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# Finite Element Model of Piston-connecting Rod System

## INTRODUCTION

### PROBLEM DEFINITION

### CONSTRAINTS

#### Natural Boundary Conditions:

### Essential Boundary conditions:

### CHALLENGES

### ASSUMPTIONS

### MODEL SETUP

### Idealization

### ELEMENT SELECTION

### MESH DEPENDENCY AT CRITICAL NODES

### ELEMENT vs. NODAL SOLUTION

### FATIGUE AND TRANSIENT ANALYSIS

## MODEL ANALYSIS

### Displacement and deformation of the connecting rod

### Von Mises Stress Analysis of the Connecting Rod

### Application of Surface pressure forces on the piston end of connecting rod at various angles

## ELEMENT vs. EXPECTATIONS

## RESULTS AND DISCUSSION

## CONCLUSION

- Category:
**Life** - Subcategory:
**Professions & Career** - Topic:
**Engineer**,**Model** -
Pages:
**4** - Words:
**1912** - Published:
**27 March 2019** - Downloads:

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“In every internal combustion engine, there are various parts, which are directly or indirectly connected to each other. To convert reciprocating motion of piston into rotary motion of crankshaft, the piston is connected to crankshaft with the help of connecting rod and gudgeon pin”. “Together with the crank, they form a simple mechanism that converts linear motion into rotating motion”.

“The process of fuel combustion in the combustion chamber leads to the loads onto the piston and is transmitted to crankshaft through the connecting rod. One end of the connecting rod known as small end, is connected to the piston through gudgeon pin while the other end known as big end, is connected to crankshaft through crank pin. Connecting rods are usually made up of drop forged I section. In large size internal combustion engine, the connecting rods of rectangular section have been employed such that its moment of inertia is easily and effectively evaluated. In such cases, the larger dimensions are kept in the plane of rotation.”

“In petrol engine, the connecting rod’s big end is generally split to enable its clamping around the crankshaft. Suitable diameter holes are provided to accommodate connecting rod bolts for clamping. The big end of connecting rod is clamped with crankshaft with the help of connecting rod bolt, nut and split pin or cotter pin. Generally, plain carbon steel is used as material to manufacture connecting rod but where low weight is most important factor, aluminum alloys are most suitable. Nickel, alloy steel are also used for heavy-duty engine’s connecting rod. The connecting rod, being rigid by itself, may transmit either a push or pull and so the rod may rotate the crank through both halves of a revolution, i.e. piston pushing and piston pulling.”

The overall model of the connecting rod and its respective forces are to be analyzed effectively using ANSYS such that all stresses which are developed within the connecting rod are also said to be determined simultaneously with perfection. Study of effects these stresses enable us to ensure that the connecting rod will not fail under the extreme loads associated with the high compression ratios of a petrol engine.

The main objective of the project is to determine Von mises stresses, Shear stresses, Total deformation, Mesh dependency, Fatigue analysis and Optimization of the existing connecting rod. The project shows only static finite element analysis of the connecting rod, which has been performed, with the help of ANSYS WORKBENCH 13.

The loading conditions on the connecting rod vary with time; therefore it is important to study pressure variation due to changes in crank angles. The drive for the project is to implement all the concepts used during the finite element classes, as a connecting rod goes through various loading conditions with multi axial stresses and deformation.

Load is applied to the piston pin end of the connecting rod which is assumed to have a sinusoidal distribution of pressure over the surface area. The piston pin end is subjected to both tensile and compressive stresses due to connecting rod motion inside the piston cylinder.

The crank end of the connecting rod had all degrees of freedom fixed for half of the internal circular area when the load is being applied to the piston pin end.

• Non-coinciding of the elemental and nodal solutions for a particular element type proved that element selection wasn’t proper. To overcome this challenge, elemental and nodal solutions for different element types were plotted against active number of degrees of freedom to obtain the optimal number of degrees of freedom.

The small end of the connecting rod experiences maximum stress due to periodic contact with the bush and pin assembly. Periodic application of stresses affects fatigue life of the connecting. Overall localizing the analysis to only the small end cannot do fatigue analysis of the connecting rod. To overcome this challenge and reduce the complexity of the analysis, fatigue analysis was done at major stress points.

The connecting rod is a solid with isentropic material.

The piston pin and bushing are not considered as a part of connecting rod assembly.

Maximum tensile load is less than maximum compressive load; therefore forces are applied to lower part of the piston end.

Thermal and shear deformations are not considered.

Crank end of the connecting rod is completely fixed in all degrees of freedom.

The geometry of the connecting rod is as per the model given below. The ANSYS command file used for modeling is given in the appendix. Considering the application of forces and geometry of the connecting rod, an axis symmetric model was designed. An axis symmetric model reduces the number of nodes used, thus helping in obtaining better mesh refinement and accurate results.

With idealization, we can convert the physical model into mathematical model of the system, thus reducing the complexity of the physical problem.

With discretization of the mathematical model we can reduce the number of degrees of freedom to a finite number. By considering spatial discretization and neglecting time discretization for the static analysis of connecting rod, the degrees of freedom are further reduced. Applying constraints to the crank end of the connecting rod helped in creating a discrete model for analysis. With the help of idealization the degree’s of freedom were defined within a fixed boundary. Stresses due to thermal expansion and stresses on the pin were ignored. Transient and static analyses were considered with the help of idealization.

For accurate results we have chosen SOLID186, which is a twenty node structural solid used for 3D solid structure. The element has three degrees of freedom at each node and translations in nodal x, y and z directions. The element supports plasticity, hyper elasticity, stress stiffening, creep, large deflection, and large strain capabilities.

Accuracy of analysis was found to change with number of degrees of freedom. Elemental and nodal solutions for different number of degrees of freedom were plotted to find a converging point. It was observed that least error percentage was obtained at 33288 degrees of freedom. Number of degrees of freedom was calculated by considering active number of nodes and multiplying it with number of degrees of freedom of a single node (UX, UY, UZ). Table for constrained node is given in appendix 4.

Confirmation of local analysis can be done only by performing global analysis. Maximum elemental and nodal solutions with number of degrees of freedom were plotted to obtain a point of convergence. The obtained results were similar to critical node analysis results i.e. the convergence was observed at 33288 number of degrees of freedom. Therefore, analysis on the model was performed considering 33288 number of degrees of freedom.

Stress amplitude (S) Vs. Cycles to failure (N) i.e. the stress life theory was used for fatigue analysis of the selected connecting rod. It implicates that the component will have infinite life for a number of load cycles over to 10¬7 load cycles.

As the loading on the connecting rod is cyclic in nature the results are to be combined for the desired value of alternative and mean stresses for each operating condition. A transient analysis was conducted by application of alternative stresses. At first a compressive load of 49.98KN was applied at the pin end for duration of 0 to 0.02 seconds and then similar load of 49.98KN in the opposite direction for duration of 0.02 to 0.04 seconds subjected to 500,000 load cycles.

The connecting rod shows a displacement of 0.060199 mm under surface forces of 49.98 kN on the piston end of the connecting rod. From figure 1.8, it can be deemed that maximum deflection takes place at the smaller eye of the connecting rod.

From the nodal solution for Von Mises stress it is observed that maximum stress is at the lower section of the piston end. With application of compressive surface pressure forces of 48.98MPa, stresses of 261.851 Mpa are seen on the connecting rod. Mesh refinement with element size of 1 mm at areas 14 and 17, which are subjected to high stresses helped in obtaining better results. Stresses obtained near the circumference of the piston end showed stress levels of up to 200MPa, whereas stresses along the beam are approximately 100Mpa.

High stresses in the bottom section of the piston end shows scope for geometric optimization which is an area of future analysis for our final project.

In static analysis of connecting rod, forces are not only applied to the bottom section of the piston end but everywhere at the inner circumference of the piston end. For correct analysis of the Von Mises stress on the rod, forces were calculated at different angular cross section by substituting the forces in the above given formula. Forces at different cross section give different stresses and displacement as shown in figure 1.12 resulting in bending, extension and compression of the connecting rod. Maximum stress value was seen when forces were applied at the bottom section of the piston end.

For choosing the correct element for analysis, different solid elements such as solid45, solid92, solid95, solid185, solid186 and solid187 were used. Von Mises stress values from these solids were compared as shown in table 1.2 in the appendix. Solid45 and solid 185 were not considered, as their results didn’t coincide with the values of the other solids.

From all the feasible solid elements, considering the results and number of nodes, solid186 was chosen for our connecting rod analysis.

Considering future analysis, which includes mapped meshing, the element that supports maximum number of nodes, that is SOLID186 was chosen for analysis.

Assuming the connecting rod to be a component of “Mitsubishi 2.5 liter V6 engine (6G73)”, pressure vs. crank angle plot was obtained by running Engine Sim software. From the above plot maximum pressure of 11070 KPa was observed at 00 crank angle.

Global and local analysis of the connecting rod was plotted against number of degrees of freedom to obtain the optimum number of degrees of freedom i.e. 33288 active number of degrees of freedom.

By performing relevant calculations, forces acting on the connecting rod were obtained. ANSYS WORKBENCH 13.0 was used for analysis of the connecting rod. Maximum deformation of 0.060199 mm, resultant displacement and Maximum Von Mises stress of 261.851 MPa were determined at relevant locations.

Similar simulations were executed by applying relevant angular forces in the inner circumference of the piston end of the connecting rod to compare effects of stresses.

Static analysis of connecting rod proved that it can withstand a load of 49.98 kN without failing. With a safety factor of 0.5 and a cyclic loading of 49.98 kN, it was observed from fatigue analysis that the rod can run for infinite number of cycles without failing.

All analysis has been restricted to a global model till now, which shows system stresses and displacements from a global point of view. Better and accurate results were obtained by performing local model simulations. Mesh refinement of an axis symmetric geometry of the connecting rod was done to obtain better results with minimum number of nodes. With all these in place, a connecting rod subject to infinite number of cycles in its lifetime, with compressive and tensile loading in each cycle, was subject to fatigue failure analysis.

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