Expected Cycle Life of Piston Rod

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Piston/Rod Assembly

Contents

Expected Cycle Life of Piston Rod

Description (target value, direction of improvement) and related User Requirements

The piston rod’s cycle life will be calculated with a target value of around 10 years of use. The mower would be used on average for an hour at a time once a week. Since the mower may be in a warmer climate, it could potentially be used 52 weeks a year. The direction of improvement is up and relates to the user requirements of durability and long lasting.

Strategy for Analyzing Engineering Specification

Testing the piston rod’s cycle life will be very similar to the crankshaft. To simplify the analysis, the piston rod will be modeled using only a quarter of the geometry.. This is possible due to the symmetry of the rod. The cuts can treated as constrained boundary conditions in ANSYS to simulate the missing material. The forces applied against the crankshaft will be applied to the bottom of the rod as a vertical force. The top of the rod will be constrained to simulate the piston head. A solution for Von Mises stresses will be ultimately found. A comparison between the maximum stresses and a fatigue life curve can be made to determine the cycle life of the piston rod.

Design Decisions/Parameters Affected

The manufacturer designed the rod to be thick enough and strong enough to resist all torques and forces applied to it throughout the combustion cycle.

Key Geometric, Inertia, and Material Properties

The material of the rod was found to be 3 series aluminum as given by the manufacturer. The important dimensional properties are the length vs. width. Important material properties are the density at 0.0986 lb/in3, modulus of elasticity at 10000 ksi, poisons ratio of 0.33, and a tensile yield strength of 18000 psi.

Type of Analysis for Obtaining Results

The piston rod’s geometry was simplified. This was necessary due to the complexity of the structure. Lines of symmetry were created down the center of the rod leaving a quarter of the original geometry. This was possible by observing the geometry of the piston at the point of the ignition which put all the stress on the rod itself, not the bolts as originally thought. After importing the model into ANSYS an analysis was performed using the intermittent force calculation from ADAMS view. This force was found to be around 70 lbs. Two simple hand calculations can be done to determine the validity of the ANSYS simulation. Determining the maximum bending stress and buckling force will allow for a comparison.

Boundary Conditions and Loading

Ansys5 mjd.JPG

Quantitative Results

The following figure displays the simulated deformed shape due to the loading. Realistically the rod would bend before being crushed due to the phenomenon of buckling under a one time load.

Ansys6 mjd.JPG

Below are the Von-mises stress contour plots. The nodal solution was taken due to the slight discontinuity of the elemental solution. The load applied was 70 pounds from the piston head. In order to apply this to the correct area and in the correct direction, only the nodes on the bottom surface where the piston head meets the rod were selected. The total force was then divided by the number of nodes and then applied in the vertical direction.

Thumbnails: Click to Enlarge
Ansys7 mjd.JPG
Ansys8 mjd.JPG

The maximum stress was found to be 5884 psi which yields a factor of safety of 3.06 under the 18000 yield strength of 3 series aluminum. This stress is only found around sharp corners with most of the piston being well under this maximum stress. When combustion is not occurring, the only resistance on the rod comes from pumping air and friction from the cylinder walls. Since these forces decelerate the rod the stresses are near zero when combustion is not occurring.

A back of the envelope calculation yielded an area of 0.0669 in2 and a stress of 1047 psi. This value is in-between the colors shown in the main part of the shaft giving the ANSYS calculation some validity. In order to calculate the critical buckling force, the moment of inertia was calculated and found to be 0.01065 in4. The critical length due to the geometric constraints is the length of the rod and is 2.225 in. This gave a critical force of 218 kips. The rod’s factor of safety was 3119 showing how unlikely a failure due to buckling would be.

Suggested Changes to Improve the Quality of this Design

The design’s factor of safety was around 3 which is a fairly high number. As stated for the crankshaft, the only real improvement to the design would be to smooth out the sharp edges on the rod. This would let the rod’s stress distribution be more normalized and much lower.


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