EP0510009B1 - Rotary valve machine - Google Patents

Abstract

PCT No. PCT/DE90/00971 Sec. 371 Date Jun. 11, 1992 Sec. 102(e) Date Jun. 11, 1992 PCT Filed Dec. 18, 1990 PCT Pub. No. WO91/10812 PCT Pub. Date Jul. 25, 1991.A rotary slide vane machine has a housing with the inner contour of a Pascalian screw. The rotary piston is a hollow circular cylinder with an axis of rotation within the zero point M' of the Pascalian coordinates. The control shaft with eccentric segments is arranged inside the piston and has an axis of rotation within the center M of the Pascalian base circle. Slide vanes of a length L are guided within grooves of the rotary piston. Bearing bushings are rotatably guided on the eccentric segments and having flanges connected to the wings of the vanes. The length L is 6D+2r8s+2d61+2e+2s (D is the diameter of the base circle, r8s the radius of the eccentric segments, d61 the wall thickness of the bearing sleeves, e the minimal displacement of the wing, and s the minimal distance between bearing sleeve and machine parts moving thereto). The flange contour is limited by an arc having the minimal distance s to the inner mantle surface of the rotary piston and by distance S'E'min of a wing to an inner edge E' of neighboring wings and by an arc with the radius rKb minus the minimal distance s.

Description

The invention relates to a rotary vane machine on the geometric basis of the conchoid of the circle, also called Pascal's screw, in which the shapes and proportions of the running machine parts are coordinated and optimized with one another according to a dimensioning rule.

Rotary vane machines of the type mentioned are known from the patents 'US 1 994 245' and 'DD 46 761' and have the structure described below:

A rotary piston 4 in the form of a hollow circular cylinder is rotatably mounted in both housing covers in a hollow cylinder 1, which is closed on both sides with a housing cover and has the inner surface profile E of the circumferential line of the Pascal screw, axially to the hollow cylinder. The axis of rotation of the rotary piston coincides with the center point M 'of the center circle K' of the Pascal screw shown on drawings 1 and 2. It follows that the rotary piston with the radius r _Dk with its outer surface at the approximation point (-xr _Dk yo) = E180 ° approaches the inner surface of the hollow cylinder to a gap.

The control shaft 8 in the design as a crankshaft with the radius r of the base circle K is, through the hollow rotary piston and through central openings in its end walls, also rotatably supported in both housing covers, the crankshaft axis arranged axially to the rotary piston being centered on the center M of the Base circle K covers.

On the crank pin of the control shaft, one or two rotary valves 6 are rotatably supported in their half length L / 2. They correspond to functionally offset pairs of wings 6f, which are connected to one or two central bearing sleeves 6l via webs 6s to form a rigid rotary slide valve. Their center points P coincide with the axes of the crank pins and describe the base circle K when the control shaft rotates.
The rotary valves are also guided in diametrical longitudinal grooves 5 of the rotary piston.

In machines with one rotary valve per work area, the mass forces of the crank pin rotating on the base circle K and the rotary valve must be balanced with two counterweights on the control shaft. The installation of more than one rotary valve per work area enables the balancing of the mass forces without counterweights if the rotary valves have a longitudinal and transverse axis symmetry and the fork of the webs 6s with the bearing sleeves 6l is arranged so that they do not overlap each other.

Since the rotary piston _{axis of rotation} lies in M ′ _{(xo yo)} = P _{(xo yo)} , the center circle K ′ touches the base circle K in P _{(x2r yo)} = S _(x2ryo) , r ′ = 2r and the points P and S die have the same web speed and direction of rotation, a central external toothing on the control shaft 8 with the pitch circle K _t may be in a centric internal toothing in a rotary piston end wall with the pitch circle K _t grab '.
The ratio of the number of teeth from external teeth to internal teeth is 1: 2. The toothing of the control shaft with the rotary piston ensures the exact synchronous rotation of these two machine parts in a ratio of 2: 1 and absorbs the bending forces acting on the rotary valve.

The machine geometry of the Pascal screw also enables constructions without toothing of the control shaft 8 with the rotary piston 4, because the sequence of movements of the rotary slide valve causes a forced rotation of the rotary piston and control shaft in the ratio 1: 2 mentioned.
However, due to the torque acting on the control shaft, an additional bending moment acts on the rotary valve. The bending of the rotary valve can lead to tilting in the longitudinal grooves 5.

With the synchronous rotation of the control shaft and the rotary piston, the mass points of the rotary valves located at the center of point P of the rotary valve move with the control shaft speed on the base circle K with simultaneous rotation of the rotary valve around their axis in point P with the rotary piston speed. Here, the rounded side edges 6e of the rotary valve describe the circumferential line E of the Pascal screw and can therefore be guided without contact along the inner surface of the hollow cylinder 1 with the same profile.
At the same time, the rotary slide valve performs a periodic thrust movement in relation to the rotary piston side wall, which reaches the maximum thrust NE _max = 2D once in each direction in each direction in the 0 ° position.

In principle, the 'DD 46 761' patent provides for the installation of only one rotary slide valve per work area with a different working principle that deviates from the problem posed by this patent.
The gas outflow takes place here via channels in the rotary lobe and the intended high internal compression is achieved by using the rotary lobe 4 as a rotary slide valve.
In the text and in the drawings, the control shaft 8 is shown as a crankshaft. The problem of the design and lubrication of the rotary slide bearing is not addressed here.

In the patent 'US 1,994,245' the control shaft is also shown in the text and on the drawings as a crankshaft with a lubricant channel leading through the crankpin and cheeks.
The design principle on which the invention is based only provides for the installation of two rotary valves per work space. The rotary valves consist of two halves divided in their bearings.

• Die Kurbelwelle kann geteilt und mit bekannten Schraub- und Steckverbindungen versehen, welche den Einsatz ungeteilter Nadellager erlauben, wegen der niedrigen Exzentrizität nur unter hohem Fertigungsaufwand bei einer stark reduzierten Biege- und Torsionsfestigkeit hergestellt werden.
  Man wird deshalb,wie bereits in der US-Patentschrift 1,994,245 vorgeschlagen, die Steuerwelle als ungeteilte Kurbelwelle mit einem Schmiermittelkanal in den Kurbelzapfen und Wangen ausführen. Die Drehschieber müssen dann mit geteilten Gleitlagern versehen werden.
• Damit entsteht jedoch ein neues Problem; das Ausbohren des Schmiermittelkanals. Er muß fertigungstechnisch aufwendig aus mehreren, durch Kurbelwelle, -zapfen und -wangen geführte Teilbohrungen zusammengesetzt werden. Wegen der geringen Exzentrizität der Steuerwelle ist das Durchbohren der Kurbelzapfen bei höheren Drehschieberzahlen besonders schwierig.
• Die Steuerwelle kann aufgrund der Maschinengeometrie an ihren Enden nur einmal in jedem Gehäusedeckel gelagert werden. Dies führt bei größeren Baulängen, die sich zwangsläufig beim Einsatz einer höheren Drehschieberzahl ergeben und wegen der günstigeren Form der Drehschiebermaschine auch erwünscht sind, in der Ausführung als Kurbelwelle wegen der kleinen Kurbelzapfenquerschnitte zu einer geringen Biege- und Torsionsfestigkeit.
  Hierdurch wird die Belastungsfähigkeit der Steuerwelle und die Genauigkeit des Spaltes zwischen Drehschieberseitenkante 6e und der Hohlzylinderinnenfläche mit einem entsprechenden Rückgang des Wirkungsgrades gemindert.
• Die auf der Steuerwelle zur Verfügung stehende Fläche für die Drehschieberlager ist in der Ausführung als Kurbelwelle in zweifacher Hinsicht reduziert:
  Wegen der kleinen Kurbelzapfendurchmesser sind auch die Lagerdurchmesser klein.
  Die maximale Gesamtlagerbreite ist durch den Steuerwellenabschnitt innerhalb des Hohlzylinders vorgegeben. Mit zunehmender Drehschieberzahl reduziert sich daher ohnehin die für den einzelnen Drehschieber zur Verfügung stehende Lagerbreite. Sie wird durch die notwendige Ausbildung der Kurbelwangen noch zusätzlich verschmälert.

The prior art described has the following disadvantages and shortcomings:
The control shaft is designed as a crankshaft. This means that the largest possible installation space in the rotary piston is available for the rotary valve; on the other hand, the crankshaft concept has serious disadvantages:

• The crankshaft can be split and provided with known screw and plug connections, which allow the use of undivided needle bearings, because of the low eccentricity only with a high manufacturing effort with a greatly reduced bending and torsional strength.
  Therefore, as already proposed in US Pat. No. 1,994,245, the control shaft will be designed as an undivided crankshaft with a lubricant channel in the crank pins and cheeks. The rotary valve must then be provided with split plain bearings.
• However, this creates a new problem; drilling out the lubricant channel. In terms of production technology, it must be composed of several partial bores that are guided through the crankshaft, journal and cheeks. Because of the low eccentricity of the control shaft, it is particularly difficult to drill through the crank pins at higher numbers of rotary valves.
• Due to the machine geometry, the ends of the control shaft can only be stored once in each housing cover. In the case of larger overall lengths, which inevitably result from the use of a higher number of rotary valves and which are also desirable because of the more favorable shape of the rotary valve machine, the design as a crankshaft results in a low bending and torsional strength due to the small crank pin cross sections.
  As a result, the load capacity of the control shaft and the accuracy of the gap between the rotary slide side edge 6e and the hollow cylinder inner surface is reduced with a corresponding decrease in efficiency.
• The area available on the control shaft for the rotary slide bearing is reduced in two ways in the design as a crankshaft:
  Because of the small crank pin diameter, the bearing diameter is also small.
  The maximum total bearing width is specified by the control shaft section within the hollow cylinder. As the number of rotary valves increases, the bearing width available for the individual rotary valve is reduced anyway. It is further narrowed due to the necessary design of the crank webs.

• Die Herstellung als geteilte Welle ist wesentlich einfacher. Wie auf Zeichnung 5 dargestellt, können die einzelnen Exzentersegmente 8s mit entsprechendem Winkelversatz auf eine durchgehende zentrale Führungswelle 9 aufgesteckt werden. Die Drehschieber können auch ungeteilte Nadellager erhalten.
• Der Schmiermittelkanal kann als zentrale Langlochbohrung 16 ausgeführt werden, von der radiale Bohrungen 17 in den Steuerwellenexzentern zu den Drehschieberlagern führen.
  Neben dem fertigungstechnischen Vorteil wird auch noch der funktionelle Vorteil der Schmiermitteldruckerhöhung durch die Zentrifugalbeschleunigung in den radialen Exzenterbohrunger 17 genutzt
• Die Biege- und Torsionsfestigkeit ist wegen der sich axial überdeckenden Exzentersegmente 8s größer.
• Durchmesser und Breite der Drehschieberlager sind größer.

The disadvantages mentioned can be avoided if the control shaft is designed as an eccentric shaft:

• The production as a split shaft is much easier. As shown in drawing 5, the individual eccentric segments 8s can be plugged onto a continuous central guide shaft 9 with a corresponding angular offset. The rotary vane can also receive undivided needle bearings.
• The lubricant channel can be designed as a central elongated hole 16, from which radial holes 17 in the control shaft eccentrics lead to the rotary slide bearings.
  In addition to the manufacturing advantage, the functional advantage of increasing the lubricant pressure by centrifugal acceleration in the radial eccentric bore 17 is also used
• The bending and torsional strength is 8s greater due to the axially overlapping eccentric segments.
• The diameter and width of the rotary slide bearings are larger.

In the design as an eccentric shaft, the control shaft cross-section becomes larger and therefore also the rotary piston diameter D _Dk with a constant maximum thrust NE _max .
There are therefore limits to the technical usability, because as the control shaft cross-section increases, the ratio of the maximum thrust NE _max to the rotary piston diameter D _Dk and thus the specific displacement _volume also decrease. The shift in the proportion of L and D _Dk to NE _max is shown in drawing 3.
At the same time, the control shaft gear diameter decreases in relation to the length of the rotary valve. The decreasing toothing is loaded with an increasing torque.

Another disadvantage of the prior art is the insufficient number of rotary valves per work area:
Analogous to the working principle of the vane cell machine, the highest possible number of rotary valves per work area should be aimed for in order to achieve comparable pulsation and internal compression values.
Rotary vane machines with one or two rotary vane are technically meaningless because of too high pulsation values and the too low maximum achievable internal compression.
The rotary vane machine of the above-mentioned US patent cannot be expanded to a higher number in the illustrated proportion from D _Dk to NE _max and the design with two rotary vane per work area.
It is not possible to derive from the prior art whether the construction principle can be extended to a higher number of rotary valves comparable to vane cell machines by using an additional eccentric shaft, which uses space.

This invention has for its object to achieve the design of the control shaft as an eccentric shaft with the simultaneous installation of a higher number of rotary valves per work space with the help of a dimensioning rule for rotary vane machines of the type mentioned above, and on the other hand the shape and proportioning of control shaft, rotary slide valves and rotary pistons with optimal use of the installation space to be designed so that the maximum thrust NE _{max is} in the greatest possible ratio to the rotary piston diameter D _Dk .

Fig. 1 und 2

zeigen die Pascalsche Schnecke in kartesischen und Polarkoordinaten.

Fig. 3

zeigt in einem Diagramm die Darstellung von D als Funktion von L

Fig. 4

zeigt die Pascalsche Schnecke in der Darstellung als Kardioide.

Fig. 5

zeigt die Bemessungsregel für die Drehschieberlängenproportion.

Fig. 6 - 8

zeigen den Teil eines Drehschiebermaschinenquerschnitts mit dem Einsatz von 2 - 7 Drehschiebern pro Arbeitsraum.

The drawing contains 8 drawings for better understanding.

1 and 2

show the Pascal snail in Cartesian and polar coordinates.

Fig. 3

shows in a diagram the representation of D as a function of L.

Fig. 4

shows the Pascal snail as a cardioid.

Fig. 5

shows the dimensioning rule for the rotary valve length proportion.

6 - 8

show the part of a rotary vane machine cross section with the use of 2 - 7 rotary valves per work area.

The state-of-the-art geometric principles are not sufficient to solve this complex task.

und Polarkoordinaten $$\text{E = D·cos(a) + L/2}$$

Dabei werden der Grundkreisdurchmesser D und die Drehschieberlänge L als willkürlich gewählte Konstanten benutzt.The machine geometry on which the prior art is based is limited to the circumferential line of the Pascal screw in the representation in Cartesian coordinates $$\text{(x² + y² - Dx) ² = L² / 4 · (x² + y²)}$$

and polar coordinates $$\text{E = DCos (a) + L / 2}$$

The base circle diameter D and the rotary valve length L are used as arbitrarily chosen constants.

From this relationship it can be deduced that it can in principle be the geometric basis for rotary vane machines, but it cannot be derived from it whether the required sizes of the machine parts of all conceivable rotary vane machine types can be coordinated.
The isolated consideration of a single curve is not sufficient for the construction of rotary vane machines, because with the choice of the value for D and L the course of the curve and the proportionality from D to D _Dk are influenced in a way that decides on the basic technical usability.
If, according to the state of the art, a single curve is isolated from the computationally possible infinite number of circumferential lines, one can even with a random choice of a usable size for rotary vane machines with a control shaft in the form of a crankshaft and one or two rotary valves per work area, make no statement as to whether this curve is also suitable for installing an eccentric shaft with a significantly higher number of rotary valves.

Sie ist auf Zeichnung 3 in einem Diagramm dargestellt. Aus ihr folgt, daß sich die Drehschieberlänge L bei einem angenommenen Wert für den Grundkreisdurchmesser D nur durch Änderung des Drehkolbendurchmessers D_Dk = 2D·f um den Faktor f ändern läßt. Die Geometrie der Steuerwelle 8 bleibt dabei mit dem Maximalschub NE_max = 2D konstant.
Der Proportionalitätsfaktor f ist hier als Hilfsgröße eingeführt. In ihm ist die Summe aller die Proportionalität von D zu L beeinflussenden Größen zusammengefaßt. Mit seiner Hilfe lassen sich in übersichtlicher Form die Eckwerte für die Proportionen der Drehkolbendurchmesser D_Dk zum Maximalschub NE_max der einzelnen Bauarten und daraus abgeleitet, das spezifische Verdrängungsvolumen festlegen:

• Für eine nutzbare Maschinengeometrie auf der Grundlage der Pascalschen Schnecke muß f ≧ 1 sein, weil auch in der einfachsten Bauart einer Drehschiebermaschine mit nur einem Drehschieber pro Arbeitsraum der Drehkolbendurchmesser D_Dk ≧ NE_max ≧ L/2 sein muß.
• Der Einbau von mehr als einem Drehschieber pro Arbeitsraum setzt einen Wert für f ≧ 2 voraus, weil der Drehkolbendurchmesser D_Dk ≧ 2NE_max ≧ 2·2D sein muß. Demnach ist die Grundgleichung für Drehschiebermaschinen mit mehr als einem Drehschieber pro Arbeitsraum $$\text{L = 2D·2 + 2D = 4D + 2D}$$
  

Aus dieser Beziehung kann die auf Zeichnung 5 dargestellte Bemessungsregel für die Proportion von Steuerwelle, Drehschieberlager, Drehschieberflügellänge und Drehkolben zur Drehschiebergesamtlänge abgeleitet werden, indem sie um die Durchmesser bzw. Länge der genannten Maschinenteile in der nachfolgenden Weise erweitert wird: $${\text{D}}_{\text{Dk}}{\text{= 2D·2 + (<figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s)·2}$$

Daraus folgt $$\text{f =}\frac{{\text{2D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s}}{\text{D}}$$

und $${\text{L = 4D + (<figure-callout id="8s" label="D" filenames="imgf0005.png" state="{{state}}">D</figure-callout>}}_{\text{8s}}{\text{+ 2d}}_{\text{6l}}\text{+ 2e + 2s) + 2D}$$

Hierbei ist:

r_8s

= Steuerwellenexzenterradius,

D_8s

= Steuerwellenexzenterdurchmesser,

d_6l

= Stärke der Drehschieberlagerhülse 6l,

e

= Drehschieberminimaleinschub in die Drehkolbenseitenwand d_Kr,

s

= Minimalabstand der beweglichen Maschinenteile.

Die genannte Bemessungsregel gilt auch für maschinen mit einem Drehschieber pro Arbeitsraum. Wenn solche Maschinen nach dieser Regel abgestimmt werden, können sie ohne Änderung der Proportionen auf eine höhere Drehschieberzahl erweitert werden.The first solution step consists in the inclusion of all arithmetically possible ratios from D to L in the construction by the representation of the one constant D as a function of the other constant L. The relationship is obtained in this way $$\text{L = 2D * f + 2D}$$

It is shown in a diagram in drawing 3. From this it follows that the length of the rotary slide valve L can only be changed by a factor of f with an assumed value for the base circle diameter D by changing the rotary piston diameter D _Dk = 2D · f. The geometry of the control shaft 8 remains constant with the maximum thrust NE _max = 2D.
The proportionality factor f is introduced here as an auxiliary variable. It summarizes the sum of all variables influencing the proportionality from D to L. With its help, the basic values for the proportions of the rotary _{lobe diameter} D _Dk to the maximum thrust NE _{max of} the individual types can be clearly defined and from this the specific displacement volume can be determined:

• For a usable machine geometry based on the Pascal worm, f must be ≧ 1, because even in the simplest design of a rotary vane machine with only one rotary slide valve per working area, the rotary piston diameter must be D _Dk ≧ NE _max ≧ L / 2.
• The installation of more than one rotary valve per work area requires a value for f ≧ 2 because the rotary piston diameter must be D _Dk ≧ 2NE _max ≧ 2 · 2D. Accordingly, the basic equation is for rotary vane machines with more than one rotary slide valve per work area $$\text{L = 2D2 + 2D = 4D + 2D}$$
  

The design rule for the proportion of control shaft, rotary slide bearing, rotary slide wing length and rotary piston to the total rotary slide length can be derived from this relationship by expanding it by the diameter or length of the machine parts mentioned in the following manner: $${\text{D}}_{\text{Dk}}{\text{= 2D2 + (<figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s) · 2}$$

It follows $$\text{f =}\frac{{\text{2D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s}}{\text{D}}$$

and $${\text{L = 4D + (<figure-callout id="8s" label="D" filenames="imgf0005.png" state="{{state}}">D</figure-callout>}}_{\text{8s}}{\text{+ 2d}}_{\text{6l}}\text{+ 2e + 2s) + 2D}$$

Here is:

r _8s

= Control shaft eccentric radius,

D _8s

= Control shaft eccentric diameter,

d _6l

= Thickness of the rotary valve bearing sleeve 6l,

e

= Minimum slide valve insertion into the rotary piston side wall d _Kr ,

s

= Minimum distance of the moving machine parts.

The specified dimensioning rule also applies to machines with one rotary valve per work area. If such machines are tuned according to this rule, they can be expanded to a higher number of rotary valves without changing the proportions.

If D _8s <D is selected, the control shaft 8 can only be constructed as a crankshaft.
The design as an eccentric shaft requires the value for D _8s > D and f> 2.5. The choice of the eccentric diameter depends on the desired load capacity and the diameter D₁₆ of the central elongated hole 16.
If the control shaft, as also shown in drawing 5, is composed as a divided eccentric shaft, consisting of a central guide shaft 9 and attached eccentric segments 8s, the value for D _8s must be expanded by D₉ - D₁₆.

The thickness d _{6l of} the rotary slide bearing sleeve 6l depends not only on the required strength, but also on the _type of bearing selected. The diameter of needle or roller bearings increases significantly compared to plain bearings and requires a higher value for d _6l .

The size e indicates the minimum insertion of the rotary valve into the side wall of the rotary piston d _Dk in the position of the maximum extension from the rotary piston at 0 °. With e = 0 in cross section, the inner edge E 'of the rotary vane 6f would be congruent with the outer surface of the rotary piston side wall. In practice one will choose a value for e> O ≦ d _Dk .

Several rotary valves can only be mounted on a common control shaft if they are divided between the webs 6s and the bearing sleeves 6l.
The construction shown in drawing 5 is advantageous in terms of production and assembly technology. The bearing sleeve 6l is here divided into two halves in the longitudinal direction of the slide valve and is provided on both sides at the point of division with a flange 6Fl, which forms a groove with the flange of the other half. The plate-shaped web 6s is inserted into this groove from the outside and fastened to the flange of each bearing sleeve half with a bolt or a fitting screw.
The described bearing sleeve shape can also be used undivided for needle and roller bearings.

The overall length L _6fl available for the flansh depends on the size of the minimum _insert e. If e = d _Dk , L can be _6Fl ≧ D. If e <d _Dk is _reduced , L _{6Fl decreases} by the amount d _Dk - e. Due to the curvature of the _{side wall} of the rotary _lobe , the height H _{6Fl of} the bearing sleeve flange in the 0 ° position is limited by the circular arc r _Dk - (d _Dk + s).

The value s indicates the minimum distance between the moving components in the rotary lobe that are not connected to each other.
The rotary valves are pushed out of the rotary piston to the maximum in the 0 ° position. Here the bearing sleeve flange 6Fl approaches the shortest distance to the inner surface of the rotary piston side wall. At the same time, the inner edge E 'of the rotary vane wing 6f approaches the bearing sleeve of the opposite rotary valve to its minimum distance. As shown in the drawings 5-8, this is somewhat larger with an odd number of rotary slide valves than with an even number of rotary slide valves, with the advantage of a slightly larger installation space for the bearing sleeves.

The value for s is primarily determined by the shape of the rotary slide bearing, especially by the shape of the bearing sleeve flange. A bulky construction increases the required clearance for the warehouse and thus increases the value for s accordingly. This in turn reduces the ratio of the maximum thrust NE _max to the rotary piston diameter D _Dk .

The maximization of this ratio without simultaneous weakening of the control shaft or the rotary slide bearing can therefore only be achieved with the optimal utilization of the installation space and is the content of the second part of the dimensioning rule described below:

Die Innenkanten E′ der Drehschieberflügel 6f bewegen sich auf einer Umfangslinie nach der Beziehung $${\text{E′ = D·cos(a) + (D + r}}_{\text{8s}}{\text{+ D}}_{\text{6l}}\text{+ s)}$$

Hieraus leitet sich eine Teilgröße für die Höhe H_6Fl des Lagerhülsenflansches ab. Sie wird durch den seitlichen Minimalabstand S′E′_min eines Drehschiebers 6 von der Innenkante E′ des benachbarten Drehschieberflügels 6f festgelegt und kann in die folgende, auf Zeichnung 7 dargestellte Beziehung gefaßt werden: $${\text{S′E′}}_{\text{min}}{\text{= M′E′}}_{\text{180°}}{\text{.sin(β) - (1/2 d}}_{\text{6f}}{\text{.cos(β) + 1/2 d}}_{\text{6f}}\text{+ s)}$$

Hierbei ist:

d_6f

= Die Drehschieberplattenstärke,

s

= Der Minimalabstand der Maschinenteile,

n

= Die Drehschieberzahl,

β

= 180/n° der Winkel zwischen zwei benachbarten Drehschiebern.

Für den Lagerhülsenflansch 6Fl steht noch ein weiterer, sehr wichtiger Bauraum zur Verfügung:
Wie auf Zeichnung 4 dargestellt ist, kann die Konchoide des Kreises auch als Kardioide, eine Sonderform der Epizykloide, aufgefaßt werden, wobei hier der Mittelkreis K′ mit seiner Innenseite auf der Außenseite des Grundkreises K abrollt und die Verlängerungspunkte E bzw. E′ dabei die Kardioiden beschreiben.
Die Kurve mit der kürzesten, technisch verwertbaren Verlängerung des Mittelkreises K′ um r_Dk = r′ = D, entspricht in der Beziehung $$\text{L = 2D·f + 2D}$$

dem Wert für f = 1.As already described and shown on drawing 2, the equation applies to Pascal's screw, based on polar coordinates $$\text{E = DCos (a) + L / 2}$$

The inner edges E 'of the rotary vane 6f move on a circumferential line according to the relationship $${\text{E ′ = DCos (a) + (D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ D}}_{\text{6l}}\text{+ s)}$$

A partial size for the height H _{6Fl of} the bearing sleeve flange is derived from this. It is determined by the lateral minimum distance S'E ' _{min of} a rotary valve 6 from the inner edge E' of the adjacent rotary valve wing 6f and can be summarized in the following relationship, shown in drawing 7: $${\text{S′E ′}}_{\text{min}}{\text{= M′E ′}}_{\text{180 °}}{\text{.sin (β) - (1/2 <figure-callout id="6f" label="d" filenames="imgf0005.png,imgf0007.png" state="{{state}}">d</figure-callout>}}_{\text{6f}}{\text{.cos (β) + 1/2 <figure-callout id="6f" label="d" filenames="imgf0005.png,imgf0007.png" state="{{state}}">d</figure-callout>}}_{\text{6f}}\text{+ s)}$$

Here is:

d _6f

= The rotary valve plate thickness,

s

= The minimum distance between the machine parts,

n

= The number of rotary valves,

β

= 180 / n ° the angle between two adjacent rotary valves.

There is another very important installation space available for the 6Fl bearing sleeve flange:
As shown in drawing 4, the conchoid of the circle can also be understood as a cardioid, a special form of epicycloid, here the center circle K 'rolls with its inside on the outside of the base circle K and the extension points E and E' thereby Describe cardioids.
The curve with the shortest, technically usable extension of the center circle K ′ by r _Dk = r ′ = D corresponds in the relationship $$\text{L = 2D * f + 2D}$$

the value for f = 1.

und die Kardioide E′ mit dem Wert (r_8s + d_6l + s) verlängert.According to the previous first part of the design rule, the cardioid is point E with the rotary lobe radius $${\text{r}}_{\text{Dk}}{\text{= 2D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s}$$

and the cardioids E ′ with the value (r _8s + d _6l + s) extended.

The cardioids are then in the same position in relation to the coordinate system of the conchoid of circle K if their vertex is on the x-axis Cartesian coordinates, or 0 °, based on polar coordinates, and the return point is in O = M ′.

Werden von der Verlängerung E′0° des auf dem Rückkehrpunkt in O = M′ stehenden Mittelkreises K′ Winkel zwischen 90° und 180° mit dem Scheitel in O = M′ abgetragen, dann beschreiben die Schnittpunkte der freien Schenkel mit dem Kreisbogen Kb beim Abrollen des Mittelkreises K′ um den Grundkreis K, verlängerte Kardioiden, welche die größte,durch E′90° gehende Kardioide in einem Punkt berühren, jedoch nicht überschneiden und mit dem doppelten Winkelbetrag um den Mittelpunkt M des Grundkreises K gedreht sind.Drawing 4 shows a circular arc Kb with the radius through the points E'90 °, E'180 ° and E'270 ° $${\text{r}}_{\text{Kb}}\text{=}\frac{{\text{(D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}{\text{+ s) ² + (<figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s) ²}}{{\text{2 (<figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s)}}$$

From the extension E'0 ° of the center circle standing on the return point in O = M 'K' angle between 90 ° and 180 ° with the apex in O = M ', then describe the intersection of the free legs with the circular arc Kb Rolling off the center circle K 'around the base circle K, extended cardioids, which touch the largest cardioids going through E'90 ° at one point, but do not overlap and are rotated at twice the angular amount around the center M of the base circle K.

This geometric relationship is of great practical importance for the rotary slide valve construction that the area within the circular arc with the radius r _Kb -s can be used as additional installation space for the bearing sleeve flange 6Fl beyond the above-mentioned height. It becomes more important as the number of rotary valves increases.

The individual parts of the bearing sleeve flange 6Fl describe corresponding cardioids with their rotation around the center M.
You can, if they are within the circular arc with the radius r _Kb -s, undercut the inner edges E 'of the adjacent rotary vane 6f. The degree of undercut increases with the height H _{6Fl of} the flange and the number n of the rotary valves.

In drawings 6-8, the rotary valve profile is shown in cross section when using 2-7 rotary valves per work area. The shape of the flange is chosen with a constant size so that it is also suitable for the highest number of rotary valves.
In each case in the 90 ° position, the boundary lines of the maximum available installation space are projected onto the profile of the bearing sleeve flange. It decreases significantly with increasing number of rotary valves n.

The value for s can be minimized to a gap width only if the profile of the rotary slide bearing is limited to the installation space described.

Formula Collection

Fläche F der Pascalschen Schnecke: $$\text{F = π·D²/2 + π·L²/4 D = 2r}$$

Mittelpunkt M der Steuerwelle 8 mit dem Exzenterradius r und Mittelpunkt der Außenverzahnung mit dem Teilkreis K_t in: $${\text{M}}_{\text{(xr yo)}}$$

Drehkolbenmittelpunkt M′ und Mittelpunkt der Innenverzahnung mit dem Teilkreis K_t′ in: $${\text{M′}}_{\text{(xo yo)}}{\text{= P}}_{\text{(xo yo)}}$$

Eingriff der Außenverzahnung in die Innenverzahnung in: $${\text{S}}_{\text{(x2r yo)}}{\text{= P}}_{\text{(x2r yo)}}$$

Näherungspunkt der Hohlzylinderinnenfläche an die Außenfläche der Drehkolbenseitenwand in: $${\text{}}_{\text{(-xrDk yo)}}\text{= E180° rDk = Drehkolbenradius}$$

Durchmesser D des Grundkreises K und Durchmesser D_t des Teilkreises K_t $${\text{D = D}}_{\text{t}}\text{= 2r}$$

Durchmesser D′des Mittelkreises K′und Durchmesser D_t′ des Teilkreises K_t′ $${\text{D′ = D}}_{\text{t′}}\text{= 2D = 4r}$$

Verhältnis der Drehschieberlänge L zum Grundkreisdurchmesser D: $$\text{L = 2D·f + 2D f = Proportionalitätsfaktor}$$

Drehkolbendurchmesser DDk: $$\text{DDk = 2D·f = L - 2D = L - D′}$$

Mittlere Drehschieberschubgeschwindigkeit c_m in der Längsnut 5: $${\text{c}}_{\text{m}}\text{= 2D·n/30 (m/sec) n = Drehkolben-U/min}$$

Verhältnis der Winkelgeschwindigkeit w der Steuerwelle 8 zur Winkelgeschwindigkeit w′ des Drehkolbens 4: $$\text{w = 2w′}$$

Verhältnis der Bahngeschwindigkeit v_p der Sekanten- bzw. Drehschiebermittelpunkte P zur Bahngeschwindigkeit v_s der Mittelkreisschnittpnkte S: $${\text{v}}_{\text{p}}{\text{= v}}_{\text{s}}$$

Umfangslinie E in Polarkoordinaten: $$\text{E = D·cos(a) + L/2 für L/2 ≧ 2D}$$

Umfangslinie E in Parameterform: $$\text{x = D·cos²(a) + L/2·cos(a)}$$

$$\text{y = D·cos(a)·sin(a) + L/2·sin(a)}$$

Fläche F_0°〉̶a der Pascalschen Schnecke in Polarkoordinaten: $${\text{F}}_{\text{0°〉̶a}}\text{= D²/4·sin(a)·cos(a) + LD/2·sin(a) + (D²/4 + L²/8)·a/360·2·π}$$

Verdrängungsvolumen V_0°〉̶a der Drehschiebermaschine: $${\text{V}}_{\text{0°〉̶a}}{\text{= (F}}_{\text{0°〉̶a}}\text{- π·rKr²·a/360).h}$$

Verdrängungsvolumen V_n pro Drehkolbendrehung bei n Drehschiebern pro Arbeitsraum:   F_Ds=Drehschieberprofilfl. $${\text{V}}_{\text{n}}{\text{= (F}}_{\text{0°〉̶90/n°}}{\text{- π·rKr²/4n - F}}_{\text{Ds}}\text{)·4·n·h}$$

Umfangsgeschwindigkeitsmaximum v_Bmax der Drehschieberendpunkte E bei E0° $${\text{v}}_{\text{Bmax}}\text{= f·(L + D′)·Π}$$

Drehschieberschub NE:   NE = D·cos(a) + D
Drehschiebermaximalschub NE_max : $${\text{NE}}_{\text{max}}\text{= NE0° = 2D = D′}$$

Drehschieberminimalschub NE_min bei n Drehschiebern pro Arbeitsraum: $${\text{NE}}_{\text{min}}\text{= D·cos(90/n)° + D}$$

Festlegung der Beziehung $$\text{L = 2D·f + 2D}$$

durch die einzelnen Bauteile der Drehschiebermaschine. Hierbei ist: $$\text{f =}\frac{{\text{2D + r}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s + e}}{\text{D}}$$

$${\text{D}}_{\text{Dk}}{\text{= 2(2D + r}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s)}$$

$${\text{L = 6D + (2r}}_{\text{8s}}{\text{+ 2d}}_{\text{6l}}\text{+ 2e + 2s)}$$

$${\text{E = D·cos(a)+ (3D + r}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s)}$$

$${\text{E′ = D·cos(a) + (D + r}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s)}$$

Begrenzung des Bauraumes für den Lagerhülsenflansch 6Fl durch: $${\text{S′E′}}_{\text{min}}{\text{= M′E′}}_{\text{180°}}{\text{.sin(β) - (1/2.d}}_{\text{6f}}{\text{.cos(β) + 1/2 d}}_{\text{6f}}\text{+ s)}$$

und $${\text{r}}_{\text{Kb}}\text{- s =}\frac{{\text{(D + r}}_{\text{8s}}{\text{+ d}}_{\text{6l}}{\text{+ s)² + (r}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s)²}}{{\text{2(r}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s)}}\text{- s}$$

Circumferential line of the secant or rotary slide end points E refer to Cartesian coordinates: $$\text{(x² + y² - 2rx) ² - L² / 4 · (x² + y²) = 0 for L / 2> 4r}$$

Area F of the Pascal snail: $$\text{F = π · D² / 2 + π · L² / 4 D = 2r}$$

Center M of the control shaft 8 with the eccentric radius r and center of the external toothing with the pitch circle K _t in: $${\text{M}}_{\text{(xr yo)}}$$

Rotary piston center M ′ and center point of the internal toothing with the pitch circle K _t ′ in: $${\text{M ′}}_{\text{(xo yo)}}{\text{= P}}_{\text{(xo yo)}}$$

Intervention of the external toothing in the internal toothing in: $${\text{S}}_{\text{(x2r yo)}}{\text{= P}}_{\text{(x2r yo)}}$$

Approximation point of the inner surface of the hollow cylinder to the outer surface of the side wall of the rotary lobe in: $${}_{\text{(-xrDk yo)}}\text{= E180 ° rDk = rotary lobe radius}$$

Diameter D of the base circle K and diameter D _{t of} the pitch circle K _t $${\text{D = D}}_{\text{t}}\text{= 2r}$$

Diameter D _'of the center circle K _' and diameter D _{t 'of} the pitch circle K _t' $${\text{D ′ = D}}_{\text{t ′}}\text{= 2D = 4r}$$

Ratio of rotary valve length L to base circle diameter D: $$\text{L = 2Df + 2D f = proportionality factor}$$

Rotary piston diameter DDk: $$\text{DDk = 2Df = L - 2D = L - D ′}$$

Average rotary slide thrust speed c _m in the longitudinal groove 5: $${\text{c}}_{\text{m}}\text{= 2Dn / 30 (m / sec) n = rotary lobe rpm}$$

Ratio of the angular velocity w of the control shaft 8 to the angular velocity w 'of the rotary piston 4: $$\text{w = 2w ′}$$

Ratio of the path speed v _{p of} the secant or rotary slide center points P to the path speed v _{s of} the center circle intersection points S: $${\text{v}}_{\text{p}}{\text{= v}}_{\text{s}}$$

Circumference line E in polar coordinates: $$\text{E = DCos (a) + L / 2 for L / 2 ≧ 2D}$$

Circumference line E in parameter form: $$\text{x = D · cos² (a) + L / 2 · cos (a)}$$

$$\text{y = Dcos (a) sin (a) + L / 2sin (a)}$$

Area F _{0 °〉 ̶a} of Pascal's snail in polar coordinates: $${\text{F}}_{\text{0 °〉 ̶a}}\text{= D² / 4 · sin (a) · cos (a) + LD / 2 · sin (a) + (D² / 4 + L² / 8) · a / 360 · 2 · π}$$

Displacement V _{0 °〉 ̶a of} the rotary vane _machine : $${\text{V}}_{\text{0 °〉 ̶a}}{\text{= (F}}_{\text{0 °〉 ̶a}}\text{- π · rKr² · a / 360) .h}$$

Displacement V _n per rotary lobe rotation with n rotary valves per work area: F _Ds = rotary valve profile fl. $${\text{V}}_{\text{n}}{\text{= (F}}_{\text{0 °〉 ̶90 / n °}}{\text{- π · rKr² / 4n - F}}_{\text{Ds}}\text{) · 4 · n · h}$$

_{Circumferential} speed _maximum v _{Bmax of} the rotary vane end points E at E0 ° $${\text{v}}_{\text{Bmax}}\text{= f · (L + D ′) · Π}$$

Rotary slide thrust NE: NE = DCos (a) + D
Rotary vane _maximum thrust NE _max : $${\text{NE}}_{\text{Max}}\text{= NE0 ° = 2D = D ′}$$

Minimum rotary slide NE _min with n rotary valves per work area: $${\text{NE}}_{\text{min}}\text{= Dcos (90 / n) ° + D}$$

Establishing the relationship $$\text{L = 2D * f + 2D}$$

through the individual components of the rotary vane machine. Here is: $$\text{f =}\frac{{\text{2D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s + e}}{\text{D}}$$

$${\text{D}}_{\text{Dk}}{\text{= 2 (2D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s)}$$

$${\text{L = 6D + (<figure-callout id="8s" label="2r" filenames="imgf0005.png" state="{{state}}">2r</figure-callout>}}_{\text{8s}}{\text{+ 2d}}_{\text{6l}}\text{+ 2e + 2s)}$$

$${\text{E = DCos (a) + (3D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ e + s)}$$

$${\text{E ′ = DCos (a) + (D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s)}$$

Limitation of the installation space for the bearing sleeve flange 6Fl by: $${\text{S′E ′}}_{\text{min}}{\text{= M′E ′}}_{\text{180 °}}{\text{.sin (β) - (1 / 2.<figure-callout id="6f" label="d" filenames="imgf0005.png,imgf0007.png" state="{{state}}">d</figure-callout>}}_{\text{6f}}{\text{.cos (β) + 1/2 <figure-callout id="6f" label="d" filenames="imgf0005.png,imgf0007.png" state="{{state}}">d</figure-callout>}}_{\text{6f}}\text{+ s)}$$

and $${\text{r}}_{\text{Kb}}\text{- s =}\frac{{\text{(D + <figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}{\text{+ s) ² + (<figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s) ²}}{{\text{2 (<figure-callout id="8s" label="r" filenames="imgf0005.png" state="{{state}}">r</figure-callout>}}_{\text{8s}}{\text{+ d}}_{\text{6l}}\text{+ s)}}\text{- see}$$

Claims (1)

1. A rotary valve machine having geometry based on Pascal's limaçon with a housing constructed from a hollow cylinder (1) closed by means of two housing covers and having the internal surface profile of Pascal's limaçon, and a rotating piston (4) shaped as a hollow circular cylinder mounted within both the housing covers so as to rotate, with the axis of its rotation axial to the hollow cylinder at the origin of the coordinates, whereby a control shaft (8) which is also located axially and rotatably mounted in both housing covers passes through the rotary piston (4) with the central point of the axis of the control shaft (8) at the axial mid-point of the base circle (M), whereby bearing sleeves (61) of the rotary valves (6) with length L which is located in diametrical longitudinal grooves (5) of the rotary piston are mounted upon the eccentric segments (8S) of the control shaft (8) with the eccentricity of the base circle radius (r), and whereby the vanes of the rotary valve (6f) are fixed onto the bearing sleeve flanges (6F1) by means of webs (6s),
   characterized in that,

   - The length of the rotary valve (L) is determined according to the following relationship: $${\text{L = 6D + 2r}}_{\text{8S}}\text{+ 2d₆₁ + 2e +2s}$$

   

   In the formula, (D) is the diameter of the base circle, (r_8S) the radius of the eccentric (8S) of the control shaft (8) which holds the bearing sleeve (61), (d₆₁) the wall thickness of the bearing sleeve (61), (e) the minimum inwards travel of the vane (6f) of the rotary valve into the lateral wall of the rotary piston (4) and (s) is the minimum clearance between the bearing sleeve (61) and machine parts which are in motion relative thereto,

   - The profile of a bearing sleeve flange (6F1) is limited by an arc with a minimum clearance (s) to the internal surface of the lateral wall of the rotary piston and by the minimum lateral clearance (S′E′_min) between a vane (6f) of the rotary valve and the internal edge (E′) of an adjacent vane of the rotary valve, which is defined by the following relationship: $${\text{S′E′}}_{\text{min}}{\text{= M′E′}}_{\text{180°}}{\text{.sin(β) - (1/2.d}}_{\text{6f}}{\text{.cos(β) + 1/2 d}}_{\text{6f}}\text{+ s)}$$

   

   In the formula, (M′E′_180°) is the clearance between the mid-point (M′) of the rotary piston and the internal edge (E′) of the vane (6f) of the rotary valve in the 180° position, (β) is the 180° angle divided by the number (n) of rotary valves, and (d_6f) is the thickness of the vane of the rotary valve,

   - The profile of a bearing sleeve flange (6F1) is furthermore limited to the area within an arc (Kb) through the coordinates (E′90°, E′180°) and (E′270°) of radius (r_Kb) minus the minimum clearance (s): $${\text{r}}_{\text{Kb}}\text{- s =}\frac{{\text{(D + r}}_{\text{8s}}{\text{+ d₆₁ + s)² + (r}}_{\text{8s}}\text{+ d₆₁ + s)²}}{{\text{2(r}}_{\text{8s}}\text{+ d₆₁ + s)}}\text{- s}$$

   

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