
Research Article


Fatigue Analysis of Connecting Rod of U650 Tractor in the Finite Element Code ANSYS


M. Omid,
S.S. Mohtasebi,
S.A. Mireei
and
E. Mahmoodi


ABSTRACT

In this study, a finite element routine was first used
to calculate the static displacements and stresses under the maximum compression
and tension loadings in the connecting rod of universal tractor (U650),
which were then used for critical points evaluation. Fatigue analysis
and longevity after a 1000000cycle load, assessed through using of ANSYS
software. Calculations based on fatigue life and accurate loading histories
permit rod to be optimized for durability without the need for expensive
and timeconsuming testing of series of prototypes. According to this
study, the critical point of the connecting rod of U650 is end of the
shank and near piston pin hole (point 46). This point is critical point
not only in tensional loading but also in compressive loading. The allowable
number of cyclic loading on this part under fully reversed loading assumption
is 10^{8} cycles that will be increased by decreasing amount of
the stress concentration coefficient. The results obtained from the present
study can be used to bring about modifications in the process of connecting
rod manufacturing.





INTRODUCTION
Despite the fact that most engineers and designers are aware of fatigue
and that a vast amount of experimental data has been generated on the
fatigue properties of various metallic and nonmetallic materials, fatigue
failures of engineering components are still common. A number of factors
influence the fatigue life of a component in service, viz., (i) complex
stress cycles, (ii) engineering design, (iii) manufacturing and inspection,
(iv) service conditions and environment and (v) material of construction.
The use of calculations and simulations is a key feature of the modern
design process. Several properties such as stress, strength, stiffness,
durability, handling, ride comfort and crash resistance can today be numerically
analyzed with varying levels of accuracy. Development time can be shortened
by ensuring that some, or rather all, of these properties fulfill established
requirements even before the first prototype is being built. Accordingly,
calculations based on fatigue life and accurate loading histories permit
structures and components to be optimized for durability without the need
for expensive and timeconsuming testing of series of prototypes. Thus
designs can be obtained that are less conservative (i.e., better optimized)
than those based on traditional criteria, such as maximum load or stress
for a series of standard load cases (Fermer and Svensson, 2001). The use
of Finite Element Method (FEM) for calculating stress and strain is a
well established procedure in analyzing fatigue and determining longevity.
For example, Del LlanoVizcaya et al. (2006) carried out stress
analysis in the FE code ANSYS and then performed multiaxial fatigue study
of helical compression springs using the fatigue software nCode. Biancolini
et al. (2003) designed a connecting rod based on fatigue analysis.
Beretta et al. (1997) presented a resistant method to failure on
connecting rod design that improved the fatigue life slightly. They found
that the occurrence of fatigue phenomenon is closely related to the appearance
of cycling stresses within the connecting rod body. Lu (1996) presented
an approach to optimize the shape of a connecting rod subjected to a load
cycle which consisted of the inertia load deducted from gas load as one
extreme and peak inertia load exerted by the piston assembly mass as the
other extreme. A FE routine was first used for calculating the displacements
and stresses in the connecting rod, which were then used in another routine
to calculating the total life. Fatigue life was defined as the sum of
crack initiation and crack growth lives, with crack growth life obtained
using fracture mechanics. Rahman et al. (2008) presented the finite
element based fatigue life prediction of a new free piston linear generator
engine mounting. The objective was to assess the critical fatigue locations
on the component due to loading conditions. They concluded that Morrow
mean stress correction method gave the most conservative (less life) results
for crack initiation method.
Very few cases of fatigue failure analysis of agricultural vehicles have
been reported in the literature. Nanaware and Pable (2003) described a
case study on the fatigue fracture of rear axle shafts of 575 DI tractors.
The failure of rear axle shafts was due to inadequate spline root radius,
which led to crack initiation and subsequent crack growth is by fatigue
under the cyclic loading conditions of field operation. In general, the
shafts in power plant systems run with a steady torsion combined with
cyclic bending stress due to selfweight or weights of components or possible
misalignment between journal bearings (Bhaumik, 2002). A similar case
study (Nanaware and Pable, 2003) was reported in Fatigue Design Handbook
AE 10 (Fatigue design handbook, 1988). This case study was of scraper
type tractor rear axle shaft failures. The rear axle shafts were failing
within six months of service, even though durability tests were done in
the laboratory. Fatigue was the predominate mode of failure due to reverse
torque. Pandey (2003) conducted a failure investigation on diesel engine
crankshafts used in tractors which were made from 0.45% carbon steel.
The failure in the crankshafts has been initiated mostly from the crank
pinweb fillet region by a fatigue mechanism. The estimated stress level
for fatigue initiation is in the range of 175 MPa. Rabb (1996) examined
the fatigue failure of a connecting rod in a mediumspeed diesel engine.
The FE analysis and fatigue data led to an improved design of connecting
rod. The difficulties in making a sufficiently good FE model with exact
geometry of fine details and with all important nonlinearities are explained.
Pistek and Novotny (2006) studied on the crank train virtual development
of a new series of a tractor inline fourcylinder diesel engine. The
new series used a modified version of a connecting rod for which the mass,
the reciprocating weight and the rotating weight were, respectively, 28,
46 and 7% less than the original one. They concluded that a fatigue of
the modified connecting rod slightly decreases but is still in the safety
limit. Akinci et al. (2005) examined the failure of the transmission
gear in a rotary tiller. The rotary tiller is attached to a tractor threepoint
linkage system and driven by the tractor PTO. The failure types were abrasion
and plastic deformation. The reasons for the failure were design errors
and material faults. More recently, Bayrakçeken et al. (2007)
carried out failure analysis of crankshafts of two single cylinder diesel
engines. The single cylinder diesel engines are extensively used in agricultural
areas for several purposes such as water pumping or driving some auxiliary
agricultural vehicles. Two different failure cases of crankshafts of these
engines were analyzed. Some characterization studies and fractographic
analysis were also carried out to asses the failure reason. However, the
cranks have some miner design differences, both failures are occurred
after a fatigue process.
In this study, fatigue analysis and longevity of the U650 universal tractor
connecting rod is carried out in the FE code ANSYS. U650 model of universal
tractor is one of the conventional tractors currently being used in Iran.
MATERIALS AND METHODS
Fatigue failure of mechanical components is a process of cyclic stress/strain
evolutions and redistributions in the critical stressed volume. It may
be imagined that due to stress concentration (notches, material defects
or surface roughness) the local material yields firstly to redistribute
the loading to the surrounding material, then follows with cyclic plastic
deformation and finally crack initiates and the resistance is lost. Therefore,
the simulations for cyclic stress/strain evolutions and redistributions
are critical for predicting fatigue failure and improving the accuracy
of fatigue life prediction of mechanical components. In general, the fatigue
process embraces two basic domains of cyclic stressing or straining, as
shown in Fig. 1, differing distinctly in character (Glodez
et al., 2002). In each domain, failure occurs by different physical
mechanisms:
• 
LowCycle Fatigue (LCF)where significant plastic straining
occurs. LCF involves large cycles with significant amounts of plastic
deformation and relatively short life. The analytical procedure used
to address straincontrolled fatigue is commonly referred to as the
strainlife, crackinitiation, or critical location approach 
• 
HighCycle Fatigue (HCF)where stresses and strains are largely
confined to the elastic region. HCF is associated with low loads and
long life. The stresslife (SN) or total life method is widely used
for highcycle fatigue applicationshere the applied stress is within
the elastic range of the material and the number of cycles to failure
is large. While lowcycle fatigue is typically associated with fatigue
life between 10 to 100,000 cycles, HCF is associated with life greater
than 100,000 cycles 
The model for the fatigue crack initiation presented here is based on
the continuum mechanics approach, where it was assumed that the material
is homogeneous and isotropic, i.e., without imperfections or damages.
Methods for analysis in that case is usually based on the CoffinManson
relation between deformations (ε), stresses (σ) and the number
of loading cycles (N). The strainlife method (εN) is usually used
to determine the number of stress cycles N required for the fatigue crack
initiation, where it is assumed that the crack is initiated at the point
of the largest stresses in the material.

Fig. 1: 
Strainlife (εN) method for the fatigue crack
initiation (Lu, 1996) 
The CoffinManson formulation
can be used for life predictions and the total cyclic strain range Δε,
comprises two components (elastic and plastic cyclic strain range Δε_{e} and Δε_{p}), can be described as (Glodez et al.,
2002; Bannantine et al., 1990):
where, E is the Young`s modulus, σ_{F} is the fatigue strength
coefficient, ε_{F} is the fatigue ductility coefficient,
b is the exponent of strength and c is the fatigue ductility exponent,
(Fig. 1). Eq. 1 can only be solved
numerically. The number of stress cycles N required for the fatigue crack
initiation can then be solved iteratively from Eq. 1
with the appropriate material parameters E, ε_{F}, σ_{F},
b and c.
Case study: U650 universal tractor connecting rod: The model discussed
above has been used for the computational determination of the fatigue
life of U650 universal tractor connecting rod with complete data set given
in (Table 1). The connecting rod is made of high strength
alloy steel with Young`s modulus E = 207 GPa and Poison`s ratio v = 0.3.
The majority of connecting rods are made of ductile metallic material
(Whittaker, 2001).
For carrying out the fatigue analysis, it is necessary to determine the
maximum load. In this study, for calculating the maximum load, the mean
effective pressure of the cylinder and mechanical yield of the engine
were assumed to be 0.8 MPa and 0.7, respectively (Artamonov et al.,
1976). The maximum load was thus found to be 9500 N. We also have to specify
fully reversed loading to create alternating stress cycles (Hancq, 2003;
Hancq et al., 2000), i.e., measure the tension/compression asymmetry
of the yield strength on connecting rod under fully reversed loading (load
is applied, removed, then applied in the opposite direction with a max
loading of 9500 N). Since this is a stresslife fatigue analysis, no mean
stress theory needs to be specified because no mean stress will exist
under fully reversed loading.
Table 1: 
Universal tractor (U650) connecting rod data for the
studied rod 


Fig. 2: 
Geometric dimensions of the connecting rod 

Fig. 3: 
Meshed model of the connecting rod 
FEmesh: In the next stage of the analysis, one virtual model
of the connecting rod was constructed. The geometric dimensions of the
connecting rod were derived from the original connecting rod of the U650
universal tractor with 0.1 mm accuracy (Fig. 2). The
virtual model was then prepared by SolidWorks software. The model was
then called in the ANSYS Perp7 processor and the parameters of material
property of the connecting rod (Table 1) were entered
into the software. For achieving high accuracy, a 10node tetragonal element
(SOLID92) was used (Fig. 3) (Hancq, 2003). Each node
in SOLID92 element has a six degree of freedom behavior and is well suited
to model irregular meshes (Rabb, 1996).
Numerical analysis: In order to carry out the analysis in the
FE code ANSYS (Hancq, 2003; Hancq et al., 2000) it is necessary
to import geometry, apply boundary conditions and loading corresponding
to the maximum developed load, i.e., 9500 N. The response of connecting
rod to fully reversed, tension/compression loading is investigated by
examining their fatigue lives, critical damage initiation points and their
residual strength. Initially the maximum load is exerted in the tensional
manner and by applying the boundary conditions, the FE results is obtained.
Next, the tensional loading is removed and the maximum load is exerted
in compression mode. Since VonMises stress is being used to compare against
fatigue material data, in both runs, the Von Misses stresses are activated
by using POST1 option in ANSYS. Now we can perform stress and fatigue
calculations for both cases. In each case, the model is solved and the
critical points are identified. The simulation results are shown in Fig.
4 and 5. The critical points, according to Fig.
4 and 5, are located at points 46 (end of the shank,
near to piston pin hole), 5232 (lateral and inner face of crank end of
the connecting rod) and 4887 (end of the shank and near to crank hole)
in tensional loading and point 46 again in compression loading of the
connecting rod.

Fig. 4: 
(a) Von Misses stress contour in tensional loading and
critical points at (b) point 46, (c) point 4887 and (d) point 5232 

Fig. 5: 
(a) Von Misses stress contour in compressive loading
and (b) critical point at point 46 
Having determined the critical points, the fatigue analysis was focused on
them. A value of 1.25 was given for stress concentration coefficient to account
for the differences between the original connecting rod and the modeled type
(Rabb, 1996). Finally, a factor of safety for a design life of 1,000,000 cycles
was considered for the loading cycle (Hancq, 2003). The partial usage which
is an indication of cumulative fatigue usage defined as the ratio of cycle used/allowed,
is also determined. A summary of the results for the critical points is shown
in Fig. 6.

Fig. 6: 
Obtained results for fatigue calculation on points 46,
5232 and 4887 in tensional loading 
RESULTS AND DISCUSSION
Based on the FE analysis of the connecting rod, it is found that, in
the tensional loading, the maximum stress was 29.4 MPa (Fig.
4). Maximum stress in compressive loading was 24 MPa (Fig.
5). Critical points of the FE model were at the points 46, 5232 and
4887 in tensional loading and point 46 in compressive loading (Fig.
4, 5). According to the results shown in Fig.
6, the cyclic stresses on points 46, 5232 and 4887 are 26.793, 4.167
and 12.994 MPa, respectively. The partial usage which is an indication
of ratio of the loading cycle number to allowable cycles is 0.01 for all
three points. Also, with respect to these results, the allowable number
of cyclic loading on the connecting rod is found to be 10^{8}
cycles.
CONCLUSIONS
According to this study, finite element method is evaluated as a useful
approach to recognize the critical points and fatigue life time of the
reciprocating components such as connecting rods. In the case of U650
tractor connecting rod the critical point, is point 46, i.e., at the end
of the shank and near piston pin hole. This point is critical point not
only in tensional loading but also in compressive loading. The allowable
number of cyclic loading on this part under fully reversed loading assumption
is 10^{8} cycles. In order to improve on fatigue life of the connecting
rods of tractors further, this value may be increased by decreasing the
stress concentration coefficient.
ACKNOWLEDGMENTS
The authors would like to thank the financial support provided by the
Research Department of University of Tehran.

REFERENCES 
Akinci, I., D. Yilmaz and M. Canakci, 2005. Failure of a rotary tiller spur gear. J. Eng. Failure Anal., 12: 400404. CrossRef 
Artamonov, M.D., V.A. Ilarionov and M.M. Morin, 1976. Motor Vehicles. 1st Edn. Mir Publisher, Moscow.
Bannantine, J., J. Comer and J. Handrock, 1990. Fundamentals of Metal Fatigue Analysis. 1st Edn. Prentice Hall, New Jersey, ISBN10: 013340191X.
Bayrakceken, H., S. Tasgetiren and F. Aksoy, 2007. Failures of single cylinder diesel engines crank shafts. J. Eng. Failure Anal., 14: 725730. CrossRef 
Beretta, S., A. Blarasin, M. Endo, T. Giunti and Y. Murakami, 1997. Defect tolerant design of automotive components. J. Fatigue, 19: 319333. CrossRef 
Bhaumik, S.K., 2002. Fatigue fracture of crankshaft of an aircraft engine. J. Eng. Failure Anal., 9: 255263. CrossRef 
Biancolini, M.E., C. Brutti, E. Pennestrì and P.P. Valentini, 2003. Dynamic, mechanical efficiency and fatigue analysis of the double cardan homokinetic joint. J. Vehicle Design, 33: 4765. CrossRef 
Del LlanoVizcaya, L., C. RubioGonzález, G. Mesmacque and T. CervantesHernández, 2006. Multiaxial fatigue and failure analysis of helical compression springs. J. Eng. Failure Anal., 13: 13031313. CrossRef 
Fermer, M. and H. Svensson, 2001. Industrial experiences of FEbased fatigue life predictions of welded automotive structures. J. Fatigue Fract. Eng. Mater Struct., 24: 489500. CrossRef  Direct Link 
Glodez, S., M. Sraml and J. Kramberger, 2002. A computational model for determination of service life of gears. J. Fatigue, 24: 10131020. CrossRef 
Hancq, D.A., 2003. Fatigue Analysis Using ANSYS. ANSYS Inc., 22 P. http://www.ansys.com/
Hancq, D.A., A.J. Walters and J.L. Beuth, 2000. Development of an objectoriented fatigue tool. J. Eng. Comput., 16: 131144. CrossRef  Direct Link 
Lu, P.C., 1996. The optimization of a connecting rod with fatigue life constraint. Int. J. Mat. Prod. Technol., 11: 357370. Direct Link 
Nanaware, G.K. and M.J. Pable, 2003. Failures of rear axle shafts of 575 DI tractors. J. Eng. Failure Anal., 10: 719724. CrossRef 
Pandey, R.K., 2003. Failure of dieselengine crankshafts. J. Eng. Failure Anal., 10: 165175. CrossRef 
Pistek, V. and Novotny P., 2006. Virtual development of fourcylinder diesel engine. Proceedings of the 24th CADFEM User's Meeting, Intl. Cong. FEM Tech. with 2006 German ANSYS Conference, October 2527, 2006, Schwabeniandhalle Stuttgart, Germany, pp: 101122.
Rabb, R., 1996. Fatigue failure of a connecting rod. J. Eng. Failure Anal., 3: 1328. CrossRef 
Rahman, M.M., A.K. Arffin, N. Jamaludin, S. Abdullah and M.M. Noor, 2008. Finite element based fatigue life prediction of a new free piston engine mounting. J. Applied Sci., 8: 16121621. CrossRef  Direct Link 
Whittaker, D., 2001. The competition for automotive connecting rod markets. J. Metal Powder Report, 56: 3237. CrossRef 



