High-Cycle Fatigue of a Reciprocating Piston Engine

In a reciprocating piston engine, the connecting rods transfer rotating motion into reciprocating motion. The connecting rods are constantly subjected to high stresses and the load increases with the engine speed. A failure of one part in the engine usually results in a replacement of the whole engine. It is therefore of crucial importance to design all engine parts so that none of them fail during the operational lifetime. The connecting rods are identified as critical parts and are here analyzed from the fatigue perspective. The fatigue lifetime is predicted using the Basquin high-cycle fatigue criterion.

This example is based on a model in the Multibody Dynamics Module Application Library, Three-Cylinder Reciprocating Engine, where one connecting rod is modeled as a flexible body while the remaining parts are modeled as rigid bodies. The connections between the different parts are obtained by using a different type of joints. This technique significantly reduces the model size while maintaining the force equilibrium in the assembly.

Model Definition

The three-cylinder engine is presented in Figure 1, and it operates at 1000 RPM. Its material data is taken from structural steel. Additional information regarding its set up can be found in the documentation for Three-Cylinder Reciprocating Engine.

The material data from fatigue tests is summarized in Figure 2. A Basquin relation with the material constants σf' = 1043 MPa and b = −0.116 gives a good fit to the experimental results. The material exhibits a fatigue limit at 210 MPa.

Figure 1: Geometry of the reciprocating piston engine.

Figure 2: Fatigue material curve.

Results and Discussion

At first, stresses in a fillet of the piston end of the connecting rod are examined, see Figure 3. A fillet is chosen since a stress concentration due to geometrical change is expected there. A few revolutions must be computed before a steady state behavior is obtained. From cycle three, the stress history for each consecutive cycles seems to repeat itself. Both the peak stress and the of the stress cycle in general is about the same. The stress history is dominated by the third principal stress since the connecting rod is in compression. The two other principal stresses are small so that the stress state at the fillet can be considered uniaxial. Therefore the third principal stress is taken as the amplitude stress in the Basquin relation as opposed to the von Mises stress which may be more appropriate in a multiaxial loading. The fatigue life prediction is shown in Figure 4.

Figure 3: Stress history in the connecting rod.

Figure 4: Fatigue life prediction in the connecting rod.

The critical point is at the fillet close to the top end of the connecting rod where the Basquin model predicts a fatigue life which is longer than ten billion (4·109) cycles. This is an extremely long life. It can therefore be expected that the stress in the assembly is below the endurance limit that for the used material is 210 MPa. By using the Basquin relation

where σa is the stress amplitude and N is the number of cycles to failure, the fatigue endurance limit life can be back-calculated to 62 million cycles. This is less than the calculated value, see Figure 4, and therefore the connecting rod is designed for infinite life. This could have been already observed in Figure 3 where the stress history is shown. Since the principal stress range is about 110 MPa the stress amplitude is about 55 MPa and that is below the endurance limit of the material.

Notes About the COMSOL Implementation

In the Basquin evaluation the parameter Cycle cutoff can be used to incorporate the effect of endurance limit. The number of cycles that gives the endurance limit must then be back-calculated.

Fatigue_Module/Stress_Life/engine_fatigue

Modeling Instructions

Root

In this example you will start from an existing model from the Multibody Dynamics Module.

Application Libraries

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