Engineering Failure Analysis 19 (2012) 77–86

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Engineering Failure Analysis

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Spalling investigation of connecting rod

Saharash Khare, O.P. Singh ⇑, K. Bapanna Dora, C. SasunTVS Motor Company Ltd., Research & Development, Hosur, Tamilnadu 635 109, India

a r t i c l e i n f o

Article history:Received 28 April 2011Received in revised form 7 September 2011Accepted 12 September 2011Available online 18 September 2011

Keywords:Connecting rodSpallingWearFinite element modelContact pressure

1350-6307/$ - see front matter � 2011 Elsevier Ltddoi:10.1016/j.engfailanal.2011.09.007

⇑ Corresponding author. Address: TVS Motor Com4344 276780x3502, +914344270502 (Direct); fax: +

E-mail addresses: omprakash.singh@tvsmotor.co

a b s t r a c t

The customers of the vehicles reported high noise and vibration in the engine at an earlystage of service lifetime. An analysis of the various components of the internal combustionengine was carried out. Subsurface cracks and pit marks were seen in the crank pin, rollerbearings and big end surfaces of the connecting rod. It was found that high wear at theinterface of these components was the main culprit. A laboratory test set-up was devel-oped to correlate and reproduce the field failures. The loads and boundary conditionsobtained from the experiments were used in the finite element model of the connectingrod assembly. Results shows high interfacial pressure and stresses near the junction ofweb and flange of the connecting rod. The modified design of the connecting rod shows sig-nificant reduction in the extreme pressure in FEM resulting in the significant enhancementof durability life in laboratory test. A discussion of the spalling problem has been providedleading to the connection of pick pressure and spalling phenomena.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction and failure background

One source of energy in automobile industry is internal combustion engine. IC engine converts chemical energy intomechanical energy in form of reciprocating motion of piston [1]. Crankshaft and connecting rod convert reciprocating motioninto rotary motion. The connecting rod experiences various forces of piston acceleration and deceleration from the strokingmotion, loads generated by friction and the load by the cylinder pressure during the combustion stroke [2,3]. Under normaloperating condition, the design of the connecting rod for infinite number of cycles is based on the these forces such that de-signed connecting rod does not exceed the desired strength during the life span. However, when the customer exceeds therev limit of the engine, it adds to a larger boost levels, increasing the cylinder pressure and thus causing higher friction be-tween the sliding members. Catastrophic engine failure results if the sign of failure modes are not investigated early [4–7]. Inthis paper we report the failure of the connecting rod and the pin due to spalling observed under such condition of motor-cycle engine.

Crankshaft assembly is shown in Fig. 1. The big end of the connecting rod is connected to the crankshaft assemblyusing a crank pin. To reduce the friction, roller bearing is used between the connecting rod and the pin. The roller bearingshave non-conformal contact surfaces (line or point contact called hertzian contacts), which support the moving parts ofthe crankshaft assembly with rolling action at the interface [8]. An oil film between the interfaces prevents metal-to-metalcontact. The customers reported failure of these components after an average driving distance of 15,000 km against thedesired life of 50,000 km. Failed sample of connecting rod, pin and the roller bearing is shown in Fig. 2. Scratches andrough marks on connecting rod, crank-pin and crankpin bearing can be seen with naked eyes. Such kind of failure is gen-erally referred as fatigue spalling. Fatigue spalling initiates in the depth of material by subsurface stress field close to the

. All rights reserved.

pany Ltd., R&D (Design analysis group), P. Box No. 4, Harita, Hosur, Tamilnadu 635 109, India. Tel.: +9191 4344 276649.

.in, om.prakashh.singh@gmail.com (O.P. Singh).

Connecting rod

Piston

Crankshaft

Crank pin

Roller bearing

Fig. 1. Computer aided design of the crankshaft assembly showing various components.

78 S. Khare et al. / Engineering Failure Analysis 19 (2012) 77–86

contacting surface [9]. It then grows to the surface in the form of micro-spall or crack [10]. When these localized inelasticsubsurface deformations such a cracks interact, the material break up, disrupt the oil film in the region of highest load andlowest oil film thickness [8].

The interesting and baffling part of the failure is that high wear is seen near the I-beam section of the connecting rod(Fig. 2c). Noise levels radiated from the engine was reported to be high when the failure occurred. In this paper, the factors,which conduce to this fatigue spalling of the connecting rod, has been investigated in this paper.

2. Initial actions and laboratory test development

2.1. Visual inspection

A close inspection of the failed components reveals the following modes of failure (see Fig. 3):

� Surface defects like burr, surface mark, pit mark, dents have observed on crankpin. Fig. 3a shows damage marks on crank-pin surface which acts as a stress concentration location [11,12].� Similarly burr and deep machine tool mark near oil hole (Fig. 3b). This can act as a stress raiser during rolling contact

phenomenon [8].

These observation suggests that external contaminants like hard particles, wear debris, and flakes that might have en-tered into the sliding zone. The hard particles abrasive contaminants cause indentation or cutting wear damage to the softermaterial. The average depth of the cavities measured at the contact surfaces was about 25 lm. It is nearly impossible to ob-serve these kinds of failures during the engine running condition. Hence, different probable causes of failures are identified.The probable causes could be many; few important factors may be due to:

� Hardness of connecting rod and crank pin.� Connecting rod to crank pin bearing clearance is more then specified limit.� High stresses at the critical location.� Design of the connecting rod.� Scarcity of the lubricating oil at sliding interfaces.

Before we investigate these failures further, a laboratory test method was developed and the failures observed in the fieldare reproduced first.

2.2. Laboratory test development

Durability testing under real environment is time consuming. To reduce testing time connecting rod spalling failure hassimulated on an engine dynamometer. The controlled testing method was established based on the four-quadrant matrix

Fig. 2. Failed pictures of (a) roller bearing spalling (a) crankpin spalling, and (b) connecting rod spalling.

S. Khare et al. / Engineering Failure Analysis 19 (2012) 77–86 79

method (e.g.) [13]. This method basically correlates the MTTF (mean time to failure) in field-testing with the MTTF in labtesting. All the failure modes of the system components are identified and failure modes are correlated with the components.In lab test, vehicle was run on dynamo and followed standard running cycle (as shown in Fig. 4) with specified engine oper-ating condition (like load on vehicle, and engine oil quantity). The time period of vehicle acceleration and deceleration con-stitute a cycle. Note that the deceleration time is smaller compared to the acceleration time. The engine speed at point A andB are 20 km/h and 60 km/h respectively. The time duration for AC is 18 s and for CD is 6 s. Engine operating condition wasdecided based on real durability test condition. Noise radiated from the engine was monitored. The normal engine (at thestart of the test) radiated noise in the range of 70–75 dB. When the noise levels was increased to 90 dB, the engine wasstopped, parts were dismantled and the condition of connecting rod and pin was checked. At such high noise level, highvibration of the engine was also noticed. The same failure (see Fig. 2) was reproduced at about 20,000 cycles in the lab test.The same process was repeated for other engine tests.

2.3. Process improvement

Surface defects as seen in Fig. 3 will act as a stress raiser during service of crank pin and connecting rod, and thus wearwill initiate at this defect and shorten the life of the component. Machining process, cleaning process and material handling

Deep machine tool mark

Burr

(a)

(b)

Fig. 3. (a) Surface damage on the crankpin and (b) Surface defect near oil hole.

Fig. 4. Engine running cycle developed for the laboratory test.

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process was improved to avoid surface defect on crankshaft and crankpin. However, no significant improvement in the lifethe component was observed.

2.4. Material improvement

Case depth hardening of the material was measured and improved as shown in Fig. 5. Near the surface, the hardness wasbelow 800 HV and it was consistently low throughout the material depth of 1.3 mm with steep variation across the depth.The supplier process of heat treatment and material check controls were improved. The case depth was increased through-out the thickness with less steep variation.

2.5. Connecting rod to crank pin clearance

In failed sample clearance between connecting rod and roller bearing was not as per the specification. The measuredclearance was 15 lm. With high clearance the crankpin can float and increase the rate of wear [8]. The machining processwas improved to match design specification accurately. The new design clearance specification was controlled to 12 lm.

Table 1Contact parameters in the contact FEM.

Contact type Surface to surface (rigid flexible contact)Contact element CONTA 174Target element TARGE 170Contact detection point Gauss pointContact algorithm Augmented lagrangianInitial penetration 10 lmEffect of initial penetration Include offset only (exclude initial geometrical penetration or gap) with ramped effectContact stiffness 2.7792e6 N/mmPenetration tolerance 0.15 mmPinball region 6.045

350

450

550

650

750

850

950

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 1.1 1.2 1.3 1.4Distance from surface (mm)

Har

dnes

s (H

V 0.

3 kg

) After improvementBefore imrovement

Fig. 5. Case depth hardness variation showing before and after improvement in the connecting rod.

Forc

e (N

)

Crank Angle (degree)

Pressure force

Inertia force

Crankpin

(a) (b)

Fig. 6. (a) Finite element model of the connecting rod and (b) Variation of pressure and inertia forces during engine operation.

S. Khare et al. / Engineering Failure Analysis 19 (2012) 77–86 81

82 S. Khare et al. / Engineering Failure Analysis 19 (2012) 77–86

2.6. Crankpin and roller bearing diameter

Hertzian contact theory explains that contact pressure developed between two cylindrical rolling surface is inversely pro-portional to contact length and diameter of rolling surfaces. High pressure at the interface can result in the wear of thematerials [8]. Eq. (1) shows the relations between rolling cylinder parameter and contact pressure develop between the sur-faces. Maximum pressure between two rolling cylinder surfaces is given by [8],

F

Pmax ¼2Fpal

ð1Þ

where,

a ¼ 4Pðk1 þ k2ÞR1R2

lðR1 þ R2Þ

� �12

Contact pressure (MPa)

VonMises stress

(a)

(b)

ig. 7. Results from FEM of the original connecting rod showing (a) pressure contours in the contact patch region and (b) VonMises stresses.

S. Khare et al. / Engineering Failure Analysis 19 (2012) 77–86 83

ki ¼1� m2

i

pEi

subscript 1 and 2 denotes first and second cylinder respectively, F is the normal load; R is cylinder diameter; E is elastic mod-ulus; m is poisson’s ratio, l is contact length between two cylinder. Note that increasing the parameter a reduces the maxi-mum pressure and hence, the stresses between the two mating cylinders.

To reduce stresses, crankpin and roller bearing diameter was increased from 25 mm to 26 mm and 3 mm to 4 mm respec-tively while maintaining the same design clearance. As per the hertzian contact theory this increase in roller diameter wouldincrease contact diameter by 14%, which in turn would reduce pressure at interfaces by 14%. After all initial checks andimprovements crankshaft assembly was tested in engine endurance test. Similar connecting rod and crankpin-spalling fail-ure was observed at MTTF of 30,000 cycles. Though a 50% improvement in life was achieved, it is still significantly less thanthe desired life of 50,000 km. Hence; further investigations were conducted to improve the design of the connecting rod. Inthe next section we present the finite element model (FEM). The main purpose of the FEM study is to understand the effect ofconnecting rod design profile on contact behavior under engine operating condition.

3. Non-linear finite element analysis

3.1. FEA model

A simplified non-linear contact model was developed to investigate the contact pattern between connecting rod andcrankpin using Ansys code [14]. Rigid to flexible surface-to-surface contact mechanics was used. Since the crankpin hardnessis much more than connecting rod, the connecting rod as considered as flexible surface and crankpin as rigid surface. Theintersection face between connecting rod big end inner surface and crankpin outer surface is modeled as contact patch.FEA contact patch consist of contact and target surfaces; connecting rod big end inner surface was considered as contact (ele-ment CONTA 174) and crankpin outer surface as target (element TARGE 170). Table 1 shows values of other parameters ofthe model. Connecting rod was meshed with second order solid 92 elements. Fig. 6a shows the finite element model withloads and boundary conditions. Crankpin is modeled as rigid body. Small end side of the connecting rod is constrained inlateral and transverse direction to avoid rigid body motion. Force on the connecting rod is obtained from the cylinder pres-sure (measured experimentally) and inertial forces of the connecting rod. Fig. 6b shows the variation of cylinder pressure andinertial forces. Maximum force experienced by the connecting rod was applied as shown. Effect of thermal distortion has notbeen considered in the model.

3.2. FEA results

Fig. 7a show contact pressure distribution between connecting rod and crankpin of the existing design. The pressure dis-tribution is localized and it is seen near the I-section of the connecting rod. The maximum contact pressure is 487.51 MPa.Von-misses tress distribution shows similar distribution with maximum value of 271.44 MPa (Fig. 7b). It is interesting tonote that high magnitude of pressure and stresses are observed near the beginning of the flange and end of the web ofthe I-section of the connecting rod. This is the location where high rate of spalling was observed (see Fig. 3c). An interpre-tation of results suggests that (a) at maximum engine load condition, the load is affectively share by only two central roller

Fig. 8. Four different designs of connecting rod. Notice how the final web design is different from the previous three design.

Contact pressure (MPa)

VonMises stress

(a)

(b)

Fig. 9. Results from FEM of the final design of the connecting rod showing (a) pressure contours in contact patch region and (b) contours of VonMisesstresses.

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bearings resulting in high-localized pressure and (b) The high-localized contact pressure would reduce the lubricating oilfilm thickness and eventually lead to asperity interactions resulting in lower fatigue life.

The high and localized pressure and stresses at the interface may not induce catastrophic failure of the crankshaftassembly. However, this would act as a feeding source for accelerated wear of the interfaces and hence the consequentreduction of useful life. Any further design modifications of the connecting rod should reduce the high interfacial contact

Table 2Magnitude of maximum pressure and stress at the interface between the connecting rod and crankpin.

Parameter Original design (MPa) New design (MPa) Change (%)

Max. contact pressure 487.51 323.40 33.70Max. VonMises stress 271.44 229.08 15.60

Fig. 10. Elastic strain contours in four designs investigated.

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pressure and stresses. Various design modifications were investigated using FEA and they are shown in Fig. 8. It is to benoted how the design of web and flange has changed progressively from the original design 1 to final design 4. Such pro-file is expected to reduce the effective stiffness of the connecting rod at the roller bearing and connecting rod interface.The web area has increased when compared to the original design. The same FE model was used in the analysis. Fig. 9ashows the pressure contours in the contact patch region. Following points are noted: (a) the contact patch area has in-creased compared to the original design. Hence, the maximum pressure has reduced to 323.40 MPa, a reduction of 34%from the original design (see Table 2), (b) It can be seen from the contours that at a given instant of time, at least fourroller bearing would share the maximum load. The magnitude of maximum VonMises stress (Fig. 9b) also reduced to192 MPa, a reduction of 16% is achieved. Fig. 10 shows the non-dimensional displacements i.e. elastic strain in all the fourdesigns (see Fig. 8) at the location where failure was observed. Elastic strains in design 2, 3 and 4 are respectively 3.4, 4.9and 5.7 times the elastic strain in design 1. Higher deformation in design 4 is due to the reduction in stiffness that man-ifested in significant reduction in interfacial pressure. It is to be noted further that how the concentrated deformation indesign 1 has spread to a much larger areas in design 4. The final design of the connecting rod passed the durability lifetarget in the lab test.

4. Discussion

The failure under investigation is a case of surface fatigue resulted from the high concentrated interfacial pressure andstresses. The roller bearing operates with EHD (elasto-hydrodynamic lubrication regime [15]. The hertzian or point contactsbetween the roller bearing and connecting rod produce extreme pressures and therefore elastically deform the surfaces lo-cally to provide small elliptical contact areas. The repetitive and cyclic formation of these elastically deformed contacts even-tually lead to the surface fatigue (see Fig. 2). The tendency of the roller element bearings to float on the hydrodynamic filmwould reduce under excessive interfacial pressure, as viscous drag will no longer be able to draw sufficient amount of oil intothe load zone. Heat generated due to viscous dissipation at the interface would further aggravate the problem [2]. The bear-ing surface may reach the temperature at which the material locally melts.

It is to be noted heat flux generated by viscous shear in the bearing is proportional to the pressure field around the bear-ing surface. Hence rise in temperature can be approximated by [8],

dT ¼ P_QqC

ð2Þ

where, dT is the rise in temperature, P is pressure, _Q is the fill ratio, q is density and C is the heat capacity. In the two designsof connecting rod discussed above (see Fig. 8), and assuming the denominator remains the same in Eq. (2), the decrease in

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bearing surface temperature in final design would be about 33.70% (see Table 2) lower than the original design. This is sub-stantial in the light of the fact that 50% reduction in oil life for every 10 �C increases in temperature [16]. If the bearing sur-faces continues to operate under high load and speed for prolonged period, lubrication oxidation may breakdown due toelevated temperature [17]. This will set the stage for very fine particles to be liberated from the rolling element and finallymaking way into the load zone. Further surface craters will be created that would act as stress concentration sites (see Fig. 3).The distinct region of pressure spike seen in the original design of the connecting rod (Fig. 7) would lead into the boundaryregime of lubrication (metal-to-metal contact) triggering the spalling related phenomenon as described earlier. Though theinitial actions taken as described earlier helped in improving the component life, it was the change in design of the connect-ing rod, which improved the desired life significantly. This corroborates the fact that durability life of the componentsexperiencing alternating forces is improved substantially by proper design [18].

The change in the profile of the connecting rod (Fig. 8b) has facilitated in reducing the extreme interfacial pressure(Fig. 9). The participation of more effective load sharing surfaces (Fig. 10) would result in reduced elliptical contact areasformation and hence, improving the lubrication. In the likely scenario of the engine being abused or revved for a significantspan of time, such design would provide enhanced the life the components especially when spalling related phenomenon isdominant. Four samples of this design of the connecting rod was tested in the lab under same operating conditions andpassed the required durability life. No abnormal noise and vibration was observed even after 50,000 km of continuousrun of the engine.

5. Conclusions

The spalling of connecting rod, crank pin and roller bearing is attributed to the high-localized interfacial pressure thatdeveloped due to the design of the web and flange of the connecting rod. Load sharing interface area was less with onlytwo roller bearings effectively participating at a given instant. The high pressure resulted in accelerated surface fatigue spall-ing leading to the early failure of the components. The modified design of the web provided significant reduction in pressureand stresses in the contact patch region, hence improving the durability considerably.

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