FR2470284A1 - SYMMETRICAL SEWING CRANKSHAFT - Google Patents

Abstract

THE CRANKSHAFT INCLUDES 1 TENSIONERS AND 2 CRANKSHAFTS WITH PASSAGE RADIUS, AND 3 FLANGES, THE SHAPE OF WHICH IS DETERMINED BY AN OPTIMIZED TENSION DISTRIBUTION, GIVING AN INVARIABLE RELATIONSHIP TO THE PASSAGE RADIUS OF THE CRANKSHAFT AND OF THE journal WITH RESPECT TO THE THICKNESS OF THE FLANGE, WHICH CAN BE REPRESENTED BY A FUNCTION WHICH IS PROPORTIONAL TO THE STRAIN IN THE CRANK PASSAGE RADIUS, THE RATIO BETWEEN THE COVER S OF THE CRANKSHAFTS AND THEIR DIAMETER D TO MEET 0.14SD0.26.

Description

The invention relates to a symmetrically bent crankshaft, comprising

journals, crankpins and

  crankshaft flanges, in which the passages of the

  crickets and crankpins towards the flanges are formed with passage radii, and where the distance from a center of touril-

  to another is determined by construction reasons

tion, and the widths of the pins and crankpins are deter-

undermined by reasons for bearing demands, the

  guration of the crankshaft being determined according to a

  optimized distribution of constraints.

The dimensioning of a crankshaft also implies knowledge in terms of stress distribution in the bends and allows, in connection with the resistance to alternating forces depending on the material, the size

and technological influence, a judgment on safety

  tee against fatigue rupture. Resistance to alternating forces is defined by the choice of material, the type of forging and the size of the projected construction. It is however different with the distribution of the constraints, which can be significantly influenced by the different

bend parameters, such as lap overlap

  with the crank pins, the thickness and the width of the flas-

ques, as well as the radii of passage of the trunnions and ma-

netons towards the flanges.

  For the determination of the stress peaks forming in the cited passage radii, we now know several approximation methods, determined empirically. Reference is made in this context in particular to MTZ 23 (1962) 12, to MTZ 29 (1968) 3, to brochure 49/1965 of the Association for Research on Internal Combustion Engines (Forschungsvereiniqunq Verbrennunqskraftmaschinen), and to the research report of

  this association, brochure 130 (1972).

  All the processes described there are based on the determination of the form factors for bending and

  torsion as an expression for the relationship between stress-

peak te and nominal stress, ie a max / anominal,

respectively a max / a nominal. Depending on each pro-

assigned, the nominal load is reported either to the ma-

  neton, or in the flange section. In accordance with the pro-

  sold described in the research report - brochure 130 -,

  stress distribution in city bends-

  brequin of fast motors, characterized by thin flanges, small piston strokes and by

following by large crankpins overlaps was examined

  born on a large number of models. We also determined-

undermined inter alia the influence of bores for the turret-

  lon and the crankpin, as well as the influence of the bias of the flas-

ques, on the stress peaks. This process also

for the first time takes into account the constraints in the

trunnions, as well as types of stresses, such as pure bending, bending with transverse force and torsion. The search result represents the state on

more current technique in terms of determination

  stress nation in crankshaft bends and,

for this reason, will serve as a basis for the present invention.

  As influence parameters, account was taken of the crankpin diameter, the crankpin overlap, the width and thickness of the crankshaft flange as well as the radii of passage between pin or crankpin and flange. The shape indices are determined using a multiplicative sequence of functions of these parameters related to the crank pin diameter, the nominal stress,

  it, being related to the section of the crankshaft flange-

quin.

When sizing a crankshaft, the gap

  the cylinder size is determined by the barrels and the stroke

  pistons, which are decisive for the real power

engine sand. In addition, the necessary width of the crankpin

bp and the width of the journal b G result from the stress

calculated level of bearings. Finally the abutment flanges ap and aG with the crank pin and the pin are determined by

  depending on the desired machining status for the flange of the vile-

brequin. The design engineer therefore has the parameters

  of influence cited. Regardless of the spacing of the cylinders

24702S4

  dres, we can then still choose the diameter of the handle-

  ton dp, the overlap s of the crankpins and the width of the flange b it being understood that an increase in these values

  always generates a reduction in absolute constraints.

  It should also be pointed out that the size of the diameter of the handle dp is limited by the requirement of the possibility of

  placing the connecting rod through the cylinder barrel.

  Unlike the three values dp, s and b, the remaining parameters, namely the passage radii rp at

crankpin, the spokes of passage rG to the trunnion and the thick-

s flange w, are dependent on the predeter spacing-

mined cylinders respectively of length 1, going from one trunnion center to the other, as well as determined widths bp and bG. There remains therefore a length k = 1 - bG - bp - 2 (aG + ap). At the same time the length

is calculated k = 2 (w + rp + r).

Using known functions for determining

  tion of the shape index and the stress, it is recognized that each time it is only an increase in the radii of passage rp and rG as well as the thickness of the flange w which results in a regression of the points of constraint. Given however that the three parameters are solidly

  attached to each other by the length k, an increase

  tion of one of these parameters means at the same time

decrease of others. It is for this reason that the de-

termination of the three quantities under the aspect of minimization

constraints cannot be done on the basis of the

state of the art.

  The object of the invention in the design of a city

brequin of the type defined at the beginning and independently of the other parameters, is to obtain a minimization of the stresses in the bends only thanks to the corresponding choice of the thickness of the crankshaft flange and the radii of

switch to crankpin and journal.

In accordance with the present invention, this result

is obtained by the fact that we determine an invaria-

ble between the radii of passage of the crank pin and the trunnion on the one hand, and the flange thickness on the other, which can be represented by a function derived from the magnitudes rp, rG and w, proportional to the stress in the radius of passage of the crank pin, the ratio between the overlap of the crank pins and their diameter having to be greater than or at most equal to 0.15 and less than or at most equal to 0.26. According to the invention, the function resulting from the three quantities: flange thickness, radius of passage of the crank pin and radius of passage of the journal is written f (w, rp, rG) = {w 1'2829 (bG / 2 + rG) + 0.5231,

  bG being the span width of the journal.

  As the torsional stresses in the mobile coupling are not yet detectable at the time of design and can be influenced by torsional vibration dampers, the invention

  is limited to bending stress, while taking into account

te however in an engine running the case of

load "bending stress and transverse force". This-

the way you do it seems reasonable, since crankshaft incidents, as experience has shown, come from

  mostly from fractures due to bending.

Optimal stress distribution is already

important for the sole reason that in vehicle engines

current high-performance vehicles, the type of lightweight construction and therefore small dimensions

foreground. It is essential to make the three para-

meters cited rp, rG and w independent of the other parameters,

  since their influence on the appearance of stress-

your is undeniable.

As we can see from the known functions,

there is no relation between the width of the city flange-

  brequin b and the three sizes to be optimized. Conversely, however, there is a dependence between the thickness of the flange w and the overlap of the crank pins s, since the two appear as reported quantities w '* = w / d and s * = s / de p

in a double function f (s *, w *). To maintain the

  fluence of the thickness of the flange w on the double function f (s, w ') as minimal as possible, respectively s / de are limited as already mentioned. Moreover, we determine the stress a p in the radius of passage rp of crankpin

  for this reason that this is where the constraints of trac-

tion are the highest.

For the described minimization of the constraint ap, it is also necessary to have the following ratio: rp / rG = c rp is, as already indicated, the radius of passage in the crank pin, and rG the radius of passage in the journal. c represents a

size can be freely chosen in certain li-

mites and is roughly between 1.0 and 1.9. It allows a weighting between the maximum tensile stress

  on the one hand, and the con

maximum compression strap in the passage radius of the

pin on the other hand.

  In addition, from the length k defined above

  above, we obtain the following relation, relating to the thickness of the flange w: k w = k rp (1 + l / c) We can thus determine for each quantity c

  selected, the optimal passage radius rp with the crankpin in function

tion of the length k. Thanks to the relationships set out above

  sus, the thickness of the flange w and the passage radius rG at

  journal are also determined.

  If we substitute w and rG and take the deri-

vee on rp, we always obtain in this way a distribution

  optimal constraints, and the three quantities can-

can be determined from the equation of rp.

For the sake of simplification, we will admit three factors a1, a2 and a3: al. r t + rpt + a2 + a3 = 0 Popt 2 3pp a1 = 0.403 (2 - 1) c k 1 a = 1, 4145. k + 0.66455 x kr 0.903. bG (1 +) 2 c c a3 = -0.26155. bG. k - 0, 26155. k2 When c = 1, we obtain as solution: rpopt = -3 a2 When c is greater than 1, we obtain an optimal radius of passage at the crankpin of r = -0.5 a- _ 4 a 3a rpopt = 0, a3 a1 The invention is explained below in more detail with reference to the attached drawings which show: FIG. 1 - part of a crankshaft, seen from the side Fig. 2 - a section along line II-II through the crankshaft of fig. 1, Fig. 3 and 4 - diagrams on the change of stress ratios,

Fig. 5 and 6 - diagrams on step radii -

  wise of crankpins, to a size c determined.

Fig. 1 allows above all to explain the desi-

  gnations used for the dimensions of the crankshaft. The

  part of the crankshaft shown consists of two half

  pins 1, in a crank pin 2 and in two flanges 3 of vile-

brequin. The pins 1 have a diameter dG and passage radii rG, which open onto stop flanges 4, of a thickness aG provided on the flanges 3. The span width of the pin 1 is designated by bG. For its part, the crank pin 2 has a diameter dp and its ends

also emerge via pas-

sage rp on stop flanges 5, provided on the

  flanges 3 and having a width ap. The width of door

pin 2 is designated by bp. The gap between the center

  of a pin 1 and the center of the adjacent pin is de-

signed by 1.

  In the context of the present invention, one poses for the length k not mentioned in FIG. 1:

  k = 1b - bG bp - 2 (aG + ap) = 2 (w + rp + rG), w signi-

trusting the thickness of the flange. Finally k is also calculated

from k = 2 (w r + rp + rG).

  According to fig. 2, the width of the flange is designated

by b.

  The diagram in fig. 3 shows the con-

drag between the journal passage and the crankpin passage.

  On the abscissa 6 is the magnitude c, which can be freely chosen, and constituted by the ratio "radius of passage of the crank pin / radius of passaae of the journal". In ordinate 7 is 247c2_4 is represented the report "compression stress

within the radius of passage of the journal / draw constraint

tion in the radius of passage of the crankpin ", therefore -aG: aP.

  We can recognize on line 8 that the ratio -aG lap increases according to the high choice of the value. Fig. 4 represents the reduction in stress

  in the radius of passage rp of the crankpin 2, bringing again

  calf on the abscissa 9 the value c, and on the ordinate 10 the diminu-

  constraint tion a p in the passage radius rp in percent.

  This results in a curve 11. The figures clearly demonstrate

ment that the tensile stress ap in the pass radius-

sage rp of crankpin 2, with increasing value c, decreases in

  favor of the compressive stress -a G in the radius of

passage rG of the journal.

  From the length k already mentioned, we obtain the following relation: w = k / 2 - rp x (1 + 1 / c)

We can thus determine, from each value

their c chooses the optimal radius of passage rp of the crankpin 2 as a function of the length k. These reports then allow

  also determine the thickness of the flange w and the

yon of passage rG of journal 1.

Figs. 5 and 6 show by way of example the optimal radii of passage rp of crank pins 2, namely in FIG. 5 by choosing the value c = 1.0 and in fig. 6 with a value c = 1.1. On the abscissa 12 is the length k an mm in both cases, and on the ordinate 13 the passage radius rp also in mm. In the two figures the upper lines 14, 14 ′ represent the optimal radius

passage rp, while the lower lines 15.15 'mon-

  trent the minimum passage radii r. The field between the two Lines 14, 15 and 14 ', 15' gives the design engineer the

  possibility, at maximum stress increase of only 2%

compared to the optimum, to achieve a decrease

  stress of the main bearing due to a lower

all crankpin width, consisting of bp + 2 rp, that is

  say this results in a decrease in the rotating masses.

  With the choice of sizes w, rpet r., We ensure

2470234!

  therefore an optimal stress distribution. To get

provide the necessary safety with regard to fatigue failure,

  the absolute level of stress can be easily influenced

inked with the remaining parameters, such as the journal and crankpin diameters dp, dG, crankpin overlap s and flange width b. So the way is set

  to aim for a separate configuration of the

bile, as studied, in a context of light construction.

  It remains to point out that the optimal pas-

  wise, as shown in Figures 5 and 6, can of course be determined in the same way for all

  the "c" values chosen, therefore up to around c = l, 9.

2470ú2 4

Claims (6)

  1. Crankshaft with symmetric bend, comprising   trunnions, crankpins and flanges, where the steps   wise of the journals and the crankpins towards the flanges are formed with passage radii, and where the distance of a   journal center to center is determined by rai-   construction sounds, while the widths of the turrets   lons and crankpins are for reasons of bearing stress, the crankshaft configuration being determined by an optimized distribution of stresses, crankshaft   characterized by the fact that we give an invaria- ble to the radii of passage (rp) of the crankpin and (rG) of the rillon with respect to the thickness of the flange (w), which can be represented by a function (f) of the quantities rprG and w, which is proportional to the stress in the radius of the crankpin, the ratio between the overlap (s ) crankpins and their diameter (d%) must be greater than or at most equal to 0.15, and less than or at most equal at 0.26 (0.14 e s / dp 0.26).   2. Crankshaft according to claim 1, character-   laughed at by the fact that the function (f) of the three quantities: flange thickness (w), radius of passage (rp) of the crank pin (2) and radius of passage (rG) of the journal (1) is given by: f ( w, rp, rG) = {w-1'2829 (bG / 2 + rG) + 0.5 w-0.2829} rp -0.5231 (bG) representing the span width of the pin (1).   3. Crankshaft according to one of claims 1 or 2,   characterized by the fact that a quantity (c) which can be bound brely chosen within a certain limit and representing the ratio between the radius of passage (rp) of the crank pin (2) and the radius of passage (rG) of the journal (1), allows the weighting between the constraints in the radius of passage (rp ) and the radius of passage (rG), and thereby the determination of a small   crankpin diameter (dp), (c = rp / rG).   4. Crankshaft according to claim 3, characterized in that the ratio (c) between the passage radius (rp) of the crank pin (2) and the passage radius (rG) of the journal (1) 2470240 ' is between 1.0 and 1.9.   5. Crankshaft according to one of claims 1 to 4,   characterized by the fact that at each gear (c), the three quantities necessary for the optimization of the constraints, either radius of passage (rp) of the crank pin (2), radius of passage (rG) of the journal (1) or thickness flange (w) can be determined exactly by taking the derivative according to (rp). 6. Crankshaft according to one of claims 1 to 5,   in which one seeks to decrease the width of the crankpin to decrease the masses in rotation and thereby discharge the main bearings, characterized in that the width   of field of the radius of passage (rp) of the crankpin (2) is deter- p   mined so as to be able to admit an elevation of con- traces of only 2% in the passage radius (rp) crankpin (2).