Mahle GmbH (Ed.) Cylinder Components: Properties, applications, materials

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4.3 Requirements 73

4.3 Requirements

4.3.1 Mass of the connecting rod

As a general principle, moving masses should be kept as small as possible, in order to help

minimize fuel consumption and to reduce vibrational excitation. One such component, the

connecting rod, affects the mass of the components with which it is linked–a lighter con-

necting rod would allow for less piston, balancer, bearing and crankshaft mass. Further

weight can also be saved by reducing the connecting rod length which allows for a lower

deck height. The changes to the lateral forces on the piston skirt, however, must be taken

into consideration.

In order to maintain quiet operation and low vibration levels, the rotating and oscillating

masses should match as closely as possible. The oscillating mass portion is located on the

piston side and the rotating portion is on the crankshaft side. There exist several potential

ways to attain this goal.

The sintering method allows tolerances in raw part weight within a tolerance range of less

than 1%. MAHLE has also comprehensively developed industrial engineering technology for

steel forged connecting rods and signiﬁcantly reduced the weight variation. The controlled,

fully automated forging process thus allows a tolerance range of less than 1% in the raw

part weight.

Another option to reduce weight variation is weight grading. The oscillating and rotat-

ing masses of the ﬁnished connecting rods are determined and the connecting rods are

divided into different weight classes. For this purpose, the connecting rod is weighed hori-

zontally with two scales, each at the center point of the small end and crank end. The value

at the small end bore corresponds to the oscillating mass, and that of the crank end to the

rotating mass (Figure 4.5).

Figure 4.5:

Distribution of moving masses

of a connecting rod

74 4 Connecting rod

When machining to a target weight, a balance pad is added on the big end boss (if required

on the small end as well), which is machined to adjust to the desired weight.

Only one connecting rod weight grade is installed in a given engine. Because different diam-

eter classes are required for the piston, depending on the ﬁnished cylinder diameter, the

assembly unit consisting of piston, piston rings, piston pin, circlips, and connecting rod can

only be assembled together as one predetermined unit right at the engine manufacturing

line for installation.

4.4 Crank end

The diameter of the crank end bore is determined from the crank journal diameter of the

crankshaft and the bearing shell wall thickness. The critical stress acting on the big end is

the inertia force. The oscillating mass loads the crank end bore in tension, and the bore is

ovalized in the axial direction. This results in bending stresses and lateral forces at the parting

surface. It is important that the joint remains closed under all operating points.

4.4.1 Cracking (fracture splitting)

Cracking, or fracture splitting, of connecting rods has become common practice in recent

years. Nearly all new designs in series production today employ this method to create the

parting in the crank end. The big end is notched inside the bore with a laser or reamer. For

sintered parts, the notch could be pressed in during the manufacture of the blank. Using a

cracking mandrel, the halves are then broken apart (cracked) hydraulically at room tempera-

ture; see Figure 4.6. The resulting parting line (fracture surface) is not machined, and dowel

Figure 4.6:

Fracture surfaces of the crank end, manufactu-

red by cracking

4.4 Crank end 75

sleeves or dowel screws are not needed. The ﬁt is provided solely by the engagement of the

uneven surfaces. The fracture surfaces experience only minimal settling. For optimal fracture

splitting, preventing plastic deformation of the connecting rod material during the process is

critical. Special steel grades with a yield-tensile strength ratio of greater than 75% are used.

4.4.2 Angle split of the crank end

If the crankshaft has a large crank journal diameter, the crank end must be split at an angle

in order to allow the connecting rod to be installed and removed through the cylinder liner.

This leads to complex loading conditions in the parting joint.

In a connecting rod that is split at an angle (Figure 4.7), the upper blind hole thread is

particularly at risk, because it is located directly in the region of highest stress of the entire

connecting rod. This is the area of alternating tensile and compressive loads, which are

increased further by the notch effect of the thread, resulting in an increased risk of fracture.

The cross section around this thread must therefore be dimensioned carefully.

Figure 4.7:

Angled or straight-split

connecting rod and required

diameter of the cylinder liner

for identical crank journal

diameter

76 4 Connecting rod

4.5 Connecting rod shank

Looking at the shank cross section in the pivoting direction (perpendicular to the crankshaft

axis), a differentiation is made between the I- and H-proﬁle. The H-proﬁle is often used in

motorsport engines due to the wipping load at high speeds. The I-proﬁle is preferred for

series production engines because of simpler blank production and thus lower costs at

higher quantities.

The connecting rod shank is subjected to an alternating tension/compression load: The

maximum loads are tension due to inertia force at TDC during the exhaust stroke; compres-

sion due to gas pressure at TDC during the combustion stroke. In addition to fatigue resis-

tance, the connecting rod shank must also feature sufﬁcient buckling resistance.

To supply oil to the small end bore, oil can be fed under pressure from the crank journal pin

through a bore along the length of the connecting rod shank.

4.6 Small end

4.6.1 Pin bearing in the small end

The small end bore accepts the piston pin and, together with the pin boss in the piston,

forms the joint about which the connecting rod pivots.

In a ﬁxed pin connecting rod, the piston pin is typically shrink-ﬁt in the small end bore

and has clearance only in the pin boss. To assemble the piston pin, the small end bore is

heated to approximately 400 °C. This assembled unit can no longer be disassembled non-

destructively.

For a ﬂoating pin design, press-ﬁt bushings are generally used in the small end bore. The

piston pin has clearance both in the connecting rod and in the pin boss. It must be held in

the piston axially by piston pin circlips (Chapter 2.7) and can “ﬂoat” circumferentially on the

oil ﬁlm. The pin boss can withstand higher loads due to the superior lubrication, or a shorter

piston pin can be used with the same load. Highly loaded engines, therefore, use ﬂoating

pins.

The advantages and disadvantages of ﬁxed pin connecting rods and ﬂoating piston pins are

summarized in Table 4.2.

4.6 Small end 77

Table 4.2: Ad

vantages and disadvantages of fixed pin connecting rods and floating design for piston

pins

Fixed pin connecting rod Floating design

Advantages Advantages

No piston pin circlips needed

No sliding bearing needed in the connecting rod,

such as a bushing

Assembled unit can be disassembled

Lower weight due to greater load capacity

Disadvantages Disadvantages

Piston pin cannot be removed easily

Difficult to assemble piston pin

Higher weight with longer piston pin, due to lower

load carrying capability

Circlip grooves and circlips must be provided

Circlips must be assembled

Connecting rod bushings generally required

4.6.2 Geometry of the connecting rod small end

Surface pressure, as a dimension for the bearing load in the small end bore, is derived from

the gas pressure, pin diameter, and bearing width. It is generally greater than 100 MPa.

Gas pressure, as the highest magnitude load, acts only in the direction of the crank end bore,

which has led to the development of various types of support in the small end bore to meet

requirements relating to capacity, weight, and cost.

The parallel connecting rod is the basic version and the easiest to manufacture. It is the

most economical for manufacturing if the crank end has the same width as the small end

(Figure 4.8). Due to weight reduction requirements, this type is often replaced with one of the

variants described below, whereby the lower part of the small end bore, which is subjected to

the gas pressure, is wider than the upper part, which is subjected only to inertia forces.

Figure 4.8: Cross-sectional

shapes of the small end bore

Left: parallel connecting rod

Center: trapezoidal connecting rod

Right: stepped connecting rod

78 4 Connecting rod

The trapezoidal connecting rod is tapered at the small end and thus gets wider in the

direction of the connecting rod shank. The pin boss is adapted accordingly, in order to

reduce surface pressure here as well. During design, particular care must be taken as the

distance between the connecting rod and the pin boss in the direction of the piston pin axis

is reduced when the connecting rod pivots (Figure 4.8).

The stepped connecting rod presents the greatest challenge to the manufacturing pro-

cess. However, it best combines load carrying capability, mass reduction and hydrody-

namic conditions. During design, here again, it must be ensured that the connecting rod

does not collide with the piston when it pivots (Figure 4.8).

4.6.3 Bushing-less pin bearing in the small end

This solution reduces oscillating masses, which is of increasing importance in terms of

smooth running behavior and fuel consumption. Adequate lubrication of the small end bore

is essential for this concept. To improve tribological behavior, the small end bore contains a

proﬁled pin bore and may require a phosphate coating (Figure 4.9).

Figure 4.9: Shape optimization of the small end bore, without bushing

4.7 FE analysis of the connecting rod 79

necting rods for inline engines. The analysis

of connecting rods for V-engines depends

on the number of asymmetries present and

needs to be determined individually. The

assembled structure is ﬁxed for the analysis

solely by means of contact boundary condi-

tions. Direct restraint of the connecting rod

structure is avoided because it would lead

to overconstraint of the conditions at the

restraint points. The assignment of material

properties concludes the modeling process.

Figure 4.10:

Three-dimensional FE model of the connecting

rod of a passenger car gasoline engine with

bolts, bearings, and piston pin, as well as a sub-

stitute piston model for load application

4.7 FE analysis of the connecting rod

4.7.1 Modeling

The starting point of every FE analysis is modeling, i.e., the partitioning of the affected struc-

ture into many volume elements. The FE model includes, in addition to the pin end bore with

the cap, the bearings, bolts, and the piston pin, as well as a suitable replacement model for

the piston and the crankshaft (Figure 4.10). Modeling of all individual components is realized

as a three-dimensional structure including all signiﬁcant details, with only minor simpliﬁca-

tions (e.g., bolt threads). Symmetrical models can be used to limit the modeling effort for con-

80 4 Connecting rod

4.7.2 Stresses from assembly

The ﬁrst load case is bolt pretensioning, which results from the assembly of the cap to the

connecting rod shank. A prerequisite for realistic determination of the resulting stresses is the

consideration of the geometry of the bolt shank, the parting line type (cracked or machined

parting line), the centering of the joint face, the bolt underhead contact face, thread depth

(number of load-bearing threads), and bearing crush.

4.7.2.1 Bolt force

Analogous to the speciﬁcation for tightening the connecting rod bolts, the load on the bolt

joint is prescribed for the assembly simulation. In an iterative process, the extension and

thus the stress in the b

olt shaft is varied until the prescribed pretensioning force of the con-

nection has been reached. The yield point of the bolt material generally limits the amount of

pretensioning force. In some cases, however, the surface pressure on the cap in the area of

the bolt underhead contact face can be the limiting factor.

As a reaction to the bolt force, the crank end bore deforms into an out-of-round base bore

(Figure 4.11).

In the connecting rod manufacturing process, however, a perfectly round crank bore is gen-

erated during the ﬁnish machining by a properly applied bolt clamp load. It ensures ideal

geometry for this highly stressed bearing point. This manufacturing process is also recreated

in the simulation by a suitable procedure. This is necessary in order to prevent any disal-

lowed stress increases in the contact zones in the subsequent steps of representing the

operating loads.

Figure 4.11:

Representation of undesired deformation and

bulging of the crank end bore after connecting

rod bolts have been tightened

4.7 FE analysis of the connecting rod 81

4.7.2.2 Bushings, bearings, and shrink fit

For the typical plain bearing of the crank end bore on the crankshaft, split bearings with

deﬁned crush are used for ﬁxing and securing the position of the bearings during operation.

The bushing in the small end bore, or shrink-ﬁt piston pin in the small end bore, generates

stresses due to interference. This interference in the bearings, piston pins, or bushings is rep-

resented in the simulation by appropriate contact boundary conditions. The resulting static

stresses are later combined with the dynamic stresses from operational loading.

4.7.3 Stresses from engine operation

It follows from the kinematics of the crankshaft drive that the piston, together with the small

end bore and the piston pin, performs an oscillating motion, and the crank end bore with

the crank journal on the crank throw primarily performs a rotating motion. The displacement

of the crank end bore leads to a pivoting motion of the connecting rod. The measure of the

pivot angle of the connecting rod is determined by the geometric dimensions of the crank

drive (crank radius and length of connecting rod).

The pivot motion of the connecting rod leads to alternating transverse acceleration of both

the big and small ends, with an approximately sinusoidal curve (Figure 4.12). The vertical

motion of the connecting rod leads to a longitudinal acceleration, which also features a

modiﬁed sinusoidal curve. The stroke to the connecting rod center distance ratio (crank

radius to length of connecting rod) determines the degree of deviation from the sinusoidal

curve and leads to the acceleration at top dead center (TDC) being greater than that at the

bottom dead center. The two accelerations would be equal only in the case of an inﬁnitely

long connecting rod.

In order to translate the dynamic operating loads on the connecting rod into suitable bound-

ary conditions for a static structural analysis, different load cases that can occur during one

or two crankshaft revolutions (depending on the working principle, 2- or 4-stroke) are cap-

tured and applied to the structure in the form of quasi-static boundary conditions.

N For the simulation of the load at TDC in the combustion cycle, the maximum of the com-

bustion chamber pressure is generally applied, so that a slight displacement of the gas

pressure maximum can occur, compared to the representation in Figure 4.12. The inertia

force directed opposite the gas pressure is taken into consideration, which counteracts

the combustion force to a degree.

N Only the inertia force, without any combustion chamber pressure, is therefore used accord-

ingly at TDC in the gas exchange cycle.

N In order to calculate the load due to transverse acceleration, the respective maximum from

the transverse acceleration curve for both the small end bore and the crank end bore

are applied in combination with the effective combustion chamber pressure at the cor-

responding point in time.

82 4 Connecting rod

The individual loads mentioned are combined appropriately to obtain a complete represen-

tation of the operating load. Ten relevant load cases are the result. Depending on the engine

type, there are different weightings of the individual load cases:

N For passenger car diesel engines, the gas pressure load on the connecting rod dominates,

whereas the load due to transverse acceleration is very small relative to the gas pressure

load and can be neglected.

N For passenger car gasoline engines, likewise, the gas pressure load on the connecting

rod dominates, while the load due to transverse acceleration is small for the typical speed

range (up to about 8000 rpm) and therefore negligible in general.

N For high-speed sport and racing engines, the inertia forces are essential and the loads due

to longitudinal and transverse acceleration are correspondingly high. Particularly at very

high speeds, the inertia forces can exceed the load due to gas pressure, and the greatest

load magnitude can result, for example, from the transverse acceleration.

4.7.3.1 Gas pressure

An example of the resulting comparative stresses on the connecting rod structure of a pas-

senger car gasoline engine, under combined assembly, gas pressure, and iner

tia forces at

rated speed at TDC of the combustion cycle, is shown in Figure 4.13.

Figure 4.12:

Plot of gas pressure and

inertia forces for a passenger

car gasoline engine in a four-

stroke combustion cycle