear end of which the pump piston 58 rests and is sealed by means of a seal 72. The pin or the holder 71 connects the two pistons 58 and 59 to one another via the bolt 70 and holds them together in a force-locking manner. As a result, the pump piston 58 is riziprokiert together with the drive piston 59, the bolt 70 the lower tensile force or compressive force during the suction stroke and the support of the outer axial ends of the two pistons 58 and 59 on each other the greater force during the pressure stroke from the driving calf to the sealing piston or pump - Piston 58 transferred. The connecting bolt 70 is also sealed by means of a plastic seal 72 against at least one of the pistons 58 or 59. The drive to the pressure stroke takes place through the circumferential eccentric ring 55. The pivotable piston shoe 52 with its running surface 57 runs on its outer surface 56. The piston shoe 52 is inserted between the driving piston 59 and the eccentric ring 55 and is pivotable on the piston 59 stored. From channel 68, lubricating fluid, for example oil, is passed into chamber 67, which surrounds the outer axial end parts of the two pistons 58, 58, so that these piston ends dip into and reciprocate in them, from where the lubricating fluid through channels 74 into the sliding surfaces between the piston shoe 52 and the eccentric ring 55, as well as in the swivel surfaces between the piston shoe 52 and the treble piston 59. It also enters corresponding pressure fluid chambers 73, if such are arranged. The drive piston is perfectly lubricated and can run well. From the chamber 67, however, the lubricating fluid also enters the fitting gap between the pump body 48 and the sealing piston and pump piston 58. As a result, this piston is also permanently lubricated over the entire length, at least when the seal 62 is arranged near the chamber 61, that is to say at the inner end of the piston 58. So it is in the of Figure 2 drawn. FIG. 2 thus provides good lubrication of the running surface of a piston in a cylinder even when the piston pumps water or a non-lubricating medium. However, since the seal 62 occasionally wears or leaks with such media, in particular if high pressure is pumped from, for example, 1000 to 4000 bar, the drainage groove 66 with the mixed fluid collection chamber 65 is also arranged according to the invention. The chamber 65 surrounds a part of the middle part of the pump piston 58. If dirty or non-lubricating fluid escapes from the cylinder 61, it flows into the collecting chamber 65 and flows from there through the line 66, so that the space 67 remains filled with clear lubricating fluid . Because in chamber 67 the water cannot enter because the chamber 67 is filled with a higher pressure than the outlet channel 66. The inventive example of FIG. 12 therefore creates lubrication of the pump piston by means of the arrangement of a second piston and a lubrication chamber In addition, the example can create a drain device for the drainage of leakage fluid, whereby mixed fluid is drained and the two fluids water and oil are kept clean in their respective chambers.

FIG. 3 again shows a pump or motor unit with conical elements, pump head 48, valves 49 and 50, and a drive 55 and a piston or shoe 52. The special feature of this exemplary embodiment of the invention is that here the centering rings with support bodies 41 are united and, on the other hand, the inner rings are joined together with the centering cylinders projecting from both ends to form outer rings 43, 44. The middle centering pieces are designated 45, 46 and they are located on the support body 41, while the outer centerings are designated 43.44 and are located on the filler bodies 42. Bores 13 for connecting the chambers 61 are arranged again and the elements 1 and 2 are each nested between two parts 41 and 42, or they touch one of these parts, the centering seats and surfaces 47, 53 in the head 48 or in the piston or shoe 52 The other item numbers show parts known from FIG.

The example of the invention in FIG. 4 shows the pump body 48 with the pump cylinder 61 in which the pump piston 60 reciprocates. Inlet and outlet valve 49.50 are arranged again. The pump piston is here provided with a foot 86 with centering 85, on which the pressure spring 84 is mounted, which pushes the piston 60 out of the cylinder 61 during the suction stroke pulls out. It itself is advantageously mounted on a recess surrounding the cylinder 61 or in an annular groove 83, because this spring must be long in order to deliver a large piston stroke with a long service life. The outer end of the piston foot is radially flat to form a bearing surface for the driving piston. The driving piston 59, which drives the pressure stroke of the pump piston 60, is reciprocally mounted in a guide 148 and has a head with the pivot surface 76 with which the piston head is pivotably supported in the pivot bed with the pivot surface 75 of the piston shoe 52. The running surface 57 of the piston shoe hes 52 runs on the running surface 56 of the eccentric ring 55, as a result of which the driving piston 59 receives the pressure stroke via the piston shoe 52 and passes this over the bearing surface of the piston foot 86 onto the pump piston 60. The piston shoe has guide parts 81 with guides 79, in which it is guided in the recesses 80 of the guide 148 in order to prevent the piston shoe 52 from falling away from the piston 59 and nevertheless to enable a long piston stroke. The space between the pump body 48 and the guide body 148 is advantageously = again filled with lubricating fluid in order to allow good lubricated running of the two pistons 60 and 59 at least for a certain period of life.

FIG. 5 shows how one can advantageously insert or vulcanize a seal 38, for example an O-ring, into a groove 37 of the radially outer part of an element 1 in order to ensure a good seal of the element 1 on the piston, pump head or the described part = get ring.

FIG. 6 shows the corresponding arrangement or vulcanization of a plastic seal (ring) on the radially inner part or end of the element 1.

FIG. 7 shows in a longitudinal section part of a Ratew pump arrangement for very high water pressures. This arrangement works in the supercritical, high pressure range. In the pump head 48, the connection 122 leads to the inlet valve 50 and the outlet channel 123 to the outlet valve 49 with spring 51. The inlet valve advantageously has spring 88 nested in piston 40, for example in its annular groove 88. A narrow ring groove, therefore, so that internal compression = losses are avoided. Between the piston 40 and the head 48, the pair of elements is substantially formed according to Figure 10, inserted. 17 shows the outer ring 8, the inner ring 6, the sealing ring 7 and the conical elements 1 and 2, the piston 40 lies on the swivel = bed of the piston shoe 52 and it is secured against rotation in the embodiment of the figure by the pin 87, which protrudes into the bore to the exhaust valve, so that the Compression spring of the inlet valve 50 can be turned away from the inlet valve 50. The piston shoe 52 is arranged between the eccentric part 55, on the eccentric sliding or guide surface of which it lies and runs, and the piston 40. The shaft around the axis 97 carries the eccentric part 55 and its eccentric running surface generates the piston stroke when the shaft rotates in the bearings 114 which are arranged in the housing 130. The shaft with the axis 97 is driven by an electric motor, internal combustion engine or, as in the case of FIG. 7, by a hydraulic gate in the housing 48. This motor has, for example, the control 120, the cylinders 116, the pistons 117, in the rotor 113, the piston shoes 119 and the piston stroke guide 121 in the housing 148. The mode of operation of this motor can be found in patents of the inventor, for example in De BP 25 00 779. The bed surfaces, swivel surfaces, bearing surfaces and running surfaces or piston stroke guide surfaces of piston 40, piston shoe 52 and eccentric ring 55 are lubricated by lubricating fluid, hereinafter referred to as "oil". Due to the high radial forces, pressure fluid pockets 73 are also arranged, which carry most of the radial load on the piston and shoe. The surfaces and pressure fluid pockets mentioned are fed with pressure oil from the pressure oil chamber 130 via channels 74. The pressure pistons 96 axially press in the piston shoe 52 and seal the transitions of the channels 74. At the other end, a pressure bearing plate 115 or a further pressure piston 96 may be arranged axially movable in a pressure chamber 95. While a pumping unit is shown in FIG. 7, in practice the unit of the figure usually has 3, 5, 7 or 9 pumping chambers 61, pistons 40, pairs of elements 1, 2 and piston shoes 52 with the corresponding pressing bodies 96 and valves 49 and 50 in corresponding number of pieces. The suction or supply channels 122 are usually connected to one another. In practice, the drain or delivery pressure channels 123 are usually combined to form a single delivery channel.

• Beim Zusommendruecken der Elemente 1 und 2 entsteht eine Veraenderung der radialen Durchmesser der Elemente. Nach innenwerden sie etwas kleiner, radial nach aussen aber etwas groesser. Daher ist es bei onischen Elementen nicht immer moeglich, sie mittels der Spannringanordnung nach der Figur10 an den Aussendurchmessern zusammen zu spannen. Denn der Ring 9 der Figur dehnt sich infolge seiner Starke radial weniger aus, als die Elemente 1 und 2. Fuer sehr hohe Drucke und Durchbiegungen kann daher die Ring = anordnung 9 der Radialausdehnung der Elemente 1 und 2 nicht folgen und sie hindert diese Elemente an ihrer freien radialen Aus = dehnung. Erfindungsgemaess werden daher nach Figur 7 mit den Teile in separierten Teilschnitten darstellenden Figuren 8 bis 10 die Ring = teile 89,90 und 91 der Figur 7 in Segmente 32 A,B,C usw. aufgeschni: tten. Das ist in Figur 8 deutlich sichtbar. Man drhet zunaechst die Rin = ge, bohrt die Bohrungen und schneidet die Gewinde und danach schneidet man die Ringe mittels Scheibenfraesen in Radialsegmente. Gelegentlich werden die Segmente jedoch auch von anfang an als Segmente produziert. Die Elemente 1 und 2 erhalten dazu vorteilhafterweise an ihren radial aeusseren Teilen die Ausnehmungsring nuten 30, in die die Spannfinger 31 des oberen und des unteren Spannringes 89 und 91 eingreifen. Dadurch. wird verhindert, dass die Ringteile oder Segmente der Ringe 89,90,91 radial von den Elementen 1 und 2 wegrutschen koennen. Es ist zu beachten, dass die Stellen, an denen die Nuten 30 eingearbeitet sind, diejenigen Stellen der Ringelemente 1 und 2 sind, an denen diese bei ihrer achsi == alen Zusammendrueckung die geringsten inneren Spannungen erfahren. An der gezeichneten und zweckdienlichen Stelle am radialem Aussenteil der Elemente 1 und 2 schwaecht die Nut 30 die Elemente zwar etwas ab, doch ist das an diesen Stellen nicht schaedlich fuer die Lebensdauer der Elemente 1 und 2, weil sie an diesen Stellen trotz der Nuten 30 geringere innere Bie gespannungen (Zug- oder Druck - Spannungen) erfahren, als an den radial inneren und aeusseren Spitzen der EleMente 1 und 2. Der mittlere Ring istm meistens plangeschliffen, damit er in der Achsialhoehe, (Figur 7) der Achsialsummer des Aussenringes 8 und der beiden Elemente 1 und 2 entspricht, um genaues unnachgiebiges achsiales Spannen der Elemente 1 und 2 zu sichern . Die drei Ringe werden vorteil = hafterweise durch Schrauben 92 zusammengeschraubt, wobei die Anzieh = kraft mit Drehmomentmesser zu messen ist, um volle Spannung zu erreichen. Der Mittelring oder Distanzring 90 der Spannanordnung 89 - 91 hat vorteil = hafterweise eine innere Ringnut 93 zur Aufnahme des Aussenringes 8 der Figur. Denn ohne diese Ringnut wuerde der Aussenring 8 in radialer Richtung zu duenn, sodass er sich unter dem hohem Innendruck im Fluid in der Kammer 61 zu stark radial ausdehnen und damit innere Kompressionsverluste und Pumpverluste verursachen wuerde. Oder, andererseits wuerden die Spannschrauben 92 zu weit von den Spann- und Halte - Fingern 31 entfernt sein , sodass die Spannung an Wirkung verlieren wuerde infolge achsial er Nachgiebigkeit. Zwischen dem Kolben 40, dem Kopf 48 einer = seits und den Elementen 1 und 2 andererseits sind die plastischen Dichtun = gen 39,40 in Figur 7 angeordnet.

FIG. 7 also shows the following further specialty according to the invention, which is further illustrated by the partial figures 8, 9 and 10:

• When elements 1 and 2 are pressed together, the radial diameter of the elements changes. They get a bit smaller on the inside, but somewhat larger radially on the outside. It is therefore not always possible for onical elements to be clamped together on the outside diameters by means of the clamping ring arrangement according to FIG. 10. Because the ring 9 of the figure expands radially less than the elements 1 and 2 due to its thickness. For very high pressures and deflections, the ring arrangement 9 cannot follow the radial expansion of the elements 1 and 2 and prevents these elements their free radial expansion. According to the invention, the ring = parts 89, 90 and 91 of FIG. 7 are therefore cut into segments 32 A, B, C, etc. according to FIG. 7 with the parts in separate partial sections from FIGS. 8 to 10. This is clearly visible in Figure 8. You first press the rings, drill the holes and cut the threads and then cut the rings into radial segments using disc milling. Occasionally, however, the segments are also produced as segments right from the start. For this purpose, the elements 1 and 2 advantageously have the recess ring grooves 30 on their radially outer parts, in which the clamping fingers 31 of the upper and the lower clamping ring 89 and 91 engage. Thereby. prevents the ring parts or segments of the rings 89, 90, 91 from slipping radially from the elements 1 and 2. It should be noted that the locations at which the grooves 30 are incorporated are those locations of the ring elements 1 and 2 at which they experience the lowest internal stresses when they are axially compressed. The groove 30 weakens the elements somewhat at the drawn and expedient point on the radial outer part of the elements 1 and 2, but this is not harmful to the life of the elements 1 and 2 at these points, because in spite of the grooves 30 they remain at these points experience lower internal bending stresses (tensile or compressive stresses) than at the radially inner and outer tips of elements 1 and 2. The middle ring is mostly plane ground so that it is at the axial height (Figure 7) the axial number of the outer ring 8 and corresponds to the two elements 1 and 2 in order to ensure precise relentless axial clamping of the elements 1 and 2. The three rings are advantageously screwed together by screws 92, the tightening force being measured with a torque meter in order to achieve full tension. The middle ring or spacer ring 90 of the clamping arrangement 89-91 advantageously has an inner annular groove 93 for receiving the outer ring 8 Figure. Because without this annular groove, the outer ring 8 would thin in the radial direction, so that it would expand too much radially under the high internal pressure in the fluid in the chamber 61 and thus cause internal compression losses and pumping losses. Or, on the other hand, the tensioning screws 92 would be too far away from the tensioning and holding fingers 31, so that the tension would lose its effect due to its axial compliance. The plastic seals 39, 40 in FIG. 7 are arranged between the piston 40, the head 48 on the one hand and the elements 1 and 2 on the other hand.

The unit of FIG. 7 thus already forms a fairly ideal pump or motor unit for the supercritical range of the extremely high pressure in the fluid. The elements 1 and 2 can expand and contract radially freely, since the segments A, B.C --- X, Y.Z of the rings 89 to 91 ideally follow this radial expansion and contraction.

In FIG. 11, the rotor 98 has recesses, for example bores 107. The seats for elements 1 and / or 2 or for element columns 1, 2 and 2. The elements 1, 2 have between their radially outer superposed = supports advantageously plastic seals 37, 38, for example according to FIG. 6, and at their radially inner ends also advantageously = also plastic seals 39, 40, for example those according to FIG. 6 Pistons or the pistons 36 have piston fingers which enter the chambers 61 and can also center the elements 1, 2 on one another. The pistons 36 carry the piston shoes 21 pivotably in pivoting pans. The piston shoes 21 run on the inner surface 156 of the piston stroke guide 99, as a result of which they form the piston stroke guide surfaces. The piston shoes have pressure fluid pockets with surrounding sealing surfaces as hydrostatic bearings with pressure fluid pockets 112 and such pressure fluid pockets 111 are also arranged between the socket in the piston and the swivel surface of the piston shoe. The pressure fluid pockets are fed from the working chambers 61 through channels 74 with pressure fluid. A distance is arranged between the axis 103 of the rotor 98 and the axis of the piston = stroke guide surface 156, so that the piston stroke = guide surface 156 is eccentric to the rotor axis 103, so that when the rotor 98 or the piston stroke guide 99 rotates, the piston stroke of the pistons 36 and the piston shoes 21 is formed, the elements 1 and 2 periodically compressing and relaxing and thereby periodically enlarging and reducing the working chambers 61 once per revolution.

The fluid supply and discharge to and from the working chambers 61 takes place when rotor channels 161 are arranged by axial fluid flow or by internal radial fluid flow. For example, by entering fluid from channel 105 of control shaft 102 via control window 150 into channels 161 and from there into working chambers 61 and channels 74, such as pressure fluid pockets 11, 112, and from these through the corresponding further channels 161, control window 149 in control body 102 and which flows through channel 106. However, an external radial flow is also possible such that the piston stroke guide receives a channel 101 through which fluid from the space between rotor 98 and piston stroke guide 99 or from a container connected to channel 101 through the pressure fluid pockets 112 and channels 74 into the Working chambers 61 flowing. The particular advantage of the unit according to FIG. 11 is that, in the case of multi-element columns with elements, 1.2 long working chamber lifts of the chambers 61 are permitted, thus allowing a large flow rate in a small space. In addition, this unit is particularly simple and inexpensive to manufacture, since it does not require any pistons with a tight fit and is therefore simpler and cheaper than the conventional radial piston units with pistons which are tightly fitted into cylinders and are provocative therein. However, this unit is only for the subcritical, low pressure range in the fluid if it has no clamping arrangements according to FIGS. 10 or 17. It is converted into a high-pressure unit for the supercritical high pressure range by assigning elements 1 and 2 clamping arrangements 9, 89 to 91 of FIG. 7.

FIG. 17 shows an element 1 or 2, here designated 307 in a half radial section in approximately natural size for the supercritical high pressure range. This figure is given to show the annular groove 358, which corresponds to the groove 30 of FIG. 9, particularly clearly and, above all, to show that the inner flange surface 359 and the outer plane surface 354 must be arranged. The surface 359 for storage on the piston, the pump head or the element of the next element = pair or the next element and the surface 354 for storage on the outer ring 8 of FIG. 7 or for storage on a piston or pump head of one of the figures or not drawn in the figures. In order to reduce the internal tension and thus the risk of breakage of the elements, the inner edge 357 is advantageously rounded,

Figure 18 describes another problem. If the elements have no flats 359, 354, then they are common disc springs and then lie on a line. At the high pressure in the supercritical pressure range, this receives a pressure of many thousands of kilograms per square millimeter, namely an infinitely high internal tension the line of contact does not form an area at all, not even a total area of one square millimeter. The material melts away under this infinitely high tension in the material. The litany deforms and damages the surface and the outer layers of element 1 or 2. The flattening to the flat surfaces 359,354 reduces this danger and appearance, but does not completely eliminate it. It is therefore advantageous, in order to further limit this phenomenon, to provide the support counterpart 360 with a bulge 355, which is shaped and dimensioned such that the edge of the element 1 or 2, here designated 307, slides on it during radial expansion and Contraction under the axial collapse and expansion of the element 1,2,307. The shape should also be such that at least approximately the line support is converted into an approximate surface support. If the axial assembly is only in size according to angle alpha, then the counter bearing receives a bearing surface in angle beta, so that after compression, the entire element 1 or 2 lies well on support 360 and cannot bulge axially. After the working chamber 307 to put a sealing surface (layer, sheet) 309, e.g. of copper, teflon, etc. on the element 307, so that it is not affected by the rust-causing fluid. The plastic seal 356 is arranged accordingly, it being expedient for it to encompass the sealing layer or protective layer 309. The tip of element 1 or 2.307 must not lift off the support 360 and curve 355 too much so that the seal 356 or 309 cannot enter the gap. In high-pressure practice, the tip must not lift more than 0.03 to 0.1 mm. Otherwise the seal 356 becomes a thin black paper layer which enters the gap and destroys the seal. While the seals in FIG. 18 are hatched, the seals 356, 309, 317 etc. in FIG. 12 are not drawn hatched, but are left open without lines, because the hatching lines in FIG. 12 would make them too indistinct.

FIG. 12 shows a particularly effective and particularly high-pressure capable unit of the invention, and a particularly important finding of the invention is described on the basis of it.

It is the case that elements 1 and 2 of FIG. 7 are similar to disc springs and are therefore only usable for limited high pressures. According to this current invention, it is recognized that conventional disc springs are absolutely unsuitable for effective use in the high supercritical pressure range, because they break after only a short operating time. The cause of the cracking could not be found for a long time because the disc springs still have the formulas of Alpine pastures and Laszio were calculated. These Americans published a math for the calculation of the internal stresses in disc springs in 1935, which was later summarized in Germany. After this summary it was easier to calculate, but the deeper connections, which were clearly recognized by the pastures and Lascio, were lost. The disc springs and their stress calculations were later standardized in DIN and JIS (German and Japanese Industry Standards). According to this standardization, elements 1 and 2 should have a long service life if the standardization, i.e. the standard sheets, is observed.

After a long search, the present invention has recognized that the standard beds and the standards according to DIN and JIS are absolutely impossible and unusable for the calculation of the lifespan and the stresses for elements 1 and 2 according to this invention.

On the right in Figure 19 you can find the calculations according to Imen and Lascio, which are more precise and extensive than those according to DIN and JIS. The curve 361 in FIG. 15 shows the largest tension sigma within the relevant element 1 or 2 as a result of the compression above the circumferential angle alpha of the eccentric ring of the piston stroke drive. This calculation is based on that of Almen and Lasczio.

The invention now recognizes that the high fluid pressure in the supercritical range causes a much higher bending stress and a much more sudden bending stress within the element 1 or 2. It is this tension which, when the radial dimensions of the plate spring are approximated, brings the elements of FIG. 7 to break quickly, even after 45 minutes at an internal pressure of 1000 bar.

The invention recognizes that these further internal stresses in element 1 or 2 result from the fluid pressure in chamber 61. This fluid pressure is shown by means of the arrows "q" in FIG. 13 and it the additional bending stresses sigma B shown in FIG. 13 around the center point "O". In the supercritical, high pressure range, these bending stresses are many times higher than the internal stresses of the disc springs after breathing and Lascio, which are simply labeled "sigma" without "B" in the figures.

The invention has now recognized this previously unknown cause of the disc springs breaking very precisely and mathematically very precisely. This can be seen from FIG. 13, which, according to the inventor, records the moments and tensions differentially and integrally. Figure 13 gives the exact geometric and mathematical relationships, as well as their integrations with the help of the differential and integral calculation. Based on equation (1), the approach of the bending moment under fluid pressure, the mathematics finally leads to formulas (12) and (13) with the values of the stresses. The indices "I" and "o" also show the indices "B" the meaning of whether they have been viewed radially from the inside or radially from the outside.

In FIG. 14, it is demonstrated methematically and geometrically where the radii of the same stresses lie when one starts to calculate radially from the inside or outside, that is to say from the outside radius or inside radius of the element, the values "Rmc" and "RMC" show the bending moments under fluid pressure without taking into account the tensioners or bearings, while the values "RCMH" and Rcmh "take into account the moments with brackets or supports. The extensive calculations and investigations can be found in the extensive Rotary Engine Kenkyusho reports RER-8109 and RER- 8206. One can see from FIG. 14 the final calculation formulas according to the piff erentations and integrations, and one also sees that these values all deviate from the arithmetic mean radius "Rm" and also from the pressure point radii the control body patents of the Inventor, for example, differ from the Rgc value of the inventor's De-Patent 23 00 639. This is due to this n that the center of the area applies to the control body, whereas here in the unit with elements 1, 2 of the invention it is not the area values but the fluid pressure torque values that apply.

The invention recognizes from this that conventional disc springs = areas for use as elements 1, 2 are only possible and usable for low pressures. The elements according to the invention must therefore still meet the condition for the supercritical high pressure range that the difference between the inside and outside radii of the elements is more than three times smaller than the outside radius of the element 1 or 2 in question. In the figures for high performance = aggregate, Figures 12 and 19 is the fluid pressure work area. this radial radius difference is approximately five times smaller than the outside radius of the element in question.

If one considers the Almen and Lascio equation in FIG. 19, then it is true according to the knowledge of the invention that the difference "(C minus eta)" under the fraction has the most important meaning. This results in the fact that the internal stresses "sigma" become smaller, the larger the diameter of the element 1 or 2, with otherwise constant thickness, radial difference and the angle of attack or deflection of the element. In order to make the element usable for the supercritical range of high pressure for a long service life, one must make the radial difference very small, for example about ten millimeters, the deflection "f" also very small, for example below 0 , 5 millimeters and make the element very thick, for example 5 to 10 millimeters, in order to keep the bending stresses "Sigma B" according to FIG. 13 of the invention small and at the same time also make the diameters and radii large in order to reduce the internal stresses "sigmo" after Almen and Lasczio according to Figure 19.

However, according to the knowledge of the invention, the problem arises that the radial loads under the high fluid pressure in the chambers or in the chamber (working chamber) 61 then become so great that there are no longer any rolling bearings which bear these loads with a sufficient service life could be in tolerable dimensions. The invention thus recognizes that neither conventional disc springs, nor conventional rolling bearings of the standard series ir _{9 and} what usable elements with a long service life for motor or pump units can deliver units for the supercritical range of high pressure, in short, that the previous technology does not have any motors or pumps Can create aggregates for the supercritical area with the high pressures present therein.

Therefore, in FIG. 12, a first piston 302 is reciprocally arranged in the cylinder, the first cylinder, 301, in order to convey a fluid pressure column, which is mostly hydraulic oil, but can also be another fluid, through the channel 303 into a second cylinder 304 of larger diameter and onto which the Second cylinder reciprocable second piston 305 presses.

The second piston 305 carries the element 307, compresses it axially and thus conveys the pressure medium out of the working chamber 311 through the outlet means 313. Several elements 307 can also be arranged and clamped together by brackets or clamping arrangements 318, 320, 321, 319 and held together. When the first piston 302 moves outward in the first cylinder 301, the pressure fluid flows out of the second cylinder 304 through the connecting channel 303 between the first cylinder and the second cylinder First cylinder 301 back. So there is periodic transport of a pressure fluid column from the first cylinder to the second cylinder and then back from the second cylinder to the first cylinder. The second piston periodically swings inversely in parallel to the first piston. If the first piston is pushed into the first cylinder 301, then the second piston 305 is pushed outwards in the second cylinder 304 and vice versa, if the second piston, for example, presses in under the relaxation pressure of the element (elements) 307 in the second cylinder, then the first piston in the first cylinder is pressed out. In practice, this happens in that the first piston 302 is forced to a piston stroke by a piston stroke drive 336, for example via a piston shoe 334. For example, the piston stroke guide 336 rotates in the bearings 338, and therefore rotes in the bearings 338. In addition, the piston stroke guide may be driven, for example, by the drive 345 346 for rotation, for example by an electric motor, combustion motor or by fluid motor. For example, if the inner surface 347 of the piston stroke guide 336 is arranged eccentrically to the bearings 338, so that the bearings 338 have a central axis, but the inner surface 347 has an eccentric axis that is parallel to the central axis but radially distant from it, then causes when the piston stroke guide 336 rotates, the eccentric piston stroke guide surface 347 has a radial stroke by the sliding surface of the piston shoe 334 on it, which the piston shoe 334 transmits to the first piston 302. The first piston 302 is moved in and out once in the first cylinder 301 per revolution. This radial movement of the first piston, namely the piston stroke of the first piston by means of the channel 303 and the fluid column between the first piston and the second piston, is reversely transmitted in parallel to the second piston 305. In practice, the Repla gets ^y linder 301 and 302 to keep the Erstkolben a smaller diameter and the second cylinder 304 to the second piston 305 a substantially larger diameter, thereby it becomes possible that Radialkraefte from Erstkolben 302 so low that for the pistons = lift guide commercial roller bearings 338 can still be used and still have a sufficiently long service life. In this arrangement according to FIG. 12 of the invention, the second piston 305 does not require any roller bearings. The big problem is that there are 305 no more handelsueblichen Bearings the Zigtonn _E n forces of the second piston to a rotating cam shaft or a rotating eccentric ring, as shown in Figure 7 for the high Kraofte of the second piston, completely overcome. First piston and second piston now run in Figure 22 in non-rotating, solid, stable bodies with stationary first and second cylinders 301, 304 in it. In the example in FIG. 12, the first cylinder with the first piston is a radial arrangement, but it could also be an axial arrangement. The second cylinder and the second piston could also be a radial arrangement, as in FIG. 7, but in FIG. 12 they are an axial arrangement. This makes it possible, for example, to arrange 3, 5, 7, 9 or any other desired number of axial cylinders around the central axis 330 of the unit and to arrange the corresponding number of first cylinders and first pistons. It should be borne in mind that at least one first cylinder must be connected to a second cylinder through the connecting line or the connecting channel 303. In the case of a plurality of first cylinders and second cylinders in the unit in FIG. 12, separate first and second pistons each oscillate, without these having to be connected to other first or second pistons. The channels 303 must remain unconnected with other first and second cylinder arrangements. Each first-second piston arrangement in a common housing thus has a pressure-fluid column in a channel 303 which oscillates independently of other fluid columns and which projects into the connected first and second cylinders, but not to others of the first and second cylinders in which the unit may be connected. But that doesn't mean that, for example, two or more first pistons can or can work together with a single second piston and corresponding cylinders.

It is also possible to make the operation of the unit according to FIG. 12 continuously variable, that is to say to make the piston stroke of the first piston and, as a result, to make the piston stroke of the second piston infinitely variable. For this purpose, the drive from drive 345 to piston stroke guide part 336 in FIG. 12 is made, for example, radially flexible. For example, by inserting variable entry pins of the tooth 344 of the shaft 345 into the teeth 346 of the piston stroke drive, or by arranging an intermediate wheel movably between the toothed wheel 344 of the shaft 345 and the toothed wheel 346 of the piston stroke guide 336.

Finally, in practice it is often important to always ensure the correct amount of fluid in the fluid column of channel 303 with connected cylinders 301 and 304. This is achieved, for example, by the arrangement of a control 326 to 329. The first piston 301 here has a control extension 326, which is provided with control grooves 328, which periodically flow the fluid supply line 329 depending on the piston stroke of the first piston 302 with the first cylinder = 301 connect and disconnect from it. For this purpose, the control extension 326 plunges sealingly into the control cylinder 327 and moves in it.

In this way it can be achieved that when the first piston reaches its outer position in the cylinder 301 and the second piston 305 has approximately reached its inner position in the second cylinder 304, the fluid feed line 329 via the control channel 328 with the first and second cylinder 301,304 is connected and fills this and the channel 303 with the required amount of fluid, or refills if some fluid from the column of rooms 301, 303,304 should have been lost through leakage during the previous work cycle. In order for this to function properly, it is useful to provide the second piston 305 with inner end position stops 306 which strike the cylinder bottom of the second oil cylinder and thus determine a precise inner end position of the second piston 305 in the second cylinder 304. This is important together with the full filling of the fluid in the mentioned fluid column, so that the subsequent pressure stroke of the first piston 302 results in an exact length and position of the pump stroke of the second piston 305 in the second cylinder 304 and thus the elements 307 are pressed together in the precisely desired stroke length . This is important because the elements 307, 1.2 in these high pressure ranges of the supercritical range only make very short strokes of less than one millimeter per element 1,2,307. Only the sets of the elements are 1,2,307 longer. The safe mode of operation of the fluid column is therefore of crucial importance for the reliable operation of the assembly of FIG. 12, as well as for full high-pressure delivery with good efficiency.

Figure 12 also has the inlet means 316 to the working chamber 311, the outer ring 320, the inner ring 308, the bore 350 therein for connecting the two working chambers 311, and the seals 309, (protective skin) 348, 322, 317, etc., wherein Figure 18 is the most perfect form of a 12 shows part, and an enlargement of the re-liner 348 of FIG. 12, axis 331 shows the eccentric axis, that is the axis of the piston = stroke guide surface 347 and its distance (radial distance) 332 from the central axis 330. The interior spaces 351 and 352 can refill with lubricant, fluid to avoid rusting of the parts inside the unit.

In Figure 13, the bars "H" indicate the bearing of the element on the piston or pump head, or the clamping of two of the elements. The important meaning of FIG. 23 is that it teaches that conventional disc springs can only be used for the low pressure ranges, but for the supercritical range of the high pressure the difference (R minus r), but about a fifth or less of the outside diameter "R " should be. "S" shows the thickness of element 1,2,307.

A comparison of the teaching from FIG. 13 with FIG. 12 shows that the main part of the flow rate through the working chamber 61 is now achieved by compressing the chamber 61 in the axial direction. The elements are now only involved in a small percentage change in the volume of the working chamber 61.311. And that's how it should be in the ideal case. The elements in FIG. 12 are now more means of limiting the working chamber in the radial direction than fluid conveyors = elements. In the subcritical area, however, the elements 1, 2 are often significant participants in the change in the volume of the working chamber and their conveying effect is then greater than that of the pure axial movement of the piston in question.

In FIG. 19, the mathematical development of the calculation of the change in chamber, that is to say the amount conveyed by chamber 61, is shown on the left. Here one sees as "V" the change of the volume of the working chamber 61 by the element 1,2, while "Q" is the sum of the volume change by elements 1,2 and axial movement of the piston concerned. If the unit works as a pump, "V" is the amount conveyed per stroke by the element 1,2 in question, while "Q" is the sum of the amount conveyed per stroke from element 1 and the axial movement of the piston (e.g. 40.305).

FIG. 19 also shows that two or more elements 1, 2 do not always have to be arranged, but that a single element 1 or 2 can also be arranged between a piston 40 and the pump head 48. Here you can also clearly see the protective skin 777, the protective layer that can be inserted as a sheet or so made of appropriate material, e.g. Copper, Teflon, rubber to prevent the element 1,2 from being touched by water or the like. Water on element 1,2 may reduce its lifespan to one eighth. The seal 888 appropriately encompasses the protective skin 777 and the seal 7 is touched through the protective skin to ensure proper sealing of the chamber 61 and proper seating of the outer part (in the ring 3) of the element 1, 2.

In the upper right part of FIG. 19 you can see the numbers 1, 2, 3 in circles. These mean the relevant places (locations, positions) of the internal stresses sigma (ring 1), (ring 2), (ring 3) for calculation according to the Alms and Lasczio equation,

FIGS. 15 and 20 also show that the bending stresses "sigma with index B", that is to say the stresses under fluid pressure in the working chamber 61, increase much more suddenly than the stresses sigma when element 1 or 2 are compressed, because the full stresses occur due to fluid pressure already when the fluid, for example the water, is fully compressed to the high pressure. It is therefore very important to consider the rotation angle alpha of the rotary part that produces the piston boy. Water, for example, roughly compresses four percent at 1000 bar, or about 16 percent at 4000 bar. The maximum voltage in element 1, 2 due to fluid pressure therefore occurs suddenly, for example according to 362 in FIG. 15 or according to sigma Bi in FIG. 20. The highest voltage due to fluid pressure is therefore often reached at less than a twentieth of the rotor revolution, while that by compression the element 1, 2 is reached by the piston 40,305 only after half a rotor revolution. Therefore, the life of the element, the rapid rise in voltage by fluid pressure a bigger impact than the Sp _Q nnungen by compression of the element by the piston. These things had to be examined intensively according to the invention and brought into balance according to FIGS. 15, 20, so that the hitherto impossible service life, which disc springs cannot offer, is achieved for the elements 1, 2, 307, etc. of the invention. In FIG. 20 the sigma B curve on the right drops a little, because in this figure it is taken into account that the element near the completion of the compression stroke already partially comes to rest on the support, the piston, or the pump, motor heads, which the tension is reduced somewhat, but only gradually over the circumferential angle.

For the calculation of the inertia forces and the fluid flow velocities, the mathematical formulas for the stroke "S", the piston velocity "Vs" and the piston acceleration "bs" are given in FIG. 19, again over the rotation angle alpha.

Thus, the patent application gives a full lesson on technical action combined with the knowledge that conventional means are not aimed. can not lead to the solution of the problem and the new means and knowledge of the present invention must be used for the supercritical area of high pressure.

From about ten thousand atmospheres of water pressure it must be taken into account that the sound waves (pressure waves) propagation speed = in the water is reached. The specification in German-language catalogs of some companies that a water jet emerges from the nozzle at twice the speed of sound can be misleading, because the speed of sound of water is around 1400 meters per second, not at around 33o meters per second, as that of the air.

• oder; dass zwei Elemente 11,12, zwischen sich einen Hohlraum 61 bildend, an den radial aoussoren Enden mittels einer achsial starken Umgrei fung 9 aneinander geklemmt sind;
• oder; dass zwischen den Elementen 11,12,1,2, ein Aussenring 8 , ein Innenring 6 mit einer Bohrung und zwischen dem Innenring und dem Aussenring eine die Elemente beruehrende plastische Dichtung (Dichtring, O-Ring) 7 angeordnet sind;
• oder; dass achsial endwaorts der genannten Elemente Stutzkoerper 10 mit Auflageflaechen und mit Ausnehmungen 17 angeordnet sind, deren Radialabmessungen aussen der Umgreifung nahe sind;
• oder; dass Zentrierkoerper 19 oder 20 mit Zentriersitzen 27,26, 28,29 angeordnet sind, die teilweise in Sitze 4,25 an Elemen = ten 1 ,2 ansetzbar sind und so jeweils mindestens zwei Elemen = te 1,2 zueinander oder Paare von Elementen 1,2 zueinander zentrieren und halten;
• oder; dass ein Treibkolben 59 mit einem Purnpkolben 58 in Achsial = richtung druckfaohig und in Radialri chtung relativ zueinander etwas nachgiebig, verschiebbar, verbunden ist, wobei der Treibkolben in einem Zylinder und der Pumpkolben in einem anderem Zylinder 61 reziprokierbar ist und zwischen den genannten Zylindern ein die benachbarten Enden der Kolben 58, 59 umgebender Raum 67 angeordnet ist, der durch eine Lei= tung 68 druckarm oder druckleer gehalten sein kann; (Figur 2.)
• oder; dass die beiden Kolben 58 und 59 du ch einen etwa achspara = Ilelen Stift 70,64,71 miteinander verbunden sind und zwischen einer Bohrungswand in einem der Kolben und dem genanntem Stift ein Raum 69 angeordnet ist, der radiale Beweglichkeit der beiden Kolben zueinander gestattet, ohne die achsiale Kraftschluessigkeit in mindestens einer der achsialen Bewegungs ₌ richtungen aufzuheben; ( Figur 2. )
• oder ein Teil des Pumpkolbens von einer Ringkkammer 65 mit Abflusskanal 66 umgeben ist und diese Kammer zur Abfueh= rung von Leckage oder Mischfluid eingesetzt sein mag. (Figur 2.)
• Oder; dass Elemente 1,2 zwischen Radialplatten 42,41 mit radial innen oder aussen vorstehenden Zylinderteilen 43,44, 45,46 zentriert und gehalten sind, wobei die endwaertigen Elemente 1,2 in Sitzen 47,53 benachbarter Teile 48,52 gehalten sein moegen; ( Figur 3.)
• oder; dass ein federbelasteter Kolben 60 einen Planflaechenkopf 86 bildet, auf den die plane Endflaeche eines Zweitkolbens 59 drueckt, wobei radiale Ungleichheit der Kolbenachsen in Kauf genommen werden kann und / oder ein Kolbenschuh 52 mit Fuehrungsfingern 81 versehen ist, deren Gleitteile 79 an Laufflaochen 82 einer Ausnehmung 80 in einem Koerperteil fuer einen langen Kolbenschuh - Hub gefuehrt sind; ( Figur 4.)
• oder; dass an einem radialem Endteil eines Elemontes eine plastische Dichtung 38 in das Element 1 eingearbeitet ist, und / oder in das radial innere Teile des Elementes 1 eine plastische Dichtung 39 eingearneitet ist, wobei die genannten Dichtungen in die Betten 37 oder 40 eingelegt oder eingeklebt oder ein = vulkanisiert sein moegen; ( Figuren 5 und 6) .
• oder; dass die genannte Umgreifung 9 in einzelne Segmente A,B,C, .. ..... X, Y, Z durch etwa radiale Schnitte untertei lt ist, und / oder die genannte Halterung aus einem oberem Spannteil 89, einem mittlerem Distanztei 90 und einem unterem Spannteil 91 besteht, wobei auch diese Teile in Segmente unterteilt sind; (Figuren 7 bis 10.)
• oder; dass der obere und/oder der untere Spannteil 89,91 mit Haltefingern 31 versehen ist, die in eine entsprechend geformte Nut 30 am radial oeusserem Endteil des betreffenden Elementes 1 oder 2 ausgebildet ist, eingreifen und dadurch das radiale Ab = gleiten des betreffenden Heltesegmentes 89 A-Z oder 91 A-Z vom Element 1 oder 2 vermieden ist, und/oder dass die genannten Halteteile oder Spannteile 89 bis 91 durch Halte- oder Spannmittel, wie z. B. Bolzen 92 miteinander gespannt verbunden sind; (Figur 7 oder Figur 12.)
• oder; dass das Aggregat fuer den superkritischen hohen Druck = bereich eingesetzt ist oder einsetzbar ist und/oder der genannte Aussonring 8, plastische Dichtring 7 und der genannte Innenring 6 zwischen den beiden Elementen 1 und 2 angeordnet sind;
• oder; dass ein Fluidmotor 113,117,119,121 eine Welle treitt, die einen Exzonterring mit zur Wellenachsa exzentrischer, zy = lindrischer Kolbenhubfehrungsflaoche am Exzenter 55 bildet, auf der ein Kolbonschuh 52 gleitet, dessen Druckfluidkammcrn 73 durch Kanaele 74 aus achsialen Anprosskammorn mit abdichtenden Druckkolben 95,96 gespeist werden, auf der Schwenkflaeche dos Kolbenschuhes ein Kolben lagert und dieser Kolben 40 auf eines der Elemente 1 oder 2 drueckt und die Zusammendrueckung des betreffenden Elementes und damit den Pumpvorgand in der Arbeitskammer 61 treibt oder die Arbeitskammer den Kolben und Schuh und damit den Exzentorring 55 treibt und in rotierende Bewegung versetzt; (Figur 7 oder /2
• oder; dass in einem Koerper 98 Sitze 107 ausgebildet sind, die Elementenpaare 1 ,2 halten, eine Kolbenhubfuohrung 156 fuer den radialen Kolbenhub des Kolbens 36 ueber Kolbenschuh 21 zwischen dem Elementensatze 1,2 und der Fuehrung 156 ange ordnet ist und Fluidzuleitungskanaele 102,105 und Ablei = tungskanaole ueber einen Umsteuervorgang den zwischen den Elementen 1 und 2 gebildeten Rrbeitskammern 61 verbunden sind, wobei ein Kolbenfortsatz des Kolbens 36 die Elemente zentrieren oder halten mag und jede dichtene Einpassung von Kolben in Zylindern vermieden ist, um so das einfache und billige Pump - oder Motoren - Aggregat der Figur 11 zu verwirklichen; ( Figur 11. )
• oder dass an dessen Achsialenden radial innen oder aus sen radial etwa plane Flaechen, Auflage-Flaechen 354 und / oder 359 angeordnet sind; ( Figur 17.)
• oder dass nach Figur 18 eine Formgebung und eine Ausbauchung 355 die Linienberuehrung in eine fast Fiaechenbe= ruehrung umwandelt um groessere Tragkraft zu erzielen und / oder Schutzschichtplatton, sheets, Bleche, 309 und / oder Dichtmittel 356 angeordnet sind; ( Figur 18.)
• oder; dass ein Erstzylinder 301 angeordnet ist, in dem ein Erstkolbon 302 roziprokiort, ferner ein Zweitzylinder 304 angeordnet ist, in dem ein Zweitkolben 305 reziprokiert und der Erstzylinder mit dem Zweitzylinder durch einen Verbin = dungskanal 303 verbunden sind, die Zylinder und der Kanal mit Fluessigkeit gefuellt sind und die Reziprokierbewegung des Erstkolbens durch die Fluessigkeitssaoule in dem Kanal und den Zylindern auf den Zweitkolbon uobortragen wird; ( Figur /2.)
• oder; dass der Erstkolben 302 kleinen Durchmesser, der Zweit = kolben 305 3inen groesseren Durchmesser hat, wodurch die Druckkraft des Zweitkolbens groesser, als die des Erstkol = bens wird und die Bewegung des Zweitkolbens umgekehrt parallel zum Erstkolben erfolgt, indem der Zweitkolben im Zweitzylinder einwaerts laeuft, wenn der Erstkolben im Erstzylinder auswaerts laeuft und umgekehrt (vice versa);
• oder; eine Steuerung 326,329,328,327 zur korrekten Auffuellung der genannten Fluidsaeule oder Fluessigkeitssaeule in den Raeumen, Zylincern, Kanaelen, 301,303,304 ange= ordnet ist, die zur Zeit niederen Druckes, bevorzugt in der aeusseren Kolbenlage oder der Noehe derselben, des Erst = kolbens 302 die Zyl inder 301 und 304 und den Verbindungs= kanal 303 mit Fluid aus dem Zuleitungskanal, z.B. 329 heraus fuellt, damit eine einwandfreie Uebertragung der Bewegung des Erstkol bens 302 auf den Zweitkolben 305 erfolgt;
• oder; dass der Kolbenhub des Erstkolbens 302 stufenlos regel = bar angeordnet ist und dadurch der Kolbenhub des Zweit = kolbens 305 auch stufenlos regelbar wird;
• oder; dass das Element 1,2,307 usw. einen Ausoendurchmesser und einen Innendurchmesser hat, deren Radien die Haelfte des betreffenden Durchmessers sind und die Differenz Aussenradius minus Innenradius mindestens dreimal kleiner, als der Aussenradius ist. ( Figur 12); (Entstanden aus Figuren 13, 14, 15, 19, 20.)
• oder; dass der Erstkolben 302 mit einer Kolbenhuh - Antriebs = vorrichtung, z.B. 334, 347, 336, 346, 344, 345 oder der = gleichen versehen ist, die den Hub des Erstkolbens 302 er= zwingt; ( Figur /2 ) .
• oder; dass nur ein Element 1 zwischen einem rezi = prokierbarem Kolben 40 und einem Kopfe 48 angeordnet ist; ( Figur 19.)
• oder; dass zwecks Erreichung einer maximalen Lebensdauer der Elemente die inneren Spannungen in den Elementen 1,2,307 usw. durch Zusammendruecken und durch Durchbiegung unter innerem Fluiddruck aus der Arbeitskammer 61 gegeneinander angeglichen sind. (Figuren 15, 20).

Summarizing the previous figures with conical ring elements, it can be seen that the special designs and features of the invention also consist in
that a pair of elements 1, 2 made from fibrous materials with connecting substances between them at the radially outer ends, forming a cavity 61 between the two elements, is bonded, bonded or adhesively bonded; Figure 1)

• or; that two elements 11, 12, forming a cavity 61 between them, are clamped to one another at the radially outer ends by means of an axially strong circumference 9;
• or; that between the elements 11, 12, 1, 2, an outer ring 8, an inner ring 6 with a bore and between the inner ring and the outer ring a plastic seal (sealing ring, O-ring) 7 touching the elements are arranged;
• or; that axially end words of said elements support body 10 with support surfaces and with recesses 17 are arranged, the radial dimensions of which are close to the encompassing outside;
• or; that centering bodies 19 or 20 are arranged with centering seats 27, 26, 28, 29, which can partially be placed in seats 4, 25 on elements 1, 2 and thus each have at least two elements 1, 2 to one another or pairs of elements 1 , 2 center and hold to each other;
• or; that a driving piston 59 is connected to a pump piston 58 in the axial direction = pressure-sensitive and in the radial direction relatively flexible, displaceable relative to one another, the driving piston in one cylinder and the pump piston in another cylinder 61 being reciprocal and between the cylinders mentioned adjacent ends of the pistons 58, 59 surrounding space 67 is arranged, which can be kept low-pressure or low-pressure by a line = 68; (Figure 2.)
• or; that the two pistons 58 and 59 are connected to each other by an approximately axial pin 70, 64, 71 and a space 69 is arranged between a bore wall in one of the pistons and the mentioned pin, which allows the two pistons to move radially with respect to one another, without the axial Kraftschluessigkeit in at least one of axial motion ₌ directions annul; (Figure 2.)
• or a part of the pump piston is surrounded by an annular chamber 65 with a drain channel 66 and this chamber may be used for the removal of leakage or mixed fluid. (Figure 2.)
• Or; that elements 1, 2 are centered and held between radial plates 42, 41 with cylinder parts 43, 44, 45, 46 projecting radially inwards or outwards, the final elements 1, 2 being able to be held in seats 47, 53 of adjacent parts 48, 52; (Figure 3.)
• or; that a spring-loaded piston 60 forms a flat surface head 86, on which the flat end surface of a second piston 59 presses, whereby radial inequality of the piston axes can be accepted and / or a piston shoe 52 is provided with guide fingers 81, the sliding parts 79 of which on sliding blocks 82 of a recess 80 are guided in a body part for a long piston shoe stroke; (Figure 4.)
• or; that a plastic seal 38 is incorporated into the element 1 at a radial end part of an element and / or a plastic seal 39 is introduced into the radially inner parts of the element 1, the seals mentioned being inserted or glued into the beds 37 or 40 or a = may be vulcanized; (Figures 5 and 6).
• or; that the mentioned gripping 9 is divided into individual segments A, B, C, .. ..... X, Y, Z by approximately radial cuts, and / or the mentioned holder consists of an upper clamping part 89, a middle spacer 90 and a lower clamping part 91, these parts also being divided into segments; (Figures 7 to 10.)
• or; that the upper and / or the lower clamping part 89, 91 is provided with retaining fingers 31, which are formed in a correspondingly shaped groove 30 on the radially outer end part of the element 1 or 2 in question, and thereby the radial sliding of the relevant holding segment 89 AZ or 91 AZ is avoided by element 1 or 2, and / or that the holding parts or clamping parts 89 to 91 mentioned by holding or clamping means, such as. B. bolts 92 are tensioned together; (Figure 7 or Figure 12.)
• or; that the unit is used or can be used for the supercritical high pressure = area and / or the said outer ring 8, plastic sealing ring 7 and the inner ring 6 are arranged between the two elements 1 and 2;
• or; that a fluid motor 113, 117, 119, 121 hits a shaft that forms an eccentric ring with a cylindrical piston stroke guide plate that is eccentric to the shaft axis, on which eccentric 55 slides, on which a piston shoe 52 slides, the pressure fluid chamber 73 of which is made by channels 74 made of axial contact chamfer with sealing pressure pistons 95, 966 a piston is supported on the swivel surface of the piston shoe and this piston 40 presses on one of the elements 1 or 2 and drives the compression of the element concerned and thus the pumping element in the working chamber 61 or the working chamber drives the piston and shoe and thus the eccentric ring 55 and in rotating movement offset; (Figure 7 or / 2
• or; that seats 107 are formed in a body, which hold element pairs 1, 2, a piston stroke guide 156 for the radial piston stroke of the piston 36 via piston shoe 21 is arranged between the element sets 1, 2 and the guide 156, and fluid supply ducts 102, 105 and discharge duct ducts The working chambers 61 formed between the elements 1 and 2 are connected via a reversing process, a piston extension of the piston 36 being able to center or hold the elements and any tight fitting of pistons in cylinders is avoided, in order to simplify the simple and cheap pump or motor operation. Realize the unit of Figure 11; (Figure 11.)
• or that at its axial ends are arranged radially on the inside or radially approximately flat surfaces, support surfaces 354 and / or 359; (Figure 17.)
• or that, according to FIG. 18, a shape and a bulge 355 converts the line contact into an almost surface contact in order to achieve greater load-bearing capacity and / or protective layer plates, sheets, sheets, 309 and / or sealants 356 are arranged; (Figure 18.)
• or; that a first cylinder 301 is arranged, in which a first piston 302 is reciprocally located, further a second cylinder 304 is arranged, in which a second piston 305 reciprocates and the first cylinder is connected to the second cylinder by a connecting channel 303, the cylinders and the channel are filled with liquid and the reciprocating movement of the first piston is carried over to the second piston through the liquid column in the channel and the cylinders; (Figure / 2.)
• or; that the first piston 302 has a small diameter, the second piston 305 3 has a larger diameter, as a result of which the compressive force of the second piston becomes greater than that of the first piston and the movement of the second piston takes place in reverse parallel to the first piston in that the second piston runs into the second cylinder, if the first piston runs outwards in the first cylinder and vice versa (vice versa);
• or; a controller 326, 329, 328, 327 for correct filling of the named fluid column or liquid column in the rooms, cylinder core, channels, 301, 303, 304 is arranged, which at the time of low pressure, preferably in the outer piston position or the vicinity of it, of the first piston 302, the cylinder indenter 301 and 304 and the connecting = channel 303 with fluid from the supply channel, for example 329 out, so that a perfect transmission of the movement of the first piston 302 to the second piston 305 takes place;
• or; that the piston stroke of the first piston 302 is arranged in an infinitely variable manner and the piston stroke of the second piston 305 is thereby also infinitely variable;
• or; that the element 1, 2, 307 etc. has an outer diameter and an inner diameter, the radii of which are half of the diameter in question and the difference between the outer radius and the inner radius is at least three times smaller than the outer radius. (Figure 12); (Made from Figures 13, 14, 15, 19, 20.)
• or; that the first piston 302 is provided with a piston crown drive device, for example 334, 347, 336, 346, 344, 345 or the like, which forces the stroke of the first piston 302; (Figure / 2).
• or; that only one element 1 is arranged between a retractable piston 40 and a head 48; (Figure 19.)
• or; that in order to achieve a maximum service life of the elements, the internal stresses in the elements 1, 2, 307, etc. are matched to one another by compressing and deflecting them under internal fluid pressure from the working chamber 61. (Figures 15, 20).

FIG. 16 shows a schematic diagram of a work system for a solid material internal combustion engine. The pumping units of the invention can advantageously be used therein. In the tank (for example the tank of the vehicle) 806 there is the solid fuel 807, which is not a liquid and not a gas, but a solid substance, for example powder, magnesium, coal or cleaned and pressed coal or the like. The first transport device 809 serves to constantly transport the solid fuel to the tank outlet and fuel inlet 805, 804. The second transport device 805 conveys the solid fuel through the inlet 804 into the combustion chamber or pre-combustion chamber 800. There, the solid fuel may encounter a pump of the invention, which is indicated in principle in 808 and which directs a fluid jet (jet), for example a water jet 802, through the jet nozzle 803 at the incoming solid fuel 801. The fluid jet jet 802 hits the solid 801 with great force of a few thousand atmospheric back pressure and crushes the solid fuel into dust. This results in the mixed stream 810 of solid fuel dust intermingled with the fluid from the fluid jet stream 802. This mixed stream is extracted in the combustion chamber 800 and burns in it in compressed or other air. If the fluid jet stream 802 is a liquid fuel, then the liquid fuel and the solid fuel dust in the combustion chamber 800 burn together. If the jet stream 802 is a high pressure water jet stream, the droplets of water in the flame in the combustion chamber 800 evaporate, while the dust parts of the solid fuel stream 801 burn in the air and convert the water of the jet into steam. The steam then participates in the expansion of the gas together with the air heating up during the combustion, which supplies the drive for the pistons or expansion devices of the combustion engine. A particularly elegant solution is, for example, to strongly press the solid fuel into a right-angled cross-section 815, for example to a carbon filament of any cross-section per se, although the rectangular or square cross-section 815 is preferred. Because the rectangle or square cross section can be rolled up into a spiral (radial spiral), as shown by 25. With such a spiral, the tank 806 can be filled completely. Then, since a cubic millimeter of pressed coal has a little more than twice the calorific value of a cubic millimeter of gasoline, you can drive almost twice as far with a tank of the same size if you use the coal or solid fuel according to the invention instead of gasoline or diesel fuel. The second conveying device 805 rolls the fuel spiral 825 continuously and conveys the solid fuel filament 835, for example of the cross section 815, rapidly and continuously through the fuel introducing nozzle 804 as the filament 801 Brennkommer 800 into where the fuel filament 835,801 then meets the fluid jet stream 802 and is crushed by it under the high back pressure into powder or dust and entrained with the GluidJef stream 802.

It is particularly preferred to flow a hot flame in the combustion chamber 820, for example from hydrogen, gasoline or the like, and air through the flame inlet 824 as a flame constant current 821 into the combustion arm 820 and thus within or around the radial center or axis to produce very hot continuous flame 831 and to maintain it continuously. The solid fuel filament 835 then moves continuously and continuously, with the speed in the car being controlled with the known accelerator pedal, through inlet guide 804 into the combustion chamber, pointing to the fluid stream Jet, e.g. the water jet stream hits from its nozzle 823. At point 822, the fuel filament and the fluid jet, for example the water jet stream, meet, the solid fuel filament being crushed into powder or dust and entrained with the fluid jet stream, directly into the hot flame 831 that flows through the line 824 is streamed in.

Since the flame 831 is kept very hot by this arrangement, it has enough calorific value to immediately convert the water of the jet stream from 823 into steam, that is to say vaporize it, without the water jet. could kill the flame 831. As the flow of the mixture of air, steam and fuel gas continues along the arrows within the combustion chamber 820, the fuel gas is preferably further swirled by means of the rotating blades 826 on the rotor shaft 825, and this rotor arrangement 825, 826 also serves to serve as a source of foreign material = fling solid or unburned solid particles radially outwards and collect them in collecting chambers in the radially outer area of the combustion chamber 820. The fuel gas that has become so perfect with at least partially achieved expansion then emerges from the outlet 828 from the combustion chamber in order to enter the chamber (s) of the expansion device of the internal combustion engine, for example the expansion stroke cylinder of the engine and the pistons therein or to drive the piston in the expansion stroke, i.e. in the work delivery hub.

Figures 21 and 22 show matching cuts along the relevant center lines of the figures of a piston-piston arrangement with parts of the sl _' e surrounding surrounding parts. This arrangement may be used, for example, in the motors or pumps of the application, or may serve in general for pumps and motors which are used outside the other units in the other figures of this patent application. The task of these figures and FIG. 23 is to create the largest possible, or more precisely, the largest possible contact surface of the swivel joint between a piston with a piston shoe which can be pivoted thereon. This is necessary in order to achieve particularly high pressures in the unit. So high pressures of many hundreds or at least three hundred bar, which are occasionally necessary. In the aggregates of the other figures one occasionally uses 500 or 1000 bar for the pistons and shoes. This high pressure can no longer be carried in the swivel joints of DE-OS 30 41 367 because the contact surface between the piston and the piston shoe is too small therein. According to the invention, the piston 501 receives the bearing surface 510 with the wheel 508 around the pivot axis 505. It is now a partial cylinder surface that extends over the entire cross section of the piston. The piston shoe is provided with a complementary oscillating surface 511, which rests on the piston bed and has the radius 509, which corresponds to the radius 509. The piston shoe bearing on the piston is provided with pressure flap pockets 515 fed by channels 516, 517, and these are also arranged in the outer surface of the piston shoe with which it slides on the piston stroke guide 514. The piston stroke guide 514 has the calf stroke guide surface 512, on which the outer sliding surface 513 of the piston shoe runs. The swivel bed surface of the piston with the radius 508 about the swivel axis is designated by 510 and the complementary cylindrical hollow part swivel surface of the piston = shoe is the swivel surface 511.

• a) Eine groesstmoegliche Lagerflaeche des Schwenkgelenkes ;
• b) Ein Eintauchen der Stifthalterung in den Zylinder, und
• c) Herausziehen des Kolbens beim Saughub mittels der lugflaechen wodurch gleichzeitig die folgenden Vorteile erzielt werden :
• d) . die Anordnung ermoeglicht hoechste Drucke infolge der grossen Tragfaehigkeit der Schwenklagerflaechen ;
• e) der Kolbenhub bleibt gross, da ein Teile der Verbindungsonordnung zwischen Kolben und Kolbenschuh in den be reffenden Zylinder eintauchen kann ; und;
• f) einen einwandfreien Selbstansaugehub, da ausgedehnte und praezise wi rkende Zu gflaechen fuer das Herausziehen des Kolbens in Verbindung mit dem Gelenkbolzen beim Saughub gewaehrleistet sind.

This arrangement, like that of FIG. 23, should also have the advantage that a relatively large, long piston stroke can also be achieved, so that the pivot axis can enter the cylinder 503 of the unit. The piston axis is 504, the piston head 533, the swivel axis of the swivel joint between the piston and piston shoe has the number 505. In order to achieve the described aim, the piston has a transverse bore 507 with the swivel axis 505 as the center line, which is perpendicular to the piston axis 504. The mounting bolt 506 is inserted into it. In addition, the Holben 501 has at the outer ends of said bore recesses 518, which are radially widened by the bore ₅ 07 and the bolus 506. The ends of the bolt 506 enter these recesses of the invention, 518, but are the bores 506 in the transverse bore 507 are shorter than the diameter of the piston 501 and the cylinder 503, so that the pins 506 do not abut the cylinder walls and cannot damage them. Retaining grooves 520 of the intermediate brackets 519 are inserted into the recesses 518 and they grip around the described ends of the bolts 50 ^6. Radially outside of the pivot axis 505, the intermediate parts 519 protrude axially and form the tensile surfaces 531 there, while the intermediate parts are otherwise radial of the piston are extended outwards until they meet and rest on the bearing surface 511 of the piston shoe. Axially outside of the intermediate parts 519 are the outer parts 5, which rest with their radially outer seats on the piston stroke = guide surface 512 of the piston stroke guide, and the outer parts 530 form the encompassing tensile members and encompassing tensile surfaces 532, which pull the surfaces 531 of the intermediate parts 519 reach around, touch them and let them slide on them. The end parts 530 thus serve as traction parts, which pull the piston-piston shoe arrangement radially out of the cylinders 503 during the intake stroke. The inventive arrangement according to FIGS. 21 and 22 thus simultaneously achieves:

• a) The largest possible bearing surface of the swivel joint;
• b) immersing the pen holder in the cylinder, and
• c) Pulling out the piston during the suction stroke by means of the lye surfaces, which at the same time achieves the following advantages:
• d). the arrangement enables high pressures due to the large load-bearing capacity of the swivel bearing surfaces;
• e) the piston stroke remains large, since a part of the connection arrangement between the piston and the piston shoe can be immersed in the reffing cylinder; and;
• f) a flawless self-priming stroke, since extensive and precisely working surfaces are ensured for pulling out the piston in connection with the hinge pin during the suction stroke.

In the figure, 23 is a part of the structure of the figures is changed 21 and 22, Wodu rch one of the parts can be saved and be allowed esungen another production method for diselben tasks and L _o. Piston, rotor 508, cylinder and piston stroke guide are the same as in FIGS. 31 and 32. The pin 506 and the recesses 518 are also the same. The parts 519 and 53 ₀ of FIGS. 21 and 22 are, however, replaced by the brackets 529, the eyes of which grip around the piston ends of the bolts 506 and which engage again in the recesses 518 of the piston 501. What differs from FIGS. 21 and 22 is that the holding parts 29 are provided with outer grips, rims, 522 ₁ . On the outside at the axially outer ends, the piston piston has the outer dimensions which form the holding surfaces 5-3, on which the ribs 522 with their holding surfaces 524 lie. Thus, the rims 522 of the parts 529 hold the piston 501 and the piston shoe 502 together and enable the shoe 502 to pivot on the piston 501. The entire arrangement can be seen in FIGS. 31 and 32, as well as in FIG the mutually facing radially planar inner surfaces 527 of the retaining rings 525 are kept nested between bearings 528 and are secured against axial displacement by storing and holding the mentioned inner surfaces 527 of the rings525 against the radially flat axial end surfaces 526 of the connecting parts 529.

Figure 24 shows the laen. Cut through a hydraulic unit that can be used as a pump or motor for high pressures. Very high pressures are required in the units of the invention, which the units with roller bearings can usually no longer meet for a sufficiently long service life. The unit of FIG. 24 with its sub-figures 25 and 26 can therefore be used particularly for the units of the invention as a pump or as a motor, but also generally as a hydraulic unit, for example around excavator and press units, as can be used in general hydrostatics. The main shaft 443 has a center L portion 401 which extends radially outward and forms an outer surface, preferably a cylindrical shape, which is used to place a hydrostatic bearing around the central portion which exposes the high radial forces from the cylinders 439,440 of the two axi als of the central part 401 arranged chamber groups 439,440. On the other hand, the shaft is mainly supported for centering purposes in the end bearings 445 and 444, the bearings 445 centering the shaft axially and the bearing 444 permitting axial expansion of the shaft under heat. On both ends axially of the central part 401, a body 441, 442 is arranged on the shaft 443, each containing a rotor or chambers, which the working chambers, e.g. cylinders 439,440 of chamber groups 439,440. In the middle part, the pressure chambers 402, 403 are also arranged, in which axially movable pressure bodies 404, 405 are arranged, which, under the pressure in the chambers 402, 403, press the rotor in question with its other end surface, the rotating control bottle or rotor surface against the control surface arranged in the housing 447 . The body 404 the rotor 441 against the upper control surface in the figure, the pistons 405 the rotor 442 against the lower stationary control surface in the figure 24.

As far as we have already tried to design such units, they have encountered difficulties in that they did not work properly or did not work for all applications. By the present invention nor the 25 through 26, the emerged difficulties been overcome and quietest _t ungsfaehige aggregates with high operational reliability created for wide application areas. For example, pressure chambers 402 and 403 are assigned pressure fluid pockets in the stationary control or storage areas at the other end of the unit. This prevents the control surfaces from overheating.

According to the invention, a channel 409 to a secondary pressure chamber 410 is arranged from the respective pressure chambers 403, the secondary pressure chamber 410 being open in the opposite direction from the rotor, like the pressure chamber 403.A channel 413 is correspondingly assigned to the pressure chamber 402, which leads into the secondary pressure chamber 412 leads, which is again open in the opposite direction from the center plate 401, like the pressure chamber 402. In the secondary pressure chambers 410 and 412, the secondary pressure pistons 611 or 411 are arranged, provided with a passage and they press on the respective adjacent rotor 441 , or 442, seal against this and open into the relevant rotor secondary duct 433 or 618, which passes axially through the rotor 441 or 442 in question and at the other end of the rotor 441 or 442 in question into the relevant pressure fluid pocket 432, 420, 434 or 435 the relevant stationary tax = area of the aggregate needs. As a result, pressure fluid is directed from chamber 403 into pocket 432 or 420, depending on the position in the rotor circulation of the pressure chamber 403 gorado in question. Correspondingly, pressure fluid is directed from the relevant pressure chamber 402 into the relevant pressure fluid bottle 435 or 434 of the control surface of the unit on the other side. What is achieved by this arrangement according to the invention is that the respective rotor 441 or 442 between its own rotor channel 421, 424, its own control port 423, 426 is pressurized with its own cylinder pressure and, at the same time, is also ainsial bcawl with the pressure from the pressure chamber of the other rotor the pressure fluid pocket concerned 432,420,434 or 435, it will daduech prevents one of the rotors or both rotors mmerngruppen by strong pressure in one of the K _a 441 or 442 402 or 403 and low pressure in the other of the working chamber group is pressed too strong and heisslaufon the Steuerflaechen . For example, the channels 421 and control pocket 423 have the pressure from chamber 4-39, while the Orackfluid pocket 432 at this time has the pressure from the other working 440.

The same applies for the same premises 424,427,440,420, and he _fu 424,426,440,432 and 439,421,423,434 424,417,440,420 or so forth. The previous overheating of the control surfaces, especially with different pressures in the working chambers 439 or 440, has been avoided and the unit is working now, with the correct dimensioning and position of the parts described, with appropriate relative speeds and pressures.

E. Also for reasons of operational safety, the storage of the middle part has been redesigned according to the invention. Thus, the middle part 401 fits tightly around it, the inner race 460, which runs in the slide bearing ring 46 / the housing 462. Between the inner race 460 and the middle part 401 there are 180 ° around the middle part 401 the neck spiral channels 406, blw. 408 and which are shown in the development in FIG. 25. There you will also find the mouths 414, which extend radially outwards from the secondary chamber channel 409 through the inner race 460 into the relevant pressure fluid pocket 407 in the inner race 460. From the respective secondary pressure chamber 412 or the secondary channel 413, the pressure therefrom is passed through the radial channel 414 into the relevant pressure fluid pocket in the radial outer surface of the inner race 460. As can be seen in FIG. 1, a pressure fluid storage pocket 407 is connected to a pressure chamber 403 and a working chamber 440 by means of channel 415 via channel 406, while the respective pressurized fluid storage pocket 477 is connected via a channel 414 or 415 and a semi-spiral channel 406 with a pressure - Chamber 402 and a cylinder 439 is connected. As a result, the pressure from the relevant cylinder 439 or 440 is passed through 180 degrees around the central part 401 into a diametrically opposite pressure fluid pocket 409 or 477. In the running surface of the inner race 460, which rotates in the bearing ring 461, one has to alternate a pressurized fluid bearing carrier bag with pressure from a diametrically opposed cylinder 439 and an adjacent one with pressure from a diametrically opposed cylinder 440. The rotor, or the middle part 401 Therefore floats radially between the forces from the cylinders 439 and 440 on the shaft and the oppositely directed forces from the pressure fluid pockets 407 and 477 on the middle part 401. The radial loads on the shaft, rotors and middle piece cancel each other out and the bearings 444,445 can be weak be. The unit, however, allows very high pressures without the bearings 444,445 failing under pressure = load. The goal according to the invention has thus been achieved.

FIG. 26 still shows the important knowledge of the invention that the pressure chambers 401, 403 must lie on the radius "rgc", which must be calculated and found according to the formula shown in the figure Control surfaces of the unit are still hot. And accordingly the diameter of the pressure chambers 402, biw, 403 must correspond to the equation for "dth" according to FIG. 26 if the control surfaces are not to run hot. These equations are therefore important and new conditions and knowledge within the scope of the invention according to FIGS. 24 to 26. In the right half of FIG. 26 one can see the control pocket 422 and the pressure fluid pocket 420, the relief groove 458 between the control mirrors around 422 and 420 , as well as the radii "Ro" and "Ri" important for the calculation according to the equations. It is also important that for most applications the control levels around 422 and 420 should give the same pressure and the secondary pressure chambers should be on the "Rgc" radius of the outside control level around 420. The other three control mirror half pairs around 423, 432-424, 435-426, 434 are designed accordingly. It can also be seen from FIG. 25 that it is sometimes expedient to separate the pressure fluid pockets and their sealing surfaces, that is to say the pressure fluid pocket zones 407 and 477, from one another by means of drainage or separation channels or grooves 453, in order to simplify and more precisely calculate the Storage and pressure fluid pockets 407,477 with their sealing surface 452 that surround them.

Further details on the figures of the patent application can be found in the partly very extensive and partly also expensive Rotary Engine Kenkyusho reports, the RER reports.

und fuer das Russenmap un 57 gilt der Durchmesser "d_ta" nach der folgenden Berechnung:

mit : fb = Balanzierungs- Faktor; Pi = 3,14; E = Druckzone. ( Warum das "G" in der Gleichung ist, hat der Erfinder z.Zt. vergessen.)The calculation _formula applies to the radius "r _gc ";

and for the Russian map un 57 the diameter "d _ta " applies according to the following calculation:

with: fb = balancing factor; Pi = 3.14; E = pressure zone. (The inventor has currently forgotten why the "G" is in the equation.)

The numbers 459,480,482 and 483 _' show the reversing bends between the relevant tax pockets of the relevant tax index.

For the evaluation of Figure 14 in the practical construction of the unit, the following calculations must be observed:

For the practical implementation of the unit of the invention, the following calculations are of importance and must be observed for FIG. 19:

In Figures 27 to 3 ₀ the piston Schun 502 is connected in similarity = Licher manner with the piston 501, as shown in Figures 21 to 23. The reference numeral 501,502,504,505,506,507,508,509,510, 511,513,515,516 and 517 show in principle the same parts as in the figures 21 to 23, Namely the piston, the piston shoe, the piston axis, the swivel bed axis, the cross pin, the cross hole, the two radii, the swivel bearing bed surface, the swivel surface, the outer surface of the piston shoe, the pressure fluid pockets and the channels through the piston to the pressure fluid pockets, whereby these Parts compared to those of Figures 21 to 23 may have had structural changes if necessary.

While in FIGS. 21 to 23 the holding arrangement for the piston and the piston shoe was possible by means of parts arranged on the rotor, FIGS. 27 to 30 show a self-retaining connection of the piston shoe to the piston. For this purpose, the intermediate holder pieces 519 of FIGS. 2 to 23 are replaced by the holder pieces 619 in FIGS. 27 to 28. But the pieces 519 and 619 serve the same purpose, namely to connect the piston 501 and the piston shoe 502 so that they can pivot relative to one another. Only the eye 520 is in the recess 518, the parts 518 and 520 being practically identical to those of FIGS. 21 to 23, so attached to the transverse bolt 506 that the pivoting movement cannot be prevented, the middle pieces or intermediate pieces 619 but are still held firmly on the cross pin 506 and thus the connection between piston 501 and piston 502 remains secured. Correspondingly, within the recesses 518, the eyes 520 grip the ends 606 of the bolt 506, which are kept smaller in diameter, held at the thicker central part of the bolt 506 and firmly connected at their axially outer ends by means of holders 570 to the bolt 506. Once this arrangement has been carried out, the connection and permanent pivoting mounting and mounting of the pivoting piston shoe on the piston no longer require any further aids. Instead of gripping the piston shoe as in FIGS. 21 to 23 _j , the gripping can also be carried out as in FIGS. 27 to 30, in that the central parts 685 of the pieces 619 are provided with bunches 574 radially on the outside engage in recesses 573 of the piston shoe 502 and advantageously use collars 574 which are provided with a holding surface 572 with a radius 571 about the pivot axis 505 in order to fit on a complementary surface before storing on Kolbenschun 502 and holding it. In comparison to the exemplary embodiment in FIGS. 21 to 23, it can also be seen in FIG. 27 that the piston 501 can be provided with the cross head 575. The bed surface 510 extends in Figure 27 over the entire length of the crosshead. The crosshead can be designed with the radius 508 around the pivot axis 505 so far parallel to the pivot axis 505 that the axial ends of the crosshead 575 together with the axial end parts of the bearing bed surface 510 protrude beyond the outer diameter 533 of the piston 501. With such an arrangement, a larger area of the bearing support surface 510 is obtained, so that the complementary piston shoe 502 mounted on it can transmit correspondingly high radial forces without swiveling the swivel surface 511 on the bearing bed surface 510.

mit "Kf = Fliehkraft in Kg; die betreffenden Radien ( Rz= aeusserer und Ri = innerer) der innersten und aeussersten Teile des Kolbens 501The 27 to 3 ₀ further comprising a further recognizes = nis show the invention and a Lo solution to the information thus detected problems. When the rotor of a pump or a motor rotates, the pistons 501 and piston shoes 502 sliding in the radially arranged cylinders 503 are subject to the centrifugal force. The centrifugal force is:

with "Kf = centrifugal force in kg; the relevant radii (Rz = outer and Ri = inner) of the innermost and outermost parts of the piston 501

in meters, if it had the same diameter throughout from ra = diot inside to outside; "ϑ" = specific weight of the material of the piston; "ϑ" = (ϑs / g) = mass of the piston shoe with "Rs" = center of gravity of the piston shoe and "ϑs' its specific weight; furthermore with pi =" π "= 3.14" d _p "= outer diameter of the piston 501 ( see Figure 51) and "ω" = the angular velocity = 0.1047x rpm.

mit " e " = Exzentrizitaet = Abstand der Achse des Rotors von der Achse der Kolbenhubfuehrungsflaeche,z.B.513, siehe zum Beispiel der Wert "e" auch aus der europaeischen Patentanmeldung 0 064 563.The radii are not constants, however, they change with the circumferential angle "α" of the rotor 608, namely the radius of the distance between the swivel axis "Rc" and the rotor axis according to Eickmann technology from older Eickmann patents:

with "e" = eccentricity = distance of the axis of the rotor from the axis of the piston stroke guide surface, eg 513, see, for example, the value "e" also from European patent application 0 064 563.

If the calculation of the defective individual sections of the piston 501 concerned can also use the values “R3”, “R4”, “Rc” etc. from FIG. 28. These are then summed up in order to maintain the total effective centrifugal force.

From the European patent application of the applicant and inventor just mentioned it emerges that for a high performance in a small space, that is to say with little external wear on the unit, the piston stroke "S" in comparison to the radius or inner diameter "da" of the piston stroke guide surface 513 is as large as possible have to be. Since the piston stroke is "S" = "2 e", the eccentricity written in equation (28) must be as large as possible in order to obtain high performance with low weight and a small external dimension of the unit in question. The consequence of this is a high fluctuation with a high serving height of the current value of the relevant radius "R" around the zero value (mean value at e = 0) "Ro" of the relevant radius. The consequence of this is a large swiveling of the centrifugal force around the mean value of the centrifugal force "Kfo". This fluctuation was present in previous radial piston units, but it was only so high in Eickmann units. When using other non-Eickmann units, the value of the angular velocity "ω" did not play a major role, since these units were mostly driven by electric motors at the same speed. Because as engines with variable speed, units without the deep-diving piston shoes of the Eickmann system are not efficient and therefore hardly used. In the current units of the invention, however, the units are often driven by hydraulic motors and the speed is regulated by steplessly adjustable pumps. Therefore, in the current radial piston units, speed differences from O rpm to 10,000 rpm can occur. If one now looks at equations (27) and (28), one finds that, for example, at 10,000 rpm the angular velocity is 1000 times as high as at 10 rpm and I follow the square of the angular velocity and consequently according to the equations mentioned the centrifugal force of the piston 501 with the piston shoe 502 at 10,000 rpm must be one million times as great as at 10 rpm.

For centrifugal force fluctuations of millions and amplitude fluctuations of over 20 percent, however, the previous outer surfaces of the piston shoes, for example according to the patent DE - PS 2 500 779, are no longer sufficient. New means must be used to remove the treads 512 of the piston shot in question hes 502 usable for a sufficiently large speed range without the surfaces 512 and 513 overheating under excessive force fluctuations. Although it is generally not possible to make the piston shoe of the radial piston machine usable for all speeds from 0 to 20,000 rpm, it is possible to use the measures of the current inventive features to paint the area between a highest and a lowest D. _r eh = to expand the number of applications.

This is achieved according to the invention, the 29 to 3 ₀ and the figures 44 to 46 in the piston shoe to be worked into 502 Tragflaechenstege 579 in the pressure pocket 515 that either the pressure = fluid pocket 515 of the Aussenflaeche 512, or the front part and Rueckteil 580 the relevant outer surface 512 of the piston shoe 502 beyond the drain grooves 577 can be provided with means for enabling larger speed ranges, or both measures can be arranged simultaneously.

The arrangement of the Tragstegflaechen 579 in the figure 3 ₀ has the advantage that these surfaces are lubricated from both ends of the running direction with pressure from the respective adjacent printing fluid pocket parts 515th This gives these surfaces 579 a higher load-bearing capacity per surface cross-section than the sealing surfaces (Sea = ling lands) 578, because the sealing lands 578 are only pressurized with fluid at one end and are therefore less lubricated, so that they also have less load-bearing capacity per surface cross-section , as the flat parts 579. It is expedient to make the webs 579 very short in the air direction, for example to arrange 1 to 5 mm and for this purpose to arrange many webs 579 because the forced lubrication from the pocket parts 515 only acts a few millimeters deep. If the webs 579 were too wide in the direction of travel, the forced lubrication would fail and the running surface 512 would run hot when the speed became high.

For the exact calculation, the area "L x W" in FIG. 30 is assumed to be the high-pressure zone and, according to the invention, this is made larger than the piston cross-section d ² pi / 4 of the piston 501. In practice, the dimension "L x W" is approximately the sum of the centrifugal force Kf divided by the difference between "L x W" and d ² pi / 4 and the difference mentioned multiplied by the fluid pressure in the pressure fluid pocket 515 and the relevant cylinder 502 from which this pressure acts on the piston crown 533

The corresponding feature of FIG. 30 therefore consists in making the dimension “LxW” larger than the piston cross section d ² pi / 4, if the unit is to run at a correspondingly high speed, the support web surfaces 579 inside the pressure fluid pocket 512 transversely to the air direction to be arranged, to clearly delimit the sealing surfaces 578 by means of the drain grooves 6 and to arrange the stabilizing surface parts 580 arranged in front of and behind the relevant drain groove 577 in the direction of the lay and, if appropriate, in the direction transverse to the direction of the run, as viewed in the direction of travel. The surface parts 580 are then, as will be described with reference to FIGS. 54-56, converted from a support surface into a load-bearing surface, so that they can contribute to the absorption of the centrifugal force of the piston 501 and the piston shoe 502.

The figures just described still show the important radius "da" as the radius of the tread = piston stroke guide = shaft diameter "da" as number 901, and the swivel joint = distance from the rotor axis as radius "Rc" = 902 and the direction 903 of the piston shoe relative to piston stroke guide surface 514.

In FIGS. 31 and 32 parts with references already known from the other figures, reference number numbers show parts with the same functions or the same parts as in the other figures, in which the same numbers appear. FIGS. 33 to 38 show the parts of FIGS. 31 and 32 in a single part = separate representation in order to be able to see better the radii and surfaces in question and the shape of the parts in question.

The special features of FIGS. 31 to 32 are that the piston shoe 502 is pivotally held on the head of the piston 501 by means of holdings 581 fastened to it. The brackets 581 can be fastened to the piston shoe 502 with fastenings 582. To hold the piston shoe 502 on the piston 501, a holding surface 583 is formed on the piston head and the holder 581 has a collar 584, the holding surface 586 of which is placed on the holding surface 583 of the piston 501 in order to partially grip it. These arrangements take place at each axial end of the piston head and parallel to the pivot axis 505. The surfaces 583 and 586 receive the only slightly different radii 587 around the pivot axis 505.

A further peculiarity of these figures is that the piston 501 ver on its head in the axial direction parallel to the pivot axis 505 beyond the outer surface 637 = piston diameter "dp" longer piston cross heads 588, over which the swivel surface 510 is extended, that is, a continuous contact surface = swivel = bed surface 510 over the entire piston head of the piston 501 formed from the piston cross section and piston cross = head.

The wing webs 579 can also be arranged on one of the surfaces 510 or 511 between the piston 501 and the piston shoe 502. Corresponding channels 716 can conduct fluid from the relevant pressure fluid pocket 515 into those between the surfaces 583 and 586, as well as connect the channels 616 pressure fluid pockets 515 through the piston shoe or through the piston 501 to the relevant cylinder 503 of the unit. The surfaces 583 and 586 described have approximately the same radii 589 about the pivot axis 505. If the fastening 582 is a screw, the threads 590 are arranged in the piston shoe 502 and, in addition, either the piston shoe 502 or the holders 581 can pull the tension surfaces 531 with the radii 592 around the axis 592 of the guide surface 513 of the piston stroke guide 514.

In FIGS. 39 to 53 too, the reference numbers already described, referentials, have the same meaning and show the same functions or corresponding parts as in the figures in which the reference number numbers have already been described.

The special feature of the invention in FIGS. 39 and 40 is that the piston 501 is provided near its upper end with a transverse bore 594, the axis 505 of which forms the swivel axis, passes through the axis 504 of the piston 501 and, as well in the other relevant figures, perpendicular to this. The transverse bolt 595 is inserted into the transverse bore 594 and carries the piston shoes 1102 and 1202 at its axial ends. These are provided with a bore into which the ends of the transverse bolt 595 fit so that the piston shoes about the axis 505 of the bolt = can pivot 595, or pivot the pin 595 with the piston shoes 1102 and 1202 about the axis 505, the center piece of the pin 595 then pivoting in the transverse bore 594 of the piston 501. The piston shoes then have the already known outer surfaces 512 with the pressure fluid pockets 515 and the other accessories already known from the other figures. Figures 39 and 40 therefore use two piston shoes 1102 and 1202 instead of the previous ei = nem piston shoe 502. The manufacture of such piston shoe 1102 or 1202 is a particularly simple one because it can be rotated out of rings and the cross hole can be drilled to receive the ends of the cross pin 595. However, if the design is to absorb particularly high radial forces or allow particularly high speeds, then it is also expedient according to the invention to also use the piston head for the radial support of piston shoe parts. Then the piston head of the piston 501 receives, in addition to the transverse bore 594, the bearing bed surface 510, as is known with the radius 509 around the swivel axis 505. The piston shoes 1102 and 1202 then receive the piston shoe axial extensions 1302 and 1402, which have the swivel surfaces, facing one another 511 are provided and store and swivel on the swivel bed surfaces 510 of the piston 501. Corresponding pressure fluid pockets and connecting lines 515 and 515 can be arranged in parts of the piston shoe, the piston head and / or the transverse pin 595 and are shown in the drawings at examples of places. FIG. 40 shows partial sections along one or more of the arrows A, B or C of FIG. 3 ₉ , the sections being laid out and the cross-sectional parts being shown in FIGS. 4 through _{0 in} such a way that the important arrangements can be seen. Parts that are not specifically described here are those with reference number that are already known from the description of other figures.

Pull rings 731 can be placed under the pull surfaces 531 in order to pull the piston shoes radially outwards during the suction stroke, for example in excavator pumps, and thereby also pull the pistons 501 connected to the piston shoes radially outwards in the cylinders 503.

The swivel surfaces 511 of the piston shoe axial extensions 1302 and 1402 can merge into angled surface parts 597 in order to simplify the manufacture and to allow perfect swiveling of the relevant piston shoe parts on the bearing bed surface 510 of the piston 501. The pull surfaces 531 run on the guide surfaces 931 of the pull rings 731.

The piston shoe 502 of FIGS 41 to 43 are equivalent to the in its principle known from the German Patent 2,500,779 and aggregates in German Patent 1,302,469, the G _r undpatent for deep plunging piston shoes usable.

But the piston shoes of the German patent 2 500 779 are not suitable for speeds that require many thousands of revolutions per minute. In addition, they lack stability. According to the invention, new arrangements are therefore made in order to make the deep-diving piston shoe even more stable, so that it lies more precisely with its running surface 512 on the piston stroke guide surface 513 of the piston stroke guide 514 in question, as a result of which its outer surface parts come into effect better and achieve less leakage and friction , and so that the piston shoe can also be used for high speeds without running hot. The new arrangements consist in the fact that the piston shoe guide parts 1502 are designed in the peripheral direction, that is to say in the circumferential direction, as long as possible, so that a long contact surface of high systems = stability and with a large cross section is created. In addition, a distance "B" is observed, which is calculated from the inner edge of the guide parts 1502 to the pivot axis and which also determines the distance "G" de pivot axis 505 from the outer surface = running surface = sliding surface 512 of the piston shoe. These two distances "B" and "" G "are now too short, as possible. In the piston shoe for about 40 cc with 7 pistons, for example" Bxcirca 2 mm "and" G = approx. 8 mm ". The peripheral length is "L". These data, namely "B", "S", "Rs", "G", "L" are shown in the figures, since = with the reader of the patent applications on inquiries for corresponding = de butt sizes exactly are given answers of the invention may be noted the equation for D _r uckfluidtaschen Further, must accordingly, namely, equation (29), which also appears in the figure.:

The pressure fluid pockets 515 are now partially radially above the cross section of the piston 501 and the slots 702 between the surfaces or sliding pieces or legs of the "H" of the deep diving = the piston shoe, i.e. the parts 602 and 602, are now, according to the invention, axially shorter than in German patent 2,500,779.

The piston rod guide parts 602 are thus located on the piston rod center part 302, which are separated from one another by the slots 702 at both ends of the piston rod center part, so that the piston arms of the piston 502 and the radial webs of the rotor can enter them. The piston arms and the radial webs are known from the cited patents. The running direction of the piston shoe = outside = surface 512 is designated with: "Directions" Mf "and" Mo "of for = ward and oppositional movements" in FIG. 53. The width of the slot 702 is designated with "S" and the pressure fluid pockets 515 have the cross sections "fp1" and "fp2" together with half of the surfaces of their sealing lands, that is to say their sealing surfaces 578. The central part 302 of the piston shoe carries the swivel roller 902 with the swivel surface 911, which extends over the entire swivel roller 902 extends and pivots and supports in the swivel bed 510 of the piston 501 in question. The swivel roller 902 extends axially, that is to say para = Ilel to the swivel axis 505, beyond the middle piece 302 and is there in one piece with the slide pieces 602 and supports them. As a result of the large length "L" of the new piston shoe, a larger number of support surfaces 999 are arranged on the side of the drain grooves 577, which are separated from one another by drain slots 577. The larger number of drainage grooves promotes the load-bearing capacity of the several support surface webs 999. Current tests have shown the load-bearing capacities of the piston shoes for different speed ranges and it is also apparent from which radial forces the surfaces 999, 578 or 579 at which pressures and carry speeds.

FIGS. 44 to 46 explain in particular which measures according to the invention can be arranged in order to make the piston shoe of one or the other type of this patent application or the general piston shoe usable for extended speed ranges. FIG. 44 shows the arrangement of a large number of support surfaces 999 beyond the inner drainage grooves 577 and their interruption by further drainage grooves 577. Care must be taken to ensure that the grooves 577 are properly filled with lubricating fluid if the unit is to work gurly. For this purpose it is often advisable to provide the support = partial surfaces 999 with lubricating fluid influence grooves 1577, which promote the influence of lubricating fluid in the running direction of the piston shoe. These are advantageously laterally offset from one another in different adjacent pieces 999, so that the grooves 577 to be filled well.

um den Durchmesser der gedachten Welle zu erhalten. In der Gleichung (30) bedeuten : "fv" = Berichtigungsfaktor, "pi" = " F " und "dv" = den Durchmesser der gedachten Welle. Der Berichtigungsfaktor "fv" liegt zwischen 1 und 2, naeher an 2. Hat man so den Durchmesser der Welle, dann kann man, wenn man die Relativgeschwindigkeit zwischen den Flaechen 513 und 580 gleich der Umfangsgeschwindigkeit der Welle mit dem Durchmesser "dv" gleichsetzt, die hydrodynamische Tragkraft des Passungsspal tes zwischen den Flaechenteilen 513 und 580 zum Beispiel nach der Huette oder nach dem Buche : "Die Grundlagen der Lagerschmie= rung" von Werner H.Kara (Erdoelbuecherei) berechnen. Daraus laesst sich der notwendige Abstand zwischen den Flaechen 513 und 580 vom Beginn an der Nut 577 zu berechnen und wenn man fuer jeden Platz der Flaeche 580 diesen Abstand hat, dann kann man die Kruemmung der Flaeche 580 berechnen und damit ihren Radius oder ihre Radien und ih= re Mittelachse oder ihre Mittelachsen. Wird das richtig ausgefuehrt, dann kann die betreffende Flaeche 580 so ausgebildet werden, dass sie eine maximale Tragkraft fuer den gewuenschten Drehzahlbereich erhaelt. Man hat also beim Kolbenschuh der Figuren 45-46 einen hydrostatischen Tragteil zwischen den Nuten 577, der den Hauptteil der Radiallast traegt und der vom Fluiddruck im betreffendem Zyl inder 503 abhaengig ist, waehrend man perpherial ausserhalb der Abflussnuten 577 mindestens ein hydrodynamisches Tragfaeld erhaelt, das nicht vom Fluiddruck im Zyl inder 503 abhaengig ist, sondern von der Drehzhal des Aggregates abhaengt. Bei richtiger Bemessung und Ausbildung ergaenzen sich beide Flaechen-Tragteile so, dass der erreeichbare Drehzahlbereich mit gu = tem Wirkungsgrade ausgedehnter wird und eine Maximalausdehnung sichert.In contrast, FIGS. 45 and 46 show how support surface fields 580 can also be appropriately designed in other ways in order to achieve the object of the invention. This is understandable ^y by consideration of Figure 5 together with the sectional figure 56. The part of the Aussnflaeche of the piston shoe between the drain grooves 577 of Figure 45 has the radius = 513 is the radius of the diameter "da" of the Kolbenhubfuehrung 514th Thereafter, the support surface (the support surfaces) 580 receives a different curvature according to the invention, namely one by which the support surfaces 580 are converted into wings 580, that is to say are converted into hydrodynamically supporting surfaces. In order to be able to form a hydro = dynamic supporting field, the surface 580 must not have the same radius as the middle part between the drain grooves 577. Instead, the gap between the guide surface 513 of the piston stroke guide 514 and the support surface 580 must be in the direction of travel to narrow a certain relationship. Therefore, the support surface in question, which now becomes the wing 580, has a radius 913 instead of the previous radius 513. In a first approximation, the correct radius 913 and its center, which is different from the rotor or stroke axis of the hybrid 514, are found (mean = axis) by the length "LL" of the area in question than half the 580 U fangsflaeche _m regards a circular cylindrical shaft. Of course only for the calculation. So you calculate:

to get the diameter of the imaginary shaft. In equation (30) mean: "fv" = correction factor, "pi" = "F" and "dv" = the diameter of the imaginary shaft. The correction factor "fv" is between 1 and 2, closer to 2. If you have the diameter of the shaft, you can, if you equate the relative speed between the surfaces 513 and 580 equal to the peripheral speed of the shaft with the diameter "dv", calculate the hydrodynamic load-bearing capacity of the fit gap between the surface parts 513 and 580, for example according to the hut or according to the beech: "The basics of bearing lubrication" by Werner H.Kara (petroleum oil storage). From this, the necessary distance between the surfaces 513 and 580 can be calculated from the start at the groove 577 and if one has this distance for each place of the surface 580, then one can calculate the curvature of the surface 580 and thus its radius or its radii and ih = re central axis or their central axes. If this is carried out correctly, the surface 580 in question can be designed in such a way that it receives a maximum load-bearing capacity for the desired speed range. 45-46 there is therefore a hydrostatic support part between the grooves 577, which carries the main part of the radial load and which is dependent on the fluid pressure in the relevant cylinder 503, while at least one hydrodynamic support field is obtained peripherally outside the drain grooves 577, which is not dependent on the fluid pressure in the cylinder 503, but depends on the speed of the unit. With correct dimensioning and training, the two surface support parts complement each other in such a way that the achievable speed range is expanded with good efficiencies and ensures maximum expansion.

FIG. 47 serves to overcome another difficulty which the invention recognizes and which also results from the particularly high pressures in the fluid which enable or secure the units and elements and the piston shoes of the invention. At such high pressures, I feel the walls of the cylinders expand very noticeably from the inside under the high fluid pressure if the wall thicknesses are not very thick. The expansion of the cylinder under internal pressure is, for example, according to a formula obtained from Mr. Igarashi from Riken Seiki (OJia, Japan):

Bei so hoher Radialausdehnung des Innendurchmessers des Zylinders kommt es vor, dass der Kolben oder dessen Dichtung nicht mehr ein= wandfrei dichten kann. Auch bei Verbrennungsmotoren mit hohem Innen= druck im Zylinder kann diese Erscheinung auftreten. Ausserdem ver= schlingen dicke Waende viel Material, machen teuer und schwer. Daher wird nach der Erfindung um den Zylinder 102 eine mit Druckfluid ge= fuellte Kammer 103 herum angeordnet, die durch die Aussenwand 104 gebildet und verschlossen ist. Der Druck in der Kammer 103 kann jetzt so bemessen werden, dass jede Radialausdehnung der Zylinderwand 102 verhindert wird, oder deren Radialausweitung unter Druck im Zylinder 100 auf ein ertraegliches Mass verringert wird. Da die Ausdehnung der Aussenwand 104 nicht schaedlich ist, koennen nach der Erfindung duennere Waende 102 und 104 verwendet werden, sodass das Aggregat leichter, billiger und betriebssicherer wird.

With such a large radial expansion of the inside diameter of the cylinder, it happens that the piston or its seal can no longer seal = wall-free. This phenomenon can also occur in internal combustion engines with high internal pressure in the cylinder. In addition, thick walls swallow a lot of material, making them expensive and heavy. Therefore, according to the invention, a chamber 103 filled with pressurized fluid is arranged around the cylinder 102, which chamber is formed and closed by the outer wall 104. The pressure in the chamber 103 can now be dimensioned such that any radial expansion of the cylinder wall 102 is prevented, or its radial expansion is reduced to an acceptable level under pressure in the cylinder 100. Since the expansion of the outer wall 104 is not harmful, thinner walls 102 and 104 can be used according to the invention, so that the unit becomes lighter, cheaper and more reliable.

• " b " = Radialaufweitung
• n = D/d
• D = Aussendurchmesser des Zylinders
• d = Innendurchmesser des Zylinders
• E = Elektrizitaetsmodul
• P = Druck und
• "b " = Spannung,

Since the pressure in the chamber 103 counteracts the pressure in the cylinder 100, the cylinder wall 102 can no longer expand as far as desired. The supply of the fluid and pressure into the chamber 103 can take place through the inlet 106. Inlet 107 directs the relevant pressure or fluid into chamber 108. Channel 119 shows the inlet to cylinder 100, which may also act as an outlet. The figure also shows that a pump unit can be built directly with the cylinder 100 and the arrangement according to FIG. 57. 57 shows the pump housing 109, the rotor 11, the cylinder groups 112 and 116 with the piston groups 113 and 117, as well as the piston shoe groups 114 and 118. If, for example, pressure fluid is sent through inlet 110 into the cylinder 112 of the unit, the action then takes place group 112, 113, 114 as a motor and drives group 116, 117, 118 as a motor so that the motor feeds through line 119 into cylinder 100. A pressure ratio can be generated in this way by giving the group with the cylinders 112 a larger swallowing volume than the group 116, 117, 118 conveying volume. The desired high pressure in the cylinder 100 and / or in the chamber 103 can then be generated from a low or medium pressure unit. The drawing therefore shows a channel 120, which may direct the pressure from line 119 into the outer chamber 103 (via connection 106) in order to have the same pressure in the cylinder 100 and the outer chamber 103, so that the radial expansion of the cylinder wall 102 becomes zero. even if this wall has only a small wall thickness. The following values apply in equation (31):

• "b" = radial expansion
• n = D / d
• D = outer diameter of the cylinder
• d = inner diameter of the cylinder
• E = electricity module
• P = pressure and
• "b" = voltage,

FIGS. 48 to 50 show sections and views through a deep-diving high-pressure piston shoe for radial piston units of the invention. These are especially for the pumps and motors

This patent application is suitable, but it can also be used in general in the high-performance radial piston units of the applicant and inventor, which have arisen in particular through the patent specification of AnmeLder. and Inventor, DB-PS 1,30 ₂ , 469. After this, the rotor 608 has the radial webs 108 for guiding the long piston stroke and the piston shoe has seen from above the "H-shape", as it is called at the German Patent Office.

This "H-shape" characterizes the "deep-diving piston shoe" of the applicant and inventor, which enables the large piston stroke in a small space and thus for pumps and motors for the invention. Because the high performance with low weight comes from the "deep diving" possible H-shape together with the design of the piston stroke guide and the rotor 608,

The piston shoe of the invention according to FIGS. 48 to 50 remains a deep-diving piston shoe in the above sense, but is intended to provide greater resistance to deflection under high load and can therefore be used not only for aircraft, but also for presses and excavator pumps. It is also a finding of the inventor that conventional piston shoes show a tendency to tip over because they require too long piston shoe feet stored within the piston and are therefore too far away from the outer sliding surface 512 of the piston shoe. So it must be doubted whether the piston shoes that appeared in the FRG around 1970 for high speeds could ever overcome this tendency, which the inventor now recognized and which must lead to leakages. Because the distance between the swivel joint center is too far from the outer surface. This cannot be prevented if the piston shoes are not deep diving. In the case of the deep-diving piston shoe, which was introduced by the applicant and inventor at the beginning of the 1960s, on the other hand, the angle between the pivot center (axis) in 505 of FIGS. 4-9 from the outer edge (the outer edges) of the sliding surface 512 can be so acute (see FIG. 49 ) that such a tendency to tip with leakage can never occur.

On the other hand, the deep-diving piston shoe has a tendency to bend, i.e. the side guide parts are bent away from the piston stroke guide under high loads and constant load changes. Not millimeters, but in the order of a thousandth of a millimeter and depending on the number of load changes and the load change speeds. The piston shoe of FIGS. 49 to 50 is to overcome or restrict these defects, in particular restrict or prevent the deflection of the side parts of the deep-diving piston shoe. For this purpose, the rotor 508 according to the invention preferably receives the further recesses 752 for receiving the support 751 of the new comfort shoe of the invention. The new marking of the piston shoe of the invention consists in that the slots between the side legs of the "H-shape" are no longer open in the running direction of the piston shoe, but are closed by a bridge that connects the outer ends of the legs of the "H". The deep-diving piston shoe of the "H-shape" thus has at its front and rear ends of the lateral guide parts 602 the bridges which connect them and which are usually in one piece with them? 51. The slots of the "H" between the legs of the "H" are thus replaced by the two windows 750 penetrating radially through the piston shoe. They are located on both sides of the piston shoe central part 302 and thus between the central part 302; one of each of the bridges 75 ₁ and the side parts 602. When the rotor rotates, one of the rotor radial webs 708, one in each of one of the windows 750, is immersed in these windows. As in FIG. 49, pistons according to our DE OS 30 41 367 are used , then, as can be seen in FIG. 49, the piston fingers 741, 742 also engage in these windows 750.

The fingers 741, 742 and the webs 708 pass partially and temporarily completely through the windows 750 when diving into the pen, that is to say during the long piston stroke. The pistons 501 slide into the cylinder 503 of the rotor 608. In the piston swivel bed, which today is mostly a partially cylindrical surface of a hollow cylinder, the central part of the piston shoe is mounted by means of the swivel joint 302, which has a cylinder part surface which is complementary and matches the surface 510 . Piston bed 510 and piston shoe swivel roller (on the center web) 302 thus form surfaces that slide against one another with swivel radius 509 around swivel axis or swivel point 505.

The piston shoes are mostly cast in this form today and are mostly made of a certain bronze. The piston shoe beds are ground in series with shape grinding wheels, with a number of pistons with axially aligned beds being clamped in a device on the surface grinding machine. In the figures, the new piston shoe also has the lateral radial inward extensions 951 which can engage in the rotor recesses 952 and which can be provided with the recesses 952 for arranging pull rings for the suction stroke for their sliding on the inner pull surfaces. Otherwise, the piston shoe 3 ₀ 2 has in the outer surface, that is to say in the sliding surface 312, the drain grooves 577 of our DE patent 25 00 779 between the surrounding sealing surfaces of the pressure fluid pockets 515 and the front and rear piston shoe support surfaces 580. They are fed through the channels 616 by pistons and shoes from the cylinder 503 with pressure fluid. But they need something new Piston shoe of the current invention no longer the longer extension in the circumferential direction of the inventor's patent 25 00 779, since the new piston shoe in the direction of the swivel axis 505 is considerably strengthened by the bridges 751 according to the invention. The support surfaces 58 _D can now either be provided with additional drainage grooves in order to create interstatic bearings according to the inventor's systems, or hydrodynamic tread parts can now be formed over the support surfaces 580, which are then in the sense of FIGS. 45, 46 of this patent application act and are trained. The central web as the upper part of the swivel joint: kes 302 has bevels 740 which make it possible to swivel between the most radial: towards 708 of the rotor or the fingers 741, 742 of the piston 503 without bumping into them. You can also see them in Figure 50-R. This figure also shows that the new piston shoe can be extended so long in the axial direction, i.e. parallel to the axis of the rotor 608 and the pivot joint 50 or the pivot axis 505, that very large and load-bearing sliding surface parts of the sliding surfaces 602 remain. They can be so far in the axial direction mentioned that the windows 750 of the invention appear to be almost short. The windows 750 of the invention must not be made too short, however, so that a sufficient encirclement of the pistons and their arms 741, 742 is encompassed by the wall of the radial web 708 of the rotor 608, in order to prevent lateral tilting or a tendency to do so, because the clamps would Piston fingers or piston arms 741, 742 of the piston 503 act on the radial webs 708 of the rotor and in the cylinder 703.

FIGS. 51 A to C show a further embodiment of a piston-piston shoe arrangement according to the invention. Here, the piston shoe of FIGS. 1 and 2 of our European OS 0 064 563 has been modified such that the central part 12 of the E-05, which has the number 302 in FIG. 51, is a completely full cylindrical roller in the radius 509 about the pivot axis 505 . The shoe is secured here against falling out of the piston bed by the pin 762 which engages in the piston arms, on which the swivel roller 302 rests and swings with its swivel surface 511. Above the central part 5 302, which here forms the completely cylindrical cross-section, the surface 511 is at a distance from the stifi 762 or they touch it only in one point. This arrangement is usually sufficient for normal engine operation. It is also cheap and prevents jamming in fits that have become too tight during production. On the other hand, however, self-priming pumps are occasionally required, and for these, as well as for motors when they are used suddenly in special machines or vehicles, it is therefore better to hold the swivel roller 302 more firmly in the piston bed. Therefore, the holding block 761 is arranged in FIG. 51-C and with the pin 761 in the piston fingers, which are also called piston arms in the publication mentioned. It is then expedient to provide the radially inner surface of the block 761 with the radius of the swivel surface and the bearing bed surface, that is to say with the radius 509 or 508 of the roller 302 and the bearing bed 510. This inner surface now has the reference number 710 and the surface 511 slides on it.

FIG. 52 shows that the pivot axis 505 can also have a greater distance than that of the radius 509 from the block 761. Then the inner surface 910 of the block 961 is given a larger radius than the radius of the swivel roller lower surface, namely the radius 909. The upper part of the swivel roller or the central web of the piston shoe is then given the matching radius, that is to say the radius 909 is almost the same 908, which forms the upper outer surface 911 that slides on the inner surface 910 when the piston shoe swivels in the piston. So that no parts can collide, the block 961 and the piston shoe swivel roller 302 beyond the fitting surfaces 910 and 911, with the radii 908 and 909, can be provided with bevels or recesses 963 and / or 964.965.

FIGS. 53A and 53B show a further design option for connecting the piston to a piston shoe. It can also be used for non "H-shaped" piston shoes. According to the invention, the piston receives the recesses extended in the radial direction somewhat more than the radial extension of the pin 966 = gen967. After the swivel roller 302 of the piston shoe has been inserted into the piston 501, the pin 966 is inserted protruding through the swivel roller 302 and protruding into the recesses 967. Since it is longer than the swivel roller, but the recesses are closed radially upward, the pin 966 prevents the piston shoe from falling out of the bearing bed 510 of the piston 501. The ends of the Pin 966 can move freely in the recesses 967 when the piston shoe 502 swings with the swivel roller 302 in the log belt 510 of the piston 501. It is preferred to arrange the line ducts 516, 517 from the relevant cylinder through the piston and the swivel roller, as well as the central part of the piston shoe, not laterally in the piston axis, but laterally therefrom, for example as shown in the figures. The pressure fluid pockets 968 in the swivel roller 302 are then placed such that they hit the channels 516, 517 through the piston 501 and the piston shoe 502 when the piston shoe is pivoted. The underside of the piston shoe guide part can have recesses 969 so that piston radial fingers can engage or immerse in them and collision of piston and piston shoe parts is avoided.

FIG. 54 is a section from FIG. 7.

The figure 54 is supplied so that it can be used as the summary for the published specification because the

Figure 7 would exceed the scale or the place in the published specification. The description of FIG. 7 therefore also applies to FIG. 54.

Claims (10)

1.) Unit with at least one pressure fluid chamber through which fluid flows, the volume of which increases and decreases periodically, characterized in that said pressure fluid chamber (61, 116, 301, 503) has a piston (, 305, 302, 501), a piston shoe (52, 21, 502,) and / or at least a conical ring element (1,2,5,6,11,12,307) with a centering and a support (94,305,315 etc.) are assigned. 2.) Unit according to claim 1, characterized in that said elements consist of fibrous materials with connections = means, for example made of carbon fiber and the radially outer parts of said elements (1, 2) according to FIG. 1, for example by means of epoxy resin are glued. 3.) Unit for example according to claim 1 or without the element of claim 1, characterized in that two pistons (59, 60 or 58, 59 of Figures 2 or 4) lie on each other radially freely movable and means (for example, 83.84 or 64,70,72,71) are arranged to hold the two pistons together when the suction stroke is depressurized. 4.) Unit according to claim 1 or unit with a pump piston, for example 58, characterized in that the piston (58) is acted upon at both ends with different fluids, wherein at least one fluid can be a non-lubricating liquid and an annular groove in the middle of the piston Example 65 is arranged with a drain line (66) (FIG. 2) which serves as a collection arrangement for collecting leakage of one, the other, or both fluids. 5.) Unit with a piston, a piston shoe and a piston = stroke guide, characterized in that the piston shoe (2, 502) is arranged between the piston (5.501) and the piston stroke guide (514) and with its outer surface (512) on the Piston stroke = guide surface (513) of said piston stroke guide slides, said piston forms a head as a bearing bed (510) with a radius (508) around the pivot axis (505) of the piston shoe on the piston (501) on which the pivot surface (511) also coexists same radius (509) rests and swivels, a holding surface (513, 531, 532, 523, 524, 572) is present in the unit or on the piston shoe, the piston (501) is perpendicular to its axis (504) and para to the swivel axis (505) = Ilele recesses (5/8), into which pins (507) arranged in its transverse bore (506) protrude and holders (529,619) are arranged which touch part of the holding surface with a part (574,537,538) and said holder with an in the named A recess (518) of an immersible eye (520) is provided, which engages around the relevant end of said pin (507), the arrangement of the eye within the recess making it possible for the eye to be inserted into the cylinder ( 503) a pump or a motor with rotor or body (608) can occur temporarily or continuously. 6.) unit, characterized in
that a rotor (401, FIG. 4) forms an outer bearing (461, 462) with pressure fluid pockets (407, 477) which can be acted on with different pressures from different working chambers (439, 440) and / or in the rotor mentioned, pressure = chambers (402, 403) with pressure pistons (404, 405, FIG. 4) are arranged, to which auxiliary pressure pistons (411, 611) are assigned in axially diametrical or opposite auxiliary pressure chambers (410, 412) and / or double control mirrors designed according to FIG. 26 are arranged with pressure fluid zones around 422 and around 420, or that Unit is provided with a piston stroke guide and a piston shoe on a piston (FIGS. 27 to 30, 42 to 46) such that a pressure fluid pocket (515) with a surrounding sealing surface (578) is arranged in the outer surface (512) of the piston shoe (502), whose pressure fluid zone has a larger cross-section parallel to the piston stroke guide (514) than the cross-section through the piston (501) perpendicular to its axis (504), or the unit with a piston and a piston shoe according to FIGS. 31 to 38, and that the piston forms a head as a bed with the bed surface 510 with the radius 508 around the pivot axis 505, on which the pivot surface 511 of the piston shoe 502 with radius 509 rests pivotably = bar and / or the piston head axial extensions 588 bil = det that protrude beyond the diameter 637 of the piston 501 and / or holding surfaces 583 are arranged and also the piston 501 and the piston shoe 502 by means of brackets 584 with pivot are connected to one another, or the unit is equipped with a piston stroke guide and a piston shoe between this and a piston according to FIGS. 39, 40 and that a transverse pin 595 is arranged in the piston, the ends of which pivotally support piston shoes 1102, 1202 and pressure fluid pockets 515 Piston shoes are assigned, which are periodically or continuously supplied with pressure fluid from the relevant cylinder 503 and lubricate the sliding movement or the swiveling movement of the piston shoe and reduce its friction, or the unit with a piston (101) that can be reciprocated in a cylinder (100) ) and that the cylinder wall (102) is surrounded by a chamber (103) filled with pressure (FIG. 47), whereby the chamber with the pressure therein contains a radial expansion or breakage of the cylinder wall (102) under internal pressure in the cylinder (100) prevents or that The unit is provided with a rotor, a piston and a piston, for example according to FIGS. 48 to 50, and that the rotor 608 is provided with radial webs 708 which can be immersed in the window 750 of the piston shoe and the named piston shoe in front and behind in the running direction the legs and the windows 750, the legs 602 of the deep-tapering, H-shaped, piston shoe connecting and increasing its rigidity and rigidity, bridges 751 are arranged, which can dip into the recesses 752, or 752 and 952 arranged in the rotor 608. 7.) unit, for example according to claim 1 or characterized,
that a piston (8,59,74) is assigned a piston shoe (52) which is pivotably mounted on its spherical piston head, the end of which slides with a partially cylindrical running surface on a cylindrical piston stroke guide surface (57 to 56) and the lateral ends of which bulge in two opposite directions (79) form radially of the shoe center (78), which hold the piston shoe on the piston by means of holding surfaces 82 of a body (eg 148) or the centering of the shoe on the underside: support or secure. 8.) Unit according to Claim 1, characterized in that a radially inner and a radially outer control part, which is provided with at least one control slot through which fluid flows, are arranged (for example according to FIG. 11, outer control part 99 with slot 101 and inner control part). part 102 with slot 149) and between the two mentioned (radially inner and radial external) control parts their volume is arranged periodically increasing and reducing chambers through which fluid flows, the fluid flow flowing from one of the slots to another of the slots (101, 149, 150) and thus from one of the control parts to the other of the control parts (99, 102) the chamber arranged between them takes place (or in the opposite direction) and the chamber by means of conical ring elements (1, 2 etc.) or by a cylinder and piston (eg 60.61 or 58.70, of FIGS ) can be formed and a piston shoe (21) or a guide surface (57, 56, 156, etc.) and a through-flow channel (161, 74, 73, etc.) are assigned to the chamber parts arranged between the slots of the control parts. 9.) Unit according to claim 1, characterized in
that the conical ring elements are replaced by a cylinder (48, 61 FIG. 2) with a piston (58) that can be reciprocated and seals in the cylinder (FIG. 2), the said piston is designed as a reciprocating piston in contact with pressure fluid, and a driving piston that contacts the reciprocating piston , is assigned and the driving piston (59) is provided with a pivotable piston shoe (52), the pivoting surface (75) and running surface (57) of which protrude beyond the cross section of the piston (58) and / or further measures such as a radially flexible connection between the pistons , Relief chamber arrangements (67) or mixed fluid or leakage arrangements (65, 66, 70, 64, etc.) according to the figures 2, 3 or other of the figures. 10.) Unit according to at least one of the figures of this patent application, characterized in that at least one training, which is shown or described in at least one of the figures or in the description, is arranged, or, unit according to claim 1, characterized in that
that a plurality of elements (1, 2) form a plurality of inner spaces, chambers (14) 61) arranged axially one behind the other, one of said supports is a piston (36) with piston shoe (21) and a further support is a seat ( 107,98), the piston shoe slides on a running surface 156 of a piston stroke guide (99), a guide extends from the piston, for example a hollow tube (74,108) through the chambers or inner spaces of the elements and optionally in a bore (161) is guided, the piston shoes (21) have extended contact surfaces for guidance on the piston stroke guide, pressure fluid may be conducted in bearing pockets 112 of the piston shoe and one of the supports is subject to periodic movement or constant rotation, so that the interior spaces or chambers (14 , 61) form a pump or a motor and act as such an aggregate without a piston seal or an element guide having to be arranged, or that the aggregate for Generation of a fluid stream (802) is used and said fluid = stream for the purpose of cutting or atomizing onto a body or a jet onto fluid or powder, or. Solid material (801) is directed (Figure 16.),
or that the aggregate corresponds to the condition of one or more of the calculation methods of this patent application or a definition or condition of the description of the exemplary embodiments of this patent application.

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