WO2003104666A1

WO2003104666A1 - Crankshaft drive, in particular for reciprocating engines

Info

Publication number: WO2003104666A1
Application number: PCT/EP2003/006026
Authority: WO
Inventors: Math Lemoure
Original assignee: Lemoure, Suheyla
Priority date: 2002-06-11
Filing date: 2003-06-10
Publication date: 2003-12-18
Also published as: DE10226672B3

Abstract

The invention relates to the regulation of the maximum compression point and/or the upper and lower dead-centre position of a crankshaft drive in a reciprocating engine for operating vehicles, in addition to the operation of stationary engines. Conventional crankshaft drives achieve the maximum compression point at the upper dead-centre when the crank angle α = 0°. The combustion chamber pressure reaches its maximum shortly after the upper dead-centre. The torque acting on the crankshaft depends on the crankshaft position and is zero at the upper dead-centre and extremely low close to the upper dead-centre. At the moment when the combustion chamber pressure achieves its maximum, the corresponding position of the crankshaft produces unfavourable geometric angular ratios of the crankshaft drive, so that only a fraction of the combustion chamber pressure can be converted into effective torque for the crankshaft. According to the invention, the maximum compression point of the crankshaft drive can be set by means of the crankshaft journal sections (1, 2, 3), the central point displacement (40) and the crankshaft journal angles, in such a way that an increased effective torque can be achieved for the crankshaft compared to a conventional crankshaft drive.

Description

Crank drive, especially for reciprocating internal combustion engines

State of the art

The invention relates to a crank mechanism of a reciprocating internal combustion engine according to the preamble of the main claim. Such a crank mechanism is known from the literature, for example from <1> and from <2>. The crank mechanism serves to convert the oscillating movement of the piston into a rotary movement. The sizes required for the kinematics of the crank mechanism are the crank radius (r), the

Push rod length (1) and the crank angle (α). The piston travel (S ₀ ) for the unrestricted crank drive is expressed by the following equation:

According to the above equation, the maximum compression at top dead center (crank angle α = 0 °) of the piston in the cylinder is achieved. The fuel is injected directly into the combustion chamber shortly before compression (e.g. in the case of direct-injection diesel engines). After the fuel-air mixture has ignited, the pressure in the cylinder rises rapidly. The pressure curve in the cylinder reaches a maximum shortly after top dead center. The torque acting on the crankshaft depends on the crankshaft position and is defying the very high combustion chamber pressures at top dead center zero and very small near top dead center with small crank angles (α). The two forces acting on the piston in the cylinder direction are the gas force and the oscillating mass force. Your resulting is the piston force (F _R. ). With the crank angle (α), the swivel angle (ß) and the connecting rod ratio (λ = r / 1) the forces in

Calculate the crank mechanism as a function of the gas force, where sin (ß) = λ * sin (α) corresponds. To determine the torques (Ma) acting on the crankshaft, the Tangential forces (F _τ ) are multiplied by the constant crank radius (r). The

Tangential forces (F _τ ) and the torque (Md) acting on the crankshaft are calculated as follows:

Tangential force _{= F.} * " ⁽ <* + ß ⁾ ^{r τ Γ} « COSQS)

Torque M = F • ^{sin (α + ß)} »r ^{lvl dr κ} COSQS)

The torque (M _d ) is therefore the product of the time course of the resulting piston force (F _K ), the geometric ratio " ^sιn ^ ^{α +} Pl" and cos (/ 3) the crank radius (r). The factor crank radius (r ) is independent of the

Piston position constant. The pressure curve in the cylinder and thus the piston force

(F _K ) reaches a maximum shortly after top dead center, is above the

Piston position changeable and decreases rapidly with increasing crank angle (α) after reaching the maximum. The factor " ^sin ^ ^{α +} " is derived from the cosQS) geometry of the crank drive used, is zero at top dead center and increases approximately linearly up to approx. Crank angle α = 40 °. Now consider the variation of the piston force over time ( F _K ) and the geometric

Factor " ^{sin q +} " it becomes clear that with increasing crank angle (α) the cos (j8) behave in opposite directions over the course of the course. That is, at the moment when the combustion chamber pressure and thus the piston force (F) unite shortly after top dead center The geometric factor " ^sin ^ ^{α +} ^" has reached the maximum in the same cos S)

Moment or at the same crank angle () a relatively small value and is zero at top dead center. This means that in and near top dead center

Gas force strained all components and assemblies of the engine only in tension and the torque on the crankshaft reached an insignificant value. With increasing crank angle (α) the factor " ^sin ^ ^{α +} ^" cos (/ 3) increases but at the same time the gas force and thus the piston force (F _K ) decrease rapidly.

Due to the relationships described above, this becomes the crankshaft

Available torque (Md) due to the unfavorable temporal overlap the combustion chamber pressure with the angular ratios of the connecting rod and the crankshaft. As a result, only a fraction of the energy supplied to the combustion chamber is converted into usable torque (Md) on the crankshaft.

Advantages of the invention

The object of the invention is to provide a crank mechanism of the type characterized in the preamble of claim 1, which ensures an increased usable torque (Md) on the crankshaft.

This object is achieved according to the invention by the characterizing features of claim 1.

By changing the crank drive geometry according to the invention, the maximum compression point can be set in such a way that crank torque conditions which are more favorable for the torque (Md) are already present during the start of combustion than in the crank drives present today. About the size of the

The center offset (40) and the crank pin angle (δ) and the corresponding assignment of the crank pin sections (1,2,3) and the connecting rod eye sections (25,26,27) can be used to set the top dead center position of the piston in the cylinder and thus the maximum compression point that it is not achieved with crank angle (α) = 0 ° as before, but for example only with crank angle (α) = 8 °. This new top dead center position of the piston at crank angle (α) = 8 ° is referred to below as the “new” top dead center. That is, the conventional piston travel (So) can, for example, depending on the crank angle (α), center offset (40) and crank pin angle (δ) can be shifted in the direction of larger or smaller crank angles (α). As a result, the injection timing of the fuel can also be adjusted in accordance with the "new" top dead center, as a result of which the combustion and the maximum combustion chamber pressure in the cylinder also increase shortly after new "top dead center. The geometrical Ratios of the crank mechanism in the area of the "new" top dead center for converting the combustion chamber pressure into torque (Md) on the crankshaft are considerably more favorable. This leads to the fact that in the "new" top dead center

Torque (Md) is no longer zero, but results from the respective combustion chamber pressure in the "new" top dead center, from the crank radius (r) and from the crank pin angle (Ö) chosen accordingly and thus from the resulting angle relationships with respect to the crank angle (α) With the crank mechanism according to the invention, a higher usable torque (M _d ) can thus be achieved on the crankshaft from a given pressure curve in the cylinder than with a conventional crank mechanism.

The fuel consumption of an internal combustion engine can also be significantly reduced with the crank mechanism according to the invention, for example by for a certain torque (Md), a smaller combustion chamber pressure or a smaller amount of fuel is already sufficient to achieve the same output at the same speed as in an internal combustion engine with a conventional crank mechanism. In this context, the smaller amount of fuel can also reduce the exhaust gas emissions accordingly, or more easily comply with the emission limits prescribed by law.

Furthermore, with the crank drive according to the invention, further possible combinations with regard to torque (M _d ), fuel consumption and exhaust gas emissions can be represented by the corresponding selection of the center offset (40) and the crank pin angle (δ).

drawing

The invention will be explained in more detail with reference to the exemplary embodiments shown in the drawings:

FIG. 1 shows a crankshaft for an internal combustion engine, for example a motor vehicle. The internal combustion engine is, for example, a self-igniting or spark-igniting internal combustion engine and has one or more cylinders.

A crankshaft for a single-cylinder internal combustion engine is shown consisting of the two crankshaft bearings (11, 41), the two crank webs (9, 10) and a crank pin, which in turn consists of three crank pin sections (1, 2, 3) arranged offset from one another. Also shown is the crank radius (r) which results from the distance between the crankshaft axis (13) and the crankpin axis (12).

FIG. 2 shows the top view of the crankshaft based on the front view of FIG. 1. Shown are the lower crank pin halves (7, 8, 38), the upper crank pin halves (4, 5, 6), the respective crank pin offset (42, 43, 44, 45) and the crank pin width (30,31,32) of the crank pin sections (1,2,3).

FIG. 3 shows the side view from the left of the crankshaft, with a section through the crank pin section (2), based on the front view of FIG. 1. The crank pin angle (δ) here has an amount of 90 °, based on the crank arm axis (14) , The sliding or contact surfaces of the crank pin sections (1, 2, 3) with the respective inner surface of the connecting rod eye sections (25, 26, 27) take place via the crank pin tips (34, 35), (the crank pin tip of the crank pin section (1) is shown in FIG. 3 not shown), and thus over the crank pin width (30,31,32). To a high

To avoid surface pressure on the crank pin tips (34, 35) and on the inner surfaces of the connecting rod eye sections (25, 26, 27) especially during the gas and inertia forces, the radii of the crank pin tips (34, 35) should be avoided the crank pin width (30,31,32) should be dimensioned accordingly. In addition, the lower crank pin halves (7,8), the upper crank pin halves (4,5), the crank pin offset (42,43) and the

Crank pin main axis (33) shown.

FIG. 4 shows the side view from the left with a section through the crank pin section (2) of the crankshaft based on the front view of FIG. 1 but with a crank pin angle (δ) = 30 °.

FIG. 5 shows the connecting rod (16) belonging to the crankshaft according to FIGS. 1 to 4 in the front view. This shows the connecting rod eye sections (25, 26, 27) with the corresponding connecting rod eye radii (20, 21, 22), as well as the common connecting rod eye center (28) and the circular center line (23), the connecting rod eye sections (25, 27) and the connecting rod eye radii ( 20, 22) are the same size in this exemplary embodiment. Furthermore, the push rod length (1) and the small connecting rod eye (15) are shown.

FIG. 6 shows in half section the side view of the connecting rod (16) in relation to the front view of FIG. 5. The connecting rod eye sections (25, 26, 27) with the corresponding connecting rod eye radii (20, 21, 22) and the connecting rod eye axis (37 ).

FIG. 7 shows an assembly drawing of the crank drive according to the invention according to FIG. 6 and FIG. 1, but with crank pin angle (δ) = 0 ° and with a crank angle (α) = 0 °. The eccentricity with respect to the connecting rod eye center (28) and the crank pin center (36) and thus between the connecting rod eye axis (37) and the crank pin axis (12) by the amount of the center offset (40) results from the fact that the upper crank pin half (4) is larger than that associated connecting rod radius (21) of the connecting rod section (26). The lower crank pin halves (7, 8, 38) of the crank pin sections (1, 2, 3) are smaller than the connecting rod radii (20, 21, 22) and can be dimensioned so large that, on the one hand, the free run of the eccentricity with respect to the connecting rod center ( 28) and crank pin center (36), or between the connecting rod eye axis

(37) and the crank pin axis (12) depending on the crank angle (α) is ensured, and on the other hand to achieve high strength on the crankshaft, in particular at the respective transitions of the crank pin sections (1, 2, 3).

Since the total height, upper crank pin half (4) plus lower crank pin half (8), of the crank pin section (2) is smaller than the diameter of the connecting rod eye section (26), at least one additional connecting rod eye section (25) is required for guiding the connecting rod (16), whereby from Another connecting rod eye section (27) is used here for reasons of symmetry. With the

Connecting rod eye sections (25, 27) can ensure that, for each crank angle (α), both the upper crank pin half (4) with the connecting rod eye section (26) and the two upper crank pin halves (5, 6) with the connecting rod eye sections (25, 27) are always in contact and therefore also the

Sliding guidance of the connecting rod (16) is guaranteed. The two connecting rod eye sections

(25,27) are in this

Embodiment equal in diameter and also have the same connecting rod eye center with the diameter of the connecting rod portion (26)

(28). The connecting rod eye radii (20, 22) of the two connecting rod eye sections (25, 27) are larger than the associated upper crank pin halves (5, 6).

FIG. 8 shows a further assembly drawing of the crank drive according to the invention according to FIG. 3 and FIG. 5, however with crank pin angle (δ) = 0 ° and with a crank angle (α) = 0 °. The maximum eccentricity between the connecting rod center line (23) and the crank pin axis (12) is in the direction of the cylinder axis (39) at crank angle (α) = 0 ° and corresponds to the center point offset (40). In this exemplary embodiment of the crank drive according to the invention, the top dead center position of the piston, as also in the conventional crank drives, is reached at a crank angle (α) = 0 °, but in this embodiment the maximum piston stroke (s) is twice the center point offset (40). increased. FIG. 9 shows an assembly drawing of the crank drive according to the invention according to FIGS. 3 and 5, but with a crank pin angle (δ) = 0 ° and with a crank angle (α) = 180 °.

FIG. 10 shows an assembly drawing of the crank drive according to the invention according to FIG. 3 and FIG. 5 at a crank angle (α) = 20 °.

FIG. 11 shows the respective course of the piston travel of the conventional crank drive and of the crank drive according to the invention as a function of the crank angle (α). A crank radius (r) = 35 mm and a push rod length (1) = 125 mm were used for both the conventional and the crank drive according to the invention, with the crank drive according to the invention also having a center offset (40) = 2.5 mm and two different crank pin angles (δ) = 0 ° or white 90 ° were taken into account. The long dashed curve with the designation "new crank mechanism with delta = 90 °" is shifted by a few crank angles (α) with respect to the un dashed curve "conventional crank mechanism" in the direction of larger crank angles (α). The short dashed curve with the designation "new crank drive with delta = 0 °", however, is only increased in amplitude by the amount of the center offset (40), but has the same phase as the un dashed curve "conventional crank drive".

FIG. 12 shows an enlarged section of the diagram according to FIG. 11. The course of the crank drive according to the invention and thus the differences from the conventional crank drive are illustrated on a larger scale.

FIGS. 13 to 20 show further exemplary embodiments of the invention according to FIGS. 3 and 5 and are explained in more detail in the description below, the crank angle (α) being varied from 0 ° to 360 ° crank angle. Description of the embodiment

If, for example, the crank pin section (2) is the smaller in diameter

Connecting rod eye section (26) and, corresponding to the crank pin sections (1,3), the other two larger diameter connecting rod eye sections (25,27), the connecting rod eye center (28) is eccentric to the crank pin center point (36) at a crank angle (α) = 0 ° , The connecting rod eye center (28) and the crank pin center (36) are in the horizontal direction around the

The center point offset (40) is offset from one another, these two center points lying at the same height in the direction of the cylinder axis (39). As a result, the connecting rod (16) is offset at the crank angle α = 0 ° by the swivel angle (βo) with respect to the cylinder axis (39), cos (βo) resulting from the ratio of the connecting rod length (1) to the center offset (40). With increasing crank angle (α) the crank pin or the crank pin sections (1, 2, 3) rotates with the crank pin center (36) at a distance of the crank radius (r) around the crankshaft axis (13), and at the same time the crank pin sections (1, 2, 3 ) around the crank pin axis (12) and in the associated connecting rod eye sections (25, 26, 27). While from the crank angle (α) = 0 ° the crank pin or crank pin sections (1, 2, 3) and the connecting rod (16) rotatably mounted on the crank pin sections (1, 2, 3) move downward in the direction of the cylinder axis (39) carry out, the connecting rod (16) and the piston which is generally mounted on the small connecting rod eye (15) additionally experience a movement directed upwards through the center point offset (40). This

Relative movement or this variable distance in the direction of the cylinder axis (39) between the connecting rod eye center (28) and the crank pin center (36) reaches the maximum eccentricity at a crank angle (α) = 90 ° or at 270 °, this maximum eccentricity being exactly the center point offset (40 ) corresponds. From a crank angle (α)> 90 ° in the direction of the cylinder axis (39) the eccentricity between the connecting rod eye center (28) and the crank pin center (36) becomes smaller and smaller until the crank angle (α) = 180 ° the eccentricity is zero or more horizontal Direction again becomes maximum. <1>"Construction and calculation of internal combustion engines", Springer-Verlag, edition 1983, page 72. <2>"Engines of snow-running internal combustion engines"

Springer-Verlag, edition 1966, page 59.

Claims

Expectations

1. Crank drive of a reciprocating internal combustion engine with a crankshaft which has at least one crank pin on which a connecting rod with its large connecting rod eye is rotatably mounted, characterized in that each crank pin has at least two crank pin sections offset radially to the crank pin axis (12) and arranged axially to one another ( 2,3), and that the axes of the respective crank pin sections (2,3) each have any

Form crank pin angle (δ) with the crank cheek axis (14), and that the large connecting rod eye of the connecting rod mounted on each crank pin has at least two connecting rod eye sections (26, 27) which have the same connecting rod eye axis (37) and a different diameter and are arranged axially to one another.

2. Crank drive according to claim 1, characterized in that at least one of the connecting rod eye radii (21, 22) is smaller than one of the upper crank pin halves (4, 5) and at least one of the connecting rod eye radii (21, 22) is larger than one of the upper crank pin halves ( 4,5), the respective sequence and assignment of the connecting rod eye sections (26, 27) with the associated crank pin sections (2, 3) being arbitrary.

3. Crank drive according to one of claims 1 or 2, characterized in that both the shape of the offset crank pin sections (2,3) and the crank pin width (31,32) and the connecting rod eye width (18,19) are arbitrary. Case number: 10226672.7-13 Dipl. Ing. M. LEMOURE

4. Crank drive according to one of claims 1 to 3, characterized in that the sizes of the respective crank pin offsets (42, 43) of respectively adjacent crank pin sections (2, 3) both in the direction of the crank pin main axis (33) and in the remaining circumferential direction of the crank pin axis ( 12) can be freely selected.

5. Crank drive according to one of claims 1 to 4, characterized in that the size of the center offset (40) by the size difference of the connecting rod eye radii (21, 22) and the upper crank pin halves (4,5) is given and can be selected as desired.

6. Crank drive according to one of claims 1 to 5, characterized in that both the diameter of the connecting rod eye sections (26, 27) and thus also the

Conrod eye radii (21, 22) and the sizes of the upper crank pin halves (4, 5) can be selected as desired.