Fatigue Analysis Correlation and Failure Analysis of Optimized Connecting Rod

8/9/2019 Fatigue Analysis Correlation and Failure Analysis of Optimized Connecting Rod

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IPASJ International Journal of Mechanical Engineering (IIJME)Web Site: http://www.ipasj.org/IIJME/IIJME.htm

A Publisher for Research Motivation........ Email: [email protected]

Volume 3, Issue 5, May 2015 ISSN 2321-6441

Volume 3, Issue 5, May 2015 Page 14

ABSTRACT

The paper deals with the static and fatigue analysis of optimized design of automotive connecting rod. Optimized design is

nothing but modification of existing design having initial weight 11.46kg. After modifying I-Section, there is about 17.37%

weight reduction achieved. Fatigue analysis of modified design having weight of 9.47kg, shows fatigue life of 1.5E006. Experimental validation is done on optimized Connecting Rod design by testing the same on servo hydraulic testing machine

which shows same fatigue life as given by FEA analysis. Then failure analysis has been carried out to check whether any other

factor is responsible for this failure except design parameters or not. These checked parameters are chemical composition,

surface finish at failure region and grain flow. These parameters are as per specifications. So we can say that failure is due to

design only. But since fatigue life is sufficient, the modified design is optimized on.

Keywords: Connecting rod, Mass Optimization, Static Analysis, Fatigue Analysis, Fatigue Factor of safety,

Experimental Validation, Failure Analysis.

1.

INTRODUCTION

1.1

Existing Design Analysis Summary

We have carried out static and fatigue analysis of existing design which is having geometrical detail is as follows. Itsmass is 11.46kg.

Figure 1- Existing Design Geometry and I-Section details

Applying the loads and boundary conditions discussed below, we found that static and fatigue analysis result summary

as follows.

Table 1 – Existing design analysis summaryLoading Condition Tension at Piston End Compression at Piston

End Fatigue F.O.S. Safety Fatigue Life

1st Principal Stress 3rd Principal Stress Loading at Piston End 101 MPa

(at cross oil hole) 350 MPa

(at cross oil hole) 2.3 1.393E7

Note: Since loading at piston end condition is worse than loading at crank end, we will consider loading at piston end

condition for simulating most worst condition, so that design will be on safer side.

1.2

Topology OptimizationHere topology optimization is carried out in order to optimize the original connecting rod. HYPERMESH OptiStruct is

used for optimization analysis. Here blue color shows scope for material removal.

Fatigue Analysis Correlation and Failure

Analysis of Optimized Connecting RodMr. Vikram A. Shedge

1, Prof. K. H. Munde

2

1 P. G. Student- ABMSP”s Anantarao Pawar College of Engineering & Research, Parvati. (Pune University)

2 Assistant Professor- ABMSP”s Anantarao Pawar College of Engineering & Research, Parvati. (Pune University)


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Figure 2 - Scope for material removal

Table 2- Topology Optimization parameters

Objective Reduce volume of the component only in the design area.

Stress constraint Yield stress is given as stress constraint parameter, 600MPa

Optimization Response Mass or volume is the response.

Boundary conditions. All the four loads cases applied simultaneously.Accordingly, we modify existing design and analyzed the under same loading and boundary conditions as discussed

below.

2. STATIC ANALYSIS OF MODIFIED DESIGN

2.1 Calculation of Forces Acting on Connecting Rod

2.1.1 Input Parameters

The input parameters of it are described below. The original mass of connecting rod is 11.47 kg. The calculations for

original design are as follows.

Maximum gas pressure, Pmax = 2175.57 psi.

Length of connecting rod, L = 345 mm

Reciprocating masses, Mr = 6000 gm

Bore diameter, D = 150 mm Crank radius, R = 100 mm

Crank speed, N = 2500 RPM

Firing angle, θ = 90 - 11

0

2.1.2 Material Properties

Material density δC = 7.85E-9 ton/mm3

Poisson's ratio µ = 0.3

Young's Modulus E = 2.1E5 MPa

Yield strength Sy = 600 MPa

Ultimate tensile strength, Su = 900 MPa

2.1.3 Force due effect of gas pressure on piston

Fg = (Maximum Gas Pressure) × (C/s Area of Piston)

2.1.4 Force due to inertia of reciprocating masses

Inertial force due to reciprocating parts is given by,

Fi = (mass of Reciprocating Mass i.e. piston assembly.) × (Acceleration)


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2.2 Load Cases and Boundary Conditions

The static analysis is carried out for tensile and compressive conditions under maximum gas load. There are four load

cases in the analysis of con rod.

3)

Loading at Piston End

Tensile loads i.e. load due to inertia 51699.555N is applied over 180º at piston end with crank end restrained over180º of contact area.

The compressive load i.e. load due to gas pressure 265071.881N is applied as a uniformly distributed load over

120º of contact surface at piston end with crank restrained over 120º of contact area.

4) Loading at Crank End

Tensile load i.e. load due to inertia 51699. 555N is applied over 180º at crank end with piston end restrained over

180º of contact area.

The compressive load i.e. load due to gas pressure 265071.881N is applied as a uniformly distributed load over

120º of contact surface at piston end with piston end restrained over 120º of contact area.

2.3 Modified Design Geometry Details

Modified Design geometrical details and I-section are shown by magenta color and that of existing design by cyan color

in fig3. Modified design has mass 9.47kg.

Figure 3 - Modified Design I-Section details (Superimposed)

2.4 FEA Results and Fatigue factor of safety Calculations for Modified Design

By applying same loading and boundary conditions, we have carried out static stress analysis for modified design in the

same manner as for existing design. Highest 1st Principal stress is 142Mpa at cross oil hole at shank as shown in fig.4

(a). Also highest 3rd

Principal stress is 439MPa at the same location as shown in fig 4(b).

Figure 4 (a) - 1st Principal Stress Fig.4 (b) - 3rd Principal Stress

Fatigue Factor of Safety for this modified design by using Goodman theory will be,

Now on the basis of above FEA analysis, we have carried out fatigue analysis of existing design in N-Code dedicative

fatigue analysis software.

3.FATIGUE ANALYSIS OF MODIFIED DESIGN

There are five important segmental windows in the N-Code work environment.

142 MPa 439 MPa


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1.

FE Input Window: - This window is starting point of fatigue analysis through which we taking the FE input data.

This data is nothing but the .rst file that we have prepared during above static analysis. This file contains two load

cases i.e. tension loading and compression loading. This FE input data is used for fatigue life calculations.

2. SN-Analysis: - This window is main part of this software, where actual fatigue life calculations are carried out. Here,

we goes through different steps related to defining material and loading conditions and accordingly this will calculatefatigue life for each and every node.

Figure.5 (a) - N-code Work environment

3. FE Display: - Visual display of fatigue analysis can be viewed through this window.

4.

Hot Spot Detection: - This window will show the most critical and mast damaged areas which is having lower life

along with corresponding node numbers.

5.

Data Value Information:- This window is used for data acquisition in which whole data related to the fatigue life is

tabulated with corresponding node numbers.

Figure 5 (b) - Fatigue Life of modified design

Now this fatigue life is in the range of E006 cycles. So we can say that this design is having sufficient life with having

17.35% less mass than the existing design and modified design is one of the optimized design.

4.EXPERIMENTAL VALIDATION OF OPTIMIZED DESIGN

4.1Test Facility

A vertical axis universal servo-hydraulic fatigue test machine of 1000 KN capacity is used for connecting rod fatigue

tests. Custom made fixtures used in the previous project are modified to suit the test machine. Pin ground to big and

small end journal dimensions are fitted to the rods and clamped in split blocks bolted together to aid assembly anddismantling of the test set up as shown in figure.


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Figure 6 (a) – Servo Hydraulic Machine Figure 6 (b) – Experimental Set Up

The rods are used to test under constant amplitude axial loading at a frequency of between 1 to 3.5Hz.

Load conditions used: Rated speed: 5270 kgf (i.e. equivalent to inertia force = 15699.555N.) to -27020 kgf (i.e.

equivalent to Gas force = -265071.8801N.) Total 10 connecting rods are tested and their results are as follow.

4.2Fatigue Life Results by Experimental Fatigue Test

Figure 7 - Failure Location (at oil way cross drill)

Table 3- Number of cycles passed and failure location

Rod

No.

No. of Cycles

Passed

Failure location

T1 1,482,326 Oil way cross drilling in shank

T2 1,43,598 Small end bush & shank, Big end bearing cracked


T4 1,358,339 Oil way cross drilling in shank, small end bush cracked


T6 5,94,612be Big End oil way






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4.3Observations

From experimental results, it is clear that out of 10 tests, 8 connecting rod is failed at oil way cross drilling which is the

same location shown by FEA Analysis and Life also matching with that of calculated by the N-Code software which in

the range of 1E006 to 1.51E006 cycles.

5.FAILURE ANALYSIS OF FAILED CONNECTING ROD

Failure analysis is done to investigate the real cause of fatigue failure. If other fatigue influencing factors will be as per

specifications then failure responsible will be inherently design only. Failure analysis report on each factor and its

remark is as follow.

5.1Chemical Composition

Table 4 – Chemical Composition

Specification C Mn Si P S Cr

Min. % 0.38 0.70 0.10 - 0.025 0.90

Max. % 0.44 1.00 0.40 0.020 0.040 1.20

Actual 0.43 0.84 0.32 0.016 0.030 1.07Specification Ni Mo Al Cu V B

Min. % - 0.15 0.010 - - -

Max. % 0.10 0.25 0.040 0.08 - -

Actual 0.09 0.20 0.036 0.08 0.007 0.0002

Remark: Chemical Compositions meets the specifications and confirms to 708M40T as per BS970

5.2Microstructure

Figure 8– Microstructure at surface and core region

Remark: No Decarb observed at surface and Tempered Martensite with intermediate structure observed.

5.3 Hardness

Specification: 248 – 302 BHNActual Value: 277 BHN

Remark: Hardness value meets the specifications.

5.4 SEM Analysis

Figure 9 – SEM Analysis


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Remark: Crack is initiated from machined surface and no inclusion observed at Fracture initiation point.

5.5 Grain Flow

Figure 10 – Grain Flow

Remark: No abnormality observed in grain flow.

5.6 Surface Roughness

Figure 11– Surface Roughness

Remark: Surface roughness at oil hole surface is Rz = 4.75 microns, which is as per specification.

5.7 Oil Hole Surface Microstructure

Figure 12– Microstructure at oil hole Remark: No compaction or reformed Martensite observed.

6.CONCLUSIONS

Existing design of connecting rod is having mass of 11.46kg and having fatigue factor of safety 2.38. Fatigue life

of this design is 1.393E007 cycles, which somewhat more than the requirement criteria of E006 cycles. In Modified design, we have removed material from existing design so as to get 17.35% weight reduction. This

design is having fatigue factor of safety 1.81 and having fatigue life of 1.581E006 cycles, which is satisfying the

criteria of range of E006 cycles. So, removing the material from different sections, stress level increases and fatigue factor of safety along with

fatigue life reduced. But if removal of material in such fashion that fatigue factor of safety is in between 1.5 to 2.0

then component will get sufficient fatigue life in the range of E006 cycles.

From experimental results, it is clear that out of 10 tests, 8 connecting rod is failed at oil way cross drilling which

is the same location shown by FEA Analysis and Life also matching with that of calculated by the N-Code software

which in the range of 1E006 to 1.51E006 cycles.


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Since all other factors contributing to fatigue life are as per specifications, connecting rod fatigue failure is due to

Design aspect only. So if that component is having sufficient fatigue life (i.e. in the range if 106Cycles) then our

Design is Optimum one.

References

[1]

P. C. Sharma and Aggarwal, “A text book of Machine Design”, fourth edition 1993, S. S. Kataria and sons’

publication.

[2]

R K. Jain, “A text book Machine Design”, Seventh edition 1995, Khanna publications.

[3] Priyanka D. Toliya, Ravi C. Trivedi, Prof. Nikhil J. Chotai, “Design and Finite Element Analysis Of Aluminium-

6351 Connecting Rod”, International Journal of Engineering Research & Technology (IJERT) ISSN: 2278-0181,

Vol. 2 Issue 5, May - 2013.

[4] M.M. Noor, M.N. Shuhaizal, K. Kadirgama, Julie J. Mohamed, M. R. M. Rejab, A. N. M. Rose, “Analysis Of

Connecting Rod Based On Finite Element Approach” , Malaysian Technical Universities Conference on

Engineering and Technology March 8-10, 2008, ISBN 978-983-42358-4-0.

[5] Manish P Vekariya, Marmik M Dave, “Static Structure Analysis of Diesel Truck Engine Connecting Rod” volume

2, issue 6, June 2013 ISSN No. 2277-8179.

[6]

Abhinav Gautam, K Priya Ajit, “Static Stress Analysis of Connecting Rod Using Finite Element Approach” IOSR

Journal of Mechanical and Civil Engineering (IOSR-JMCE) e-ISSN: 2278-1684,p-ISSN: 2320-334X, Volume 10,

Issue 1 (Nov.-Dec. 2013), PP 47-51.

AUTHOR

Vikram Appa Shedge received the Bachelor of Engineering degree with distinction in

Mechanical Engineering from Rajarambapu Institute of Technology, Sakharale in 2012 and

currently working as a student of M. E.- Design in Pune University (ABMSP’s Anantarao

Pawar College of Engineering & Research, Parvati- Pune). During 2012-2013, he worked as an

Assistant Professor at PES’s, College of Engineering, Phaltan.

Fatigue Analysis Correlation and Failure Analysis of Optimized Connecting Rod

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