GB2064719A

GB2064719A - Crankshaft with symmetrical crank-throws

Info

Publication number: GB2064719A
Application number: GB8038090A
Authority: GB
Original assignee: MAN Maschinenfabrik Augsburg Nuernberg AG
Current assignee: MAN AG
Priority date: 1979-11-27
Filing date: 1980-11-27
Publication date: 1981-06-17
Also published as: IT8025911A0; FR2470284A1; JPS5686222A; GB2064719B; HU183243B; FR2470284B1; IT1134223B; TR21149A; SE8008341L; DE2947699C2; DE2947699A1

Abstract

In order to minimize stresses in the crank-throws there is a fixed relationship for the fillet radii rP of the crank-pin (2) and rG of the journal (1) to the crank web thickness W as represented by a function f of the magnitude rP, rG and w. This function is proportional to the stress at the fillet radius of the crank-pin (2). The ratio of pin/journal overlap s to the crank pin diameter dP is equal to or greater than 0.15 and less than or equal to 0.26 (0.15</=s/dP = 0.26). <IMAGE>

Description

SPECIFICATION Crankshaft with symmetrical crank-throws This invention relates to a crankshaft having symmetrical crank-throws comprising journals, crank pin and crank-webs, in which the fillets from the journals and from the crank pins to the crank-webs are formed by fillet radii and in which the distance between the journal centres is predetermined by physical reasons, and the bearing widths and the big end bearing widths are predetermined for reasons of bearing stressing, the shape of the crankshaft being determined by optimization of the stress distribution.

Dimensioning of a crankshaft also calls for information on the stress distribution in the crankthrows and, in conjunction with the fatigue strength, being a function of the material, size and technological factors, permits an assessment of its safety against fatigue failure. The fatigue strength is given by the selection of the material, the method of forging and the size of the proposed design. A different case exists regarding the stress distribution which can be substantially influenced via the various crank-throw parameters, such as the overlapping of the journals with the crank pins, the thickness and width of the crank-webs as well as the fillet radii from the crank pins and journal to the crank-webs.

Several empirically developed methods of approximation have meanwhile become known in order to determine the stress concentrations arising at the fillet radii referred to. In this connection, special reference is made to MTZ 23 (1962) 12, MTZ 29 (1 968) 3, the Research Bulletin 49/1965 of the Forschungsvereinigung Verbrennungskraftmaschinen (Research Association for Internal Combustion Engines) and the Research Report No. 1 30 (1972) of the Research Association for Internal Combustion Engines.

All the methods described in those publications are based on the determination of the form factors for bending and torsion as an expression for the relationship of peak stress to nominal stress amaX/anOm and, respectively Tmax/Tnom. Dependent on the specific method, the nominal stress is referred either to the crank pin or the crank-web cross-sectional area. According to the method described in Research Report-No. 13Sthe stress distribution in crank-throws of high-speed engines, characterised by thin crank-webs, small piston strokes and, consequently, large journal/pin overlap is studied on a large number of models.Inter alia, the influence of drilled holes in the crank pin and main journal as well as the influence of the chamfer on the crank-webs on stress concentrations was determined. For the first time, this method also ascertains the stresses in the journal fillets and also takes into consideration the stress modes, such as pure bending, bending with shear force and torsion. The results of the research work reflect the latest status of the art in respect of the determination of stresses in crankshaft throws and are proposed to form the basis of the present invention.

Relevant parameters include the crank pin diameter, journal/pin overlap, crank-web width and thickness as well the fillet radii from the crank pin and journal to the crank-web.

The stress-concentration factors are determined by multiplicative linking of functions of these parameters referred to the crank pin diameter, with the nominal stress being referred to the crank-web area.

In dimensioning a crankshaft, the cylinder spacing is given by the cylinder bore and the piston stroke which determine the attainable engine output. Furthermore, the big end bearing width bp required and the main bearing width bG are the result of the anticipated bearing stresses. Finally, it is possible to determine the rubbing shoulders ap and aG at the crank pin and the journal as a function of the required finish of the crank-web. Thus, the designer can choose the relevant parameters referred to. Independent of the cylinder spacing, it is then also possible to select the crank pin diameter dp, the pin/journal overlap s and the crank-web width b, an increase of these values being invariably associated with a reduction of the absolute stress levels.It should also be mentioned that the size of the crank pin diameter dp is limited by the requirement that the connecting rod should be capable of being installed through the cylinder bore.

In contrast to the three values dp, s and b, the remaining parameters, namely the fillet radii rp at the crank pin, the fillet radii r6 at the journal and the crank-web thickness w are a function of the predetermined cylinder spacing or, respectively the length I between centres of main journals as well as the selected bearing widths bp and bG. Consequently, there remains a length k = I - b6 - bp - 2. (a + ap). At the same time, the length is calculated by k = 2. (w + rp + rug).

On the strength of the known functions for the determination of stress concentrations and stress levels, it can be seen that only an increase in the fillet radii rp and rG and an increase in the crank-web thickness w will result in a reduction of stress concentrations. Since the three parameters are, however, firmly interelated by the length k, any increase in one of these parameters will at the same time mean a reduction of the others. This is the reason why the determination of the three parameters from the point of view of minimizing the stresses does not follow from the known status of the art.

An object of the present invention is, in designing a crankshaft of the type initially described, to accomplish a minimization of stresses in the crank-throws independent of the remaining parameters and only by an appropriate selection of the crank-web thickness and the fillet radii at the crank pin and journal.

This object is achieved by establishing a definite relationship of the fillet radii of the crank pin and the journal to the web thickness, which can be represented by a function of the magnitudes rp, rG and w which is proportional to the stresses in the fillet radius of the crank pin, with the ratio of pin/journal overlap to crank pin being greater than or equal to 0.1 5, and less than or equal to 0.26.

The function is derived from the three magnitudes, i.e. crank-web thickness, fillet radius of the crank pin and fillet radius of the journal from the expression

where bG is the bearing width of the journals.

Since the torsional loads in the piston/connecting rod/crankshaft system cannot be accurately predicted in the design stage and can be influenced by torsional vibration dampers, the invention is limited to the bending load, although the load case "bending load and shear force occurring in engine operation is included. This approach appears meaningful since crankshaft failures in most cases are attributed to bending failures as experience has shown.

An optimum stress distribution is important, if only because in present-day high-performance automotive engines the emphasis is on light-weight construction and, consequently, minimum dimensions. In this connection, it is essential that the three parameters rp, rG and w referred to should be made independent of the other parameters because their influence on the stresses arising is definite.

As can be seen from the known functions, there is no interrelationship between the crank-web width b and the three parameters to be optimized. In contrast to this, there is however an interdependence between the crank-web thickness w and the pin/journal overlap s because both occur as related parameters wx = w/dp and sX= s/dp in a double function f (sx, wx). In order to minimize the influence of the crank-web thickness w on the double function f (sx, wx), s" or s/dp is limited as mentioned. Generally, the stress Up is determined in the fillet radius rp of the crank pin because this is the point where the highest tensile stresses arise.

In order to minimize the stress Up as described, the following relationship is additionally required: rP/rG = C.

As repeatedly mentioned, rp in this expression is the fillet radius at the crank pin and rug the fillet radius at the journal. c represents a parameter which can be freely selected within certain limits and lies approximately between 1.0 and 1.9. This parameter permits weighting between the maximum tensile stress at the fillet radius of the crank pin and the maximum compression stress at the fillet radius of the journal.

Furthermore, the following relationship is derived from the initially mentioned length k regarding the crank-web thickness w: k w =rp(1 + 1 /c) 2 As a result, the optimum fillet radius rp at the crank pin can be determined for any selected parameter c as a function of the length k. On the strength of the relationship described, the crank-web thickness w and the fillet radius rG at the journal are determined.

Entering w and rG and differentiating with respect to rp, one obtains always an optimum stress distribution, and the three parameters can be derived from the equation for rp.

Let, for the sake of simplicity, the three factors a1, a2 and a3 be assumed: a, . r2pOp, + a2 . rpOpt + a3 = 0 a1 = 0.403 (--1) c2 k a2= 1.4145.-- + 0.66455.k + 0.903.b,(l ±) c c a3 -0.26155.b6.k-0.26155.k2 If the parameter c = 1, then - a3 rPopt = a2 If c is greater than 1, we obtain an optimum fillet radius at the crank pin of

The invention will now be explained in greater detail with reference to the accompanying drawings, in which:: Figure 1 is a part of a crankshaft seen from the side, Figure 2 is a section taken along the line ll-ll of Fig. 1, Figures 3 and 4 are graphs showing the variation of the stress conditions, Figures 5 and 6 are graphs showing the optimum fillet radii of crank pins for a given parameter c.

Fig. 1 is intended to define the designations used for the dimensions of the crankshaft. The part of the crankshaft shown consists of two half journals 1, a crank pin 2 and two crank-webs 3. The journals 1 having a diameter d6 and fillet radii r6 which end in rubbing shoulders 4 provided on the crank-webs 3 with a thickness aG. The bearing width of the journal 1 is bG. The crank pin 2 in turn has a diameter dp and its ends merge through fillet radii rp into rubbing shoulders 5 provided also on the crank-web 3 and having a width rp. The bearing width of the crank pin 2 is bp. The distance from the centre of one journal 1 to the centre of the adjacent journal is I.The length k which is important for the present invention and not shown in Fig. 1 is described by k = I - b6 - bp - 2(aG + ap) = 2(w + rp + rG), where w is the crank-web thickness. Finally, k is given by the expression: k = 2. (w + rp + rG).

In Fig. 2, the crank-web width is designated b.

The graph in Fig. 3 shows the stress ratio between journal and crank pin fillets. The freely selectable parameter c is plotted on the abscissa 6, and is composed of the ratio "fillet radius of crank pin to fillet radius of journal". The ratio "compressive stress at fillet radius of journal to tensile stress at fillet radius of journal" is shown on the ordinate 7, i.e. - U6Up. It can be seen from line 8 that the ratio --#G:## rises with increasing value selected.

Fig. 4 illustrates the reduction of the stress at the fillet radius rp of the crank pin 2 by plotting on the abscissa 9 again the value c, and on the ordinate 10 the reduction of the stress Up at the fillet radius rp in per cent. A curve 11 is obtained from which it can be clearly seen that the tensile stress Up at the fillet radius rp of the crank pin 2 decreases as the parameter c rises at the expense of the compressive stress U6 at the fillet radius r6 of the journal.

From the length k previously referred to, the following relationship is now obtained: w = k/2rp . (1 +1/c) This permits the optimum fillet radius rp of the crank pin 2 to be determined for any value c selected as a function of the length k. On the strength of this relationship, the web thickness w and the fillet radius r6 of the journal 1 are then determined.

Figs. 5 and 6 show, for instance, optimum fillet radii rp of the crank pin 2, Fig. 5 being for a selected parameter c = 1.0 and Fig. 6 for a parameter c = 1.1. The length k in mm is plotted on the abscissa 12 in each case, and the fillet radius rp in mm on the ordinate 13.The upper line 14, 14' respectively represents the optimum fillet radius rp, whereas the lower lines 15, 15' indicate the minimum fillet radius rp. The band width between the two lines 14, 1 5 and 14', 15' respectively makes it possible for the designer to achieve a reduction in the main bearing load on the strength of a smalier crank pin width consisting of bp + 2 rp with a maximum stress increase of only 2% compared with the optimum, i.e. there is a reduction in the rotating masses.

It follows that by selecting the parameters w, rp and r6 an optimum stress distribution is ensured. In order to obtain the necessary safety against fatigue failure, the absolute stress level can readily be influenced with the remaining parameters, such as pin and journal diameters dp, d6, pin/journal overlap s and crank-web width b. This provides a method of consistently designing the connecting rod/crankshaft system at the time of engineering design for lightweight construction. In conclusion, it may be mentioned that the optimum fillet radii as shown in Figs. 5 and 6 can of course be determined in the same manner for a full range of "c" values that can be selected, i.e. approximately up to c = 1.9.

Claims

1. A crankshaft having a symmetrical crank-throw comprising journals, a crank pin and crank-webs, in which the fillets from the journals and the crank pins to the crank-webs are formed by fillet radii and in which the distance between journal centres are predetermined and the main bearing widths as well as the big end bearing widths are predetermined for reasons of bearing stress, the shape of the crankshaft being determined by an optimized stress distribution, a fixed relationship being given for the fillet radii rp of the crank pin and r6 of the journal to the web thickness w representable by a function f of the magnitude rp, r6 and w which is proportional to the stress at the fillet radius of the crank pin, the ratio of pin/journal overlap s to the crank pin diameter dp being greater than or equal to 0.15 and less than or equal to 0.26.

2. A crankshaft as claimed in claim 1, wherein the function f of the three parameters, i.e.

crank-web thickness w, fillet radius rip of crank pin and fillet radius r6 of of journal is given by

where b6 is the bearing width of the journal.

3. A crankshaft as claimed in claims 1 and 2, wherein by means of a parameter c which is freely selectable within limits and represents the ratio of the fillet radius rp of the crank pin to the fillet radius r6 of the journal 1, a weighting is made possible between the stresses at the fillet radius rp and the fillet radius r6 and, consequently, the determination of a small crank pin diameter dp.

4. A crankshaft as claimed in claim 3, wherein the ratio c of the fillet radius rp of the crank pin to the fillet radius r6 of the journal lies between 1.0 and 1.9.

5. A crankshaft as claimed in claim 4, wherein that for any ratio c the three parameters required for optimizing the stresses, i.e. fillet radius rp of crank pin, fillet radius r6 of the journal and crank-web thickness w, can be exactly determined by differentiating with respect to rp.

6. A crankshaft as claimed in claims 1 to 5, where a reduction in crank pin width is aimed at to reduce the rotating masses and, thereby relieve the main bearings, the band width of the fillet radius rp of the crank pin being fixed so that a stress increase at the fillet radius rp of the crank pin of only 2% is permitted.

7. A crankshaft substantially as herein described with reference to the accompanying drawings.