 Open Access
 Total Downloads : 3158
 Authors : Priyank D. Toliya, Ravi C. Trivedi, Prof. Nikhil J. Chotai
 Paper ID : IJERTV2IS50371
 Volume & Issue : Volume 02, Issue 05 (May 2013)
 Published (First Online): 16052013
 ISSN (Online) : 22780181
 Publisher Name : IJERT
 License: This work is licensed under a Creative Commons Attribution 4.0 International License
Design And Finite Element Analysis of Aluminium6351 Connecting Rod
Priyank D. Toliya, Ravi C. Trivedi, Prof. Nikhil J. Chotai Department of Mechanical Engineering
Marwadi education foundations group of institutions, Rajkot, 360 003, Gujarat, India
Abstract
The connecting rod is the most relevant pats of an automotive engine. The connecting rod is subjected to an extremely complex state of loading. High compressive and tensile loads are due to the combustion and connecting rods mass of inertia respectively. The objective of this research is to investigate the failure analysis of the connecting rod of the automotive engine. Apart from conventional material of connecting rod I choose the connecting rod of FM70 Diesel engine which is made of Aluminium 6351. static analysis is done to determine the von Misses stress, elastic strain, total deformation in the present design connecting rod for the given loading conditions using the FEM Software Ansys
12.1 .In the starting of the work, the static loads acting on the connecting rod, After that the work is carried out for safe design and life in fatigue. Fatigue Analysis is compared with the Experimental results.

Introduction
Connecting rod is the main component of the combustion engines which main purpose are transfer the energy from the pistons to crankshafts and convert the linear, reciprocating motion of a piston into the rotary motion of a crankshaft, from the viewpoint of functionality; also it connects reciprocating piston to rotating crankshaft and some other design connected direct to the crosshead and then transmitting the thrust of the piston to the crankshaft of the Automobile Engine. Where in the automobile engine con rod moves are in both a rotating motion in to a big end and reciprocating motion in to a small end.
Connecting rods must have the highest possible rigidity at the lowest weight. In Automobile internal combustion engine connecting rod is a high volume production component subjected to complex loading.
Bending stresses appear due to eccentricities, crankshaft, case wall deformation, and rotational mass force. Therefore, a connecting rod must be capable of transmitting axial tension, axial compression, and lower cost due to their simple
bending stresses caused by the thrust and pull on the piston and by centrifugal force.
During the operation of the engine, the connecting rod undergoes is prone to tensile, compression, and buckling loading. In many cases, the major reason behind or causing catastrophic engine failure is the occurrence of the connectingrod failure and sometimes, such a failure can be attributed to the broken connecting rods shank especially when there is a probability of being pushed through the side of the crankcase, thereby making the engine irreparable [1].
However, specifically describing such a failure, it is important to point out at the different reasons for this failure such as fatigue near a physical defect in the rod, the overheating of the engine, cracking, lubrication failure in a bearing which is usually caused by inaccurate or faulty maintenance, failure of the rod bolts which is due to defect, improper tightening ,and reuse of already used (stressed) bolts where not recommended. Figure 1 shows the failure on connecting rod [2].

Specification of Existing Connecting rod
Table 1 shows the specifications of the connecting rod for Aluminium 6351 (FM 70 Diesel Engine.). Where tested chemical composition of the material is 0.87%Si, 0.49% Mn, 0.006%Ch, 0.005%Co, 0.013%
Ti, 0.14% Fe, 0.02%Pb, 0.005% Sn, 0.87% Mg, and
97.53% Al.
ENGINE DIMENSION
1
LENGTH (mm.)
355
2
WIDTH (mm.)
552
3
HEIGHT (mm.)
465
4
ENGINE WEIGHT
18
5
SPARK PLUG
MICO M 45 Z48
6
FUEL CAPACITY
4.5 (Lit)
7
OPERATING HOUR
2.3
8
P.T.O. SHAFT
ANTI CLOCKWISE
ENGINE DIMENSION
1
LENGTH (mm.)
355
2
WIDTH (mm.)
552
3
HEIGHT (mm.)
465
4
ENGINE WEIGHT
18
5
SPARK PLUG
MICO M 45 Z48
6
FUEL CAPACITY
4.5 (Lit)
7
OPERATING HOUR
2.3
8
P.T.O. SHAFT
ANTI CLOCKWISE
Table: 1 Specification of Al 6351 Connecting rod
ROTATION
9
METHOD OF STARTING
ROPE / RECOIL
10
FUEL
PETROL (START) DIESEL (RUN)
11
IGNITION SYSTEM
ELECTRONIC

Theoretical Calculation of Connecting Rod
A connecting rod is a Reciprocating part which is deformed in the compressive and tensile forces. Since the Tensile forces are much lower than the Compressive forces, therefore, design of the cross section of the connecting rod we consider the Rankins formula.
Here,
c = Compressive yield stress A = C/S area of Connecting rod L = Length of Connecting rod Pcr = Critical buckling load
Ixx and Iyy =moment of inertia of the section about
xaxis and yaxis respectively.
Kxx and Kyy =radius of gyration of the section about xaxis and y axis respectively. Rankin formula = (Ixx=4Iyy).
3.1 Pressure Calculation for 256 CC Diesel Engine
Let,
Con rod length to crank ratio, l = 3.80
r
Length of the connecting rod= 3.80*33.35 = 127 mm.
Figure 1: ISection of the connecting rod
Thickness of the flange and web of the section = t Width of the section = B =2t
Height of the section = H = 4t
Area of the section = A = [2 (4t * t) + 2t*t] = 6t2
xx
xx
Moment of inertia about x axis = I = 120 t4
12
yy
yy
Moment of inertia about yaxis = I = 18 t4
12
So now, Ixx / Iyy = 6.6667
We know that radius of gyration of the section about Xaxis,
Kxx = Ixx / A = 1.291 t
After deciding the proportions for Isection of the connecting rod, its dimensions are determined by considering the buckling of the rod about Xaxis (assuming both ends hinged) and applying the Rankins formula.
1+a L
1+a L
We know that buckling load, Pcr = c
K xx
By putting all the defined values we can find the,
t4 – 21.37t2 – 88.199 = 0
So, t = 5mm. So by the calculation, B = 10 mm And
Indicated power, I
= BP = 2.238 = 2.7975 Kw.
p
p
0.80
H = 20 mm.
Depth at crank end H
Indicated power, Ip = pm L A n, (n=3000 rpm)
1
Depth at piston end H = 0.8H = 17.2 mm
2
Mean Pressure, pm = 0.4295 N/mm2
C/S area of piston = D2 = A =3848.45 mm2
4
Force on the piston due to gas pressure, Pc =
max c
ax c
D2 P = P = 14876.83 N
4
3.3 Dimension at small end or at piston end
Max gas load = Bearing load, Pc = dp lp pb p
Here lp = 1.7
Critical buckling load is given by, Pcr = Pc fs = 37190 N
3.2 Design Calculation for the Al6351 Connecting rod
The most suitable section for the connecting rod is I section with the proportions as shown in Fig. 1
dp
By making calculation we find out;
Diameter of the piston pin, dp = 15.60 mm. And Length of piston pin, lp = 26.2 mm.

Dimension at big end or crank end
lc
Pc = dc lc pb c,
= 0.89, Bearing pressure of crank
dc
pin, pb c = 18.926 N/mm2 , so we can find out,
dc = Diameter of the crank pin = 30 mm.
lc = Length of crank pin = 26.2 mm.

Big End Cap and Bolts
The maximum force acting on the cap and two bolts consists only of inertia force at the top dead centre on the exhaust stroke. The inertia force acting on the bolts or cap is given by,
P = m 2 r cos + cos 2
i r n1
Where, the angular velocity of the crank is,
= 2N , and the mass of reciprocating parts is
60
given by,mr = [mass of piston assembly + (1/3) mass of connecting rod] = 2.51 kg. = 24.615 N
Figure 2: Relation between Inertia Force to Crank Angle.
Now force acting in the connecting rod at any instant due to effect of inertia force,
The inertia force on the connecting rod will be
maximum at the top dead centre position where ( =
P = P
r
r
cos
. Where sin = sin
n1
0). So when = 0 then cos = 1 and cos 2 = 1 and substituting the above values and get;
Let, Pr = Force on Connecting rod at different crank angle.
= Inclination angle of line of stroke, =crank angle
Pi Max
= mr
2 r 1 + 1 so by solving the equation
n1
get the value of; Pi Max = 10431.1867 N.
d2
So Pi Max =2 co t ,Core dia of bolt dco=6.4mm

Fs
Nominal diameter of the bolt is calculated by, d =
dco = 10 mm
0.7
The distance between the centers of bolts is, Lh = dc + d + CL = 40.2 mm.
It is treated as a beam freely supported at the bolt centers and loaded in a manner intermediate and centrally concentrated load in which case the bending moment is WL/6 so,
b h
b h
M = Pi max . * L = 10431 .1867 * 40.2
6 6
Mb = 69888.95 N.mm
For thickness of the cap (tc) is obtained by,
Figure 3: Crank angles to Force acting on the
connecting rod.


Modeling and FEM Analysis of
= Mb Y , Y = tc , I = bc t 3
pb I
2 12 c
Connecting rod
Solving the above equations we get tc = 7.5 mm
It may be noted that the inertia force of reciprocating parts opposes the force on the piston when it moves during its downward stroke (i.e. when the piston moves from the top dead centre to bottom dead centre). On the other hand, the inertia force of the reciprocating parts helps the force on the piston when it moves from the bottom dead centre to top dead centre. Figure 2 shows the Inertia Force acting on the connecting rod at different crank angles.
Aluminium 6351 Connecting rod was modeled by taking the designed parameter of rod and then by using the Proe Wild Fire 5.0 software solid modeling has done which is shown in Fig.4. And saved within this program in *.IGES format. The model is imported in Ansys and then the mechanical characteristics of the connecting rod are established: density – 2800 kg/m3, Youngs modulus 68.9 GPa, Poissons ratio – 0.3, etc.
the crank end. In the analysis carried out, the axial load was 14.88 KN in both tension and compression. The pressure constants for 14.88 KN are as follows. [8]
Compressive Loading Stress:
Crank End P
= Pc
= 14876 .83
= 25.637 MPa
co rp H1 3
15 22.34 3
Piston End P
= Pc = 14876 .83
= 64.269 MPa
Figure 4: Schematic of a typical Connecting rod [Pro e modal]
co rc H2 3
7.8 17.2 3

Meshing
Tensile Loading Stress:
2
2
Crank End P
= Pc = 14876 .83
=28.263 MPa
Here Meshing element chooses is 10 nodes
to rp H1
2
2
15 22.34
Tetrahedron named Solid187 as shown in fig 5. First
Piston End P
= Pc = 14876 .83
= 70.594 MPa
convergence was checked by finding deformation against different element size then plotting graph of
to rc H2
2
2
2
2
7.8 17.2
deformation versus no of elements. Here element size is found out to be 2mm for working in convergence zone. Total No of element was 13403 and Nodes were 23962. Fig. 6 shown below is meshed model of connecting rod.
Figure 5: Schematic of 10 tetrahedron element shape used for FEA meshing.
Figure 6: Meshed connecting rod in ANSYS Workbench 12.1 software.

Loading and Boundary Conditions
In this study four finite element models were analysed. FEA for both tensile and compressive loads were conducted. Two cases were analysed for each case, one with load applied at the crank end and restrained at the piston pin end, and the other with load applied at the piston pin end and restrained at
Since the analysis is linear elastic, for static analysis the stress, displacement and strain are proportional to the magnitude of the load. Therefore, the obtained results from FEA readily compare with the above values.
So, Net force acting on connecting rod = Gas Forces Inertia Forces Acting on piston
P = 14876.508 10431.1867, P = 4469.6 N
Boundary Condition:
Table 2: Loading Conditions
Type
of Loading
Applied Net force at
Restrained end at
Tensile Loading
Crank end (1800)
Pin end (1800)
Pin end (1800)
Crank end (1800)
Compressive Loading
Crank end (1200)
Pin end (1200)
Pin end (1200)
Crank end (1200)
FEA Result:
The load analysis was carried out to obtain the loads acting on the connecting rod at any given time in the loading cycle and to perform FEA. Most investigators have used static axial loads for the
design and analysis of connecting rods. In the static analysis von Mises stress, critical locations observed under tension and compression loadings at the piston and crank ends. The most highly stressed areas are in the transition regions between the shank and the crank end, as well as the shank and the pin end. Stresses are all symmetric over the entire rod, since geometry and loading were symmetrical.
Figure 7 Equivalent (vonMises) Stress (Tensile load at piston end)
Figure 8 Equivalent (vonMises) Stress (Tensile load at crank end)
Figure 9 Equivalent (vonMises) Stress (Compressive load at piston end)
Figure 10 Equivalent (vonMises) Stress (Compressive load at crank end)
TABLE 3: Comparative result of Equivalent Stress
Type
of Loading
Applied Net force at
Restrain ed
end at
Equivalent (vonMises) Stress (MPa)
Analytical
FEM Analysis
At Crank end
At pin end
At Crank end
At pin end
Tensile Loading
Crank end (1800)
Pin end (1800)
28.263
70.59
24.73
67.68
Pin end (1800)
Crankend (1800)
28.263
70.59
22.44
66.34
Compressive Loading
Crank end (1200)
Pin end (1200)
25.637
64.27
20.90
57.83
Pin end (1200)
Crank end (1200)
25.637
64.27
19.74
61.12


Fatigue Behavior and Life prediction
Fatigue failure of mechanical components is a process of cyclic stress/strain evolution and redistribution in the critical stressed volume. It may be imagined that due to stress concentration (notches, material defect or surface roughness) the local material yields firstly to redistribute the loading to the surrounding material, and then follows with cyclic plastic deformation and finally crack initiates and the resistance is lost. Therefore, the simulation for cyclic stress/strain evolutions and improving the accuracy of fatigue life prediction of mechanical components.
The details view of the fatigue tool is used to define the various aspects of a fatigue analysis such as loading type, handling of mean stress effects and more. Several results for evaluating fatigue are available to the user. Outputs include fatigue life,
damage, factor of safety, stress biaxiality, fatigue sensitivity.
`
Figure 11 Reversible cyclic loading
Figure 12 Fatigue life
Figure 13 Safety Factor
Figure 14 Biaxiality Indication
Fatigue Experimental Analysis
The experimental fatigue analysis has been done for finding the life of and probably to find failure area of the Aluminium 6351 connecting rod.
Figure 15 shows the modal of fatigue Experiment. Where Fig 1617 shows the experimental setup of the analysis.
Figure 15 Aluminium6351 Connecting rod
Figure 16 High and Low Cycle Fatigue machine
Figure 17 Arrangement of jig and Fixture for Experiment
Result of Analysis
Figure 18 Displ. Vs. Load in experimental Fatigue Testing
Figure 19 Stresses vs. Strain in Experimental Fatigue Testing
Figure 20 SN curve of Aluminium 6351
Figure 21 Fatigue test of Spacimen1
Figure 22 Fatigue test of Spacimen2
Figure 23 Fatigue test of Spacimen3
Figure 24 Failure result of 3 specimen
Factor of Safety
Material Properties
Aluminium 6351
Yield Strength (Mpa)
285
Theotitical Factor of Safety
4
Allow Stress (Mpa)
71.25
Ansys Result (Mpa)
67.68
Working Factor of Safety
4.21

Conclusions
Figure 7 to 10 shows the FEM Analysis of the connecting rod of Aluminium 6351 applying load 4469.6 at the crank and piston pin end while loading condition are in tensile and compressive.
As from calculation of Factor of Safety here it is (4.21) appropriate with considering designer experience. The working factor of safety is nearer to theoretical factor of safety in aluminum 6351 connecting rod.
As showing the results the peak stresses mostly occurred in the transition area between pin end, crank end and shank region. The value of stress at the middle of shank region is well below allowable limit.
Based on the results of the experiments obtained by analyzing the finite element in the present study, it can be concluded that the occurrence of the connecting rod failure was due to the fatigue crack growth mechanism which came as a result of higher stress being combined with the porosity (manufacturing defect) in initiation and growth of a fatigue crack followed by catastrophic failure. Finally, lubrication engine system should be regularly checked, and all these are highly recommended to ensure long life connecting rod.
Figure 1824 shows the Ansys and experiments results of the fatigue of three spacimen.Experiment result for spacimen1 life was 3.125Ã—105,Spacimen2 life is 2.984Ã— 105 ,and spacimen3 life is 3.096Ã—105 .where Ansys software result was 3.0817
Ã—105.So here Experimental results are nearly equal to the Analysis software results. So we can now successfully implement this part in to a FM70 Diesel Engine.
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