One-piece crankshaft for piston machines, especially for internal combustion multi-cylinder engines.

0001The invention is based on the preamble of the independent claim from DE-OS 32 35 376.

0002The calculation method using shape numbers is used for the design of crankshafts for internal combustion engines of the known type. These shape numbers depend, among other things, on the parameters of cheek thickness and cheek width. In contrast, z. B. the cheek shape is not taken into account in the shape numbers. The shape of the cheek is largely determined by the external and internal transition of the crank arm from the crank pin to the bearing pin.

0003In a simple embodiment of the outside transition, it is known to provide flat bevels to reduce the mass of the crankshaft on the crank webs in areas of material that are not subject to high loads. Furthermore, it is known from the publication mentioned at the outset to make the outer transition of the crank arm from the crank pin to the bearing pin rotationally symmetrical, the initially moderately increasing cheek thickness increasing relatively rapidly as the bearing pin approaches the maximum value of the cheek thickness. Stress studies on such crankshafts show an unfavorable stress curve both in the crank arm itself and in the subsequent pins. On the other hand, high stress values ​​can be observed in the transition from the relatively weakly increasing cheek thickness to the relatively strongly increasing cheek thickness. The disadvantage of this is an unfavorable ratio of cheek mass to cheek stiffness.

0004The invention has for its object to improve the ratio of cheek mass to cheek stiffness such that a higher stiffness is achieved with a substantially the same cheek mass.

0005This object is achieved by the features specified in the characterizing part of the independent claim. The advantage of the invention is better material utilization with more mutually spaced voltage profiles with a significant reduction in voltage peak values.

0006Further advantageous embodiments of the invention are specified in the subclaims.

0007With the invention, an optimal ratio of cheek mass to cheek stiffness is achieved. This results in a reduction in the inertial forces with a lower weight of the crankshaft. Furthermore, due to the higher rigidity, the structure-borne noise excitation in the main bearings of the internal combustion engine is reduced and at the same time the bearing misalignment is reduced.

0008The invention is described with reference to an embodiment shown in the drawing. Show it:<ul id="ul0001" list-style="none"><li>1 is a partially shown, one-piece crankshaft for an internal combustion engine with half a crank in side view,</li><li>1 in front view with the outside transition from the crank pin to the bearing pin,</li><li>Fig. 3 shows the crank arm of FIG. 1 in an end view of the inside transition from the crank pin to the bearing pin.</li></ul>

0009In a one-piece crankshaft 1 for a multi-cylinder internal combustion engine, not shown, a crank pin 2 of radius r and a bearing pin 3 of radius R are arranged with their axes 4 and 5 in a lifting plane EH and through one with their width transverse to the lifting plane E._H directed crank arm 6 connected. The crank arm increases in thickness on the outside from the crank pin 2 to the bearing pin 3, the maximum cheek thickness on the crank arm 6 in the region of an overlap 7 of the crank pin 2 and the bearing pin 3 and / or of the respective pins 2, 3 assigned to the crank arm 6 " a "is given. The degree of coverage 7 is by a in the lifting plane E._H between the upper edge 10 of the crank pin 2 and the upper edge 11 of the collar 9 determined distance dimension "b" determined.

0010The crank arm 6 has an outside transition 12 from the crank pin 2 to the bearing pin 3, which is essentially defined by the curve "y" = (<maths id="math0001" num=""><img file="EP0243896A2_D0001.tif" /></maths>)ⁿ is limited, the exponent "n" being determined from the relationship: n = log b. For the coordinates y and x in Fig.<maths id="math0002" num=""><img file="EP0243896A2_D0002.tif" /></maths>on the crankshaft 1 a coordinate cross with the ordinate y and with the abscissa for x. The aforementioned curve "y" runs between a cheek thickness x = 2 (0.2 r) n determined at a distance y = 0.2 r from the upper edge 10 of the crank pin 2 and according to the aforementioned and modified equation and that from the upper edge 10 of the crank pin corresponding to a distance y = b on the upper edge 11 of the bearing collar-side thrust collar 9, the maximum cheek thickness "a", FIGS. 1 and 2. With this cheek shape an optimal utilization of the cheek mass is achieved to achieve a high stiffness against bending and torsion loading.

0011The transition 12 arched flat by the curve "y" is at the location of the cheek thickness x_WD delimited by an edge 13 which, as in FIG. 2, by the dashed line 13¹ is illustrated, can run across the entire width of the crank arm. Parallel to the edge 13 'is shown a dashed line 14 tangent to the run-up collar 9 on the upper edge 11, which represents a further edge when the transition 12 extends over the entire width of the crank arm. However, in order to keep the weight of the crankshaft low, the transition goes on both sides of the lifting plane E._H in a cone-shaped section and a rotationally symmetrical region 15 with respect to the axis 5 of the bearing journal 3, in such a way that the length of the edge 13 remains as a chord over the circle delimiting the crankpin 2 by the radius r.

0012For further weight reduction and for simpler manufacture, a transition 12 ′ designed as a flat bevel can be provided instead of the curved transition 12. The transition 12 'illustrated by a dashed line extends from that through the cheek thickness x_WD = 2 (0.2 r)<maths id="math0003" num=""><img file="EP0243896A2_D0003.tif" /></maths> fixed edge 13 and tangent to the upper edge 11 or a circumferential region of the thrust collar 9 on the bearing journal side.

0013Independently of the respective transition 12 or 12 ', the edge 13 is adjoined by an extension 16 which can be designed, the essentially radial extent of which is determined by the radius of the thrust collar 8 on the crank pin side.

0014For the determination of the exponent "n" from the relationship n = log<maths id="math0004" num=""><img file="EP0243896A2_D0004.tif" /></maths> b is given with the basis "a" an average value which belongs to a value range of a (0.45 - 0.55).

0015Furthermore, the crank arm 6 has a decreasing thickness on the inside from the crank pin to the bearing pin 3, part of this taper being brought about by a flat bevel 17. This slope 17 is, as can be seen from the side view of the crankshaft 1 in FIG. 1, with a in the lifting plane E._H extending connecting straight line 18 rectified. On the one hand, the connecting straight line 18 cuts the bearing journal 3 over half the length 1/2 on the lower edge 19 and, on the other hand, it affects the crank pin-side thrust collar 8 on its lower edge 20 or its peripheral region. The bevel 17 also merges in the region of the axis 5 of the bearing journal through a rounding 21 which extends over the width of the crank arm 6 into a counterweight 22 arranged on the crank arm 6. Close to the bevel 17, as can be seen from FIG. 3, on both sides of the lifting plane E._H designed as cone-shaped sections and to the axis 4 of the crank pin 2 rotationally symmetrical areas 23, in such a way that the connection point 24 of the fillet 21 to the bevel 17 remains in a width corresponding approximately to the diameter dimension 2 R of the bearing pin. Thus, the crank arm 6 is also optimally designed on the inside in terms of weight and rigidity, so that overall an improvement in the static and dynamic properties of the crankshaft 1 is achieved.