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Indian Streams Research Journal
Vol.2,Issue.IV/May; 12pp.1-4

Prof. N. G. Alvi Research Papers

ISSN:-2230-7850

“Dynamic Load and Stress Analysis of a Crankshaft”

Dr. S. V. Deshmukh Prof.& Head, Bapurao Deshmukh College of Engineering, Sevagram, Wardha, India

Ram.R.Wayzode Lecturer, Suryodaya College of Engineering & Technology, Nagpur, India

Prof. N. G. Alvi Prof,Suryodaya College of Engineering & Technology, Nagpur,India

ABSTRACT
In this study a dynamic simulation was conducted on a crankshaft from a single cylinder four stroke engine. Finite element analysis was performed to obtain the variation of stress magnitude at critical locations. The pressure-volume diagram was used to calculate the load boundary condition in dynamic simulation model, and other simulation inputs were taken from the engine specification chart. The dynamic analysis was done analytically and was verified by simulation in ADAMS which resulted in the load spectrum applied to crank pin bearing. This load was applied to the FE model in ABAQUS, and boundary conditions were applied according to the engine mounting conditions. The analysis was done for different engine speeds and as a result critical engine speed and critical region on the crankshaft were obtained.

Stress variation over the engine cycle and the effect of torsional load in the analysis were investigated. Results from FE analysis were verified by strain gages attached to several locations on the crankshaft. Results achieved from aforementioned analysis can be used in fatigue life calculation and optimization of this component. INTRODUCTION Crankshaft is a large component with a complex geometry in the engine, which converts the reciprocating displacement of the piston to a rotary motion with a four link mechanism. This study was conducted on a single cylinder four stroke cycle engine.ƒnRotation output of an engine is a practical and applicable input to other devices since the linear displacement ofan engine is not a smooth output as the displacement is caused by the combustion of gas in the combustion chamber. A crankshaft changes these sudden displacements to a smooth rotary output which is the input to many devices such as generators, pumps, compressors. A detailedprocedure of obtaining stresses in the fillet area of a crankshaft was introduced by Henry et al. [1], in which FEM and BEM (Boundary Element Method) were used. Obtained stresses were verified by experimental results on a 1.9 liter turbocharged diesel engine with Ricardo type combustion chamber configuration. The crankshaft durability assessment tool used in this study was developed by RENAULT. The software used took into account torsional vibrations and internal centrifugal loads. Fatigue life predictions were made using the multiaxial Dang Van criterion. The procedure developed is such it that could be used for conceptual design and geometry optimization of crankshaft. Guagliano et
Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Research Journal
Vol.2,Issue.IV/May; 2012

al. [2] conducted a study on a marine diesel engine crankshaft, in which two different FE models were investigated. Due to memory limitations in meshing a three dimensional model was difficult and costly. Therefore, they used a bi-dimensional model to obtain the stress concentration factor which resulted in an accuracy of less than 6.9 percent error for a centered load and 8.6 percent error for an eccentric load. This numerical model was satisfactory since it was very fast and had good agreement with experimental results. Payer et al. [3] developed a two-step technique to perform nonlinear transient analysis of crankshafts combining a beam-mass model and a solid element model. Using FEA, two major steps were used to calculate the transient stress behavior of the crankshaft; the first step calculated time dependent deformations by a step-by-step integration using the newmark-beta method. Using a rotating beam-mass-model of the crankshaft, a time dependent nonlinear oil film model and a model of the main bearing wall structure, the mass, damping and stiffness matrices were built at each time step and the equation system was solved by an iterative method. In the second step those transient deformations were enforced to a solid-elementmodel ofthe crankshaft to determine its time dependent stress behavior. The major advantage of using the two steps was reduction of CPU time for calculations. This is because the number of degrees of freedom for because the number of degrees of freedom for because the number of degrees of freedom for solid element model for step two needed only to be built up once.In order to estimate fatigue life of crankshafts, Prakash et al. [4] LOAD ANALYSIS The crankshaft investigated in this study is shown in Figure 1 and belongs to an engine with the configuration shown in Table 1 and piston pressure versus crankshaft angle shown in Figure 2. Although the pressure plot changes for different engine speeds, the maximum pressure which is much of our concern does not change and the same graph could be used for different speeds [9]. The geometries of the crankshaft and connecting rod from the same engine were measured with the accuracy of 0.0025 mm (0.0001 in) and were drawn in the I-DEAS software, which provided the solid properties of the connecting rod such as moment of inertia and center of gravity (CG). These data were used in ADAMS software to simulate the slider-crank mechanism. The dynamic analysis resulted in angular velocity and angular acceleration of the connecting rod and forces between the crankshaft and the connecting rod.

Figure 1: Crankshaft geometry and bending (Fx),torsional (F ), and longitudinal (F ) force directions Table 1: Configuration of the engine to which the crankshaft belongs

Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Reserach Journal
Vol.2,Issue.IV/May; 2012

Figure 2: Piston pressure versus crankshaft angle diagram used to calculate forces at the connecting rod endsThere are two different load sources acting on the crankshaft. Inertia of rotating components (e.g. connecting rod) applies forces to the crankshaft and this force increases with the increase of engine speed. This force is directly related to the rotating speed and acceleration of rotating components. Variation of angular acceleration and angular velocity of the connecting rod Total for the engine speed of 3600 rpm is shown in Figure 3.The second load source is the force applied to the crankshaft due to gas combustion in the cylinder. The slider-crank mechanism transports the pressure applied to the upper part of the slider to the joint between crankshaft and connecting rod. This transmitted load depends on the dimensions of the mechanism.

Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Reserach Journal
Vol.2,Issue.IV/May; 2012

Figure 3: Variation of angular velocity and angular acceleration of the connecting rod over one complete engine cycle at a crankshaft speed of 2800 rpm Forces applied to the crankshaft cause bending and torsion. Figure 1 demonstrates the positive directions and local axis on the contact surface with the connecting rod. Figure 4 shows the variations of bending and torsion loads and the magnitude of the total force applied to the crankshaft as a function of crankshaft angle for the engine speed of 3600 rpm. The maximum load which happens at 355 degrees is where combustion takes place, at this moment the acting force on the crankshaft is just bending load since the direction of the force is exactly toward the center of the crank radius (i.e. Fy = 0 in Figure 1). This maximum load situation happens in all types of engines with a slight difference in the crank angle. In addition, most analysis done on engines with more cylinders (e.g. 4, 6, and 8) is on a portion of the crankshaft that consists of two main journal bearings, two crank webs, and a connecting rod pin journal. Therefore, analysis done for this single cylinder engine can be extended to larger engines.

Figure 4: Bending, torsional, and the resultant force at the connecting rod bearing at the engine speed of 3600 rpm In many studies the torsional load is neglected for the load analysis of the crankshaft, and this is because torsional load is less than 10 percent of the bending load [10]. In this specific engine with its dynamic loading, it is shown in the next sections that torsional load has no effect on the range of von Mises stress at the critical location. The main reason of torsional load not having much effect on the stress range is that the maxima of bending and torsional loading happen at different times (see Figure 4). In addition, when the peak of the bending load takes place the magnitude of torsional load is zero.Figure 5 compares the magnitude of maximum torsional and bending loads at different engine speeds. As can be seen in this figure, the maximum of total load magnitude, which is equal to the maximum of bending load decreases as the engine speed increases. The reason for this situation refers to the load sources that exist in the engine at 355 degree crank angle. At this crank angle these two forces act in opposite directions. The force caused by combustion which is greater than the inertia load does not change at different engine speeds since the same pressure versus crankshaft angle is used for all engine speeds. The load caused by inertia increases in magnitude as the engine speed increases. Therefore, as the engine speed increases, a larger magnitude of inertia force is deducted from the combustion load, resulting in a decrease of the total load.

Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Reserach Journal
Vol.2,Issue.IV/May; 2012

Figure 5: Comparison of maximum and range of bending and torsional loads at different engine speed FE MODELING OF THE CRANKSHAFT The FE model of the crankshaft geometry has about 105 quadratic tetrahedral elements, with the global element length of 5.08 mm and local element length of 0.762 mm at the fillets where the stresses are higher due to stress concentrations. As a crankshaft is designed for very long life, stresses must be in the linear elastic range of the material. Therefore, all carried analysis are based on the linear properties of the crankshaft material. There are two ways to find the stresses in dynamic loading. One method is running the FE model as many times as possible with the direction and magnitude of the dynamic force. An alternative and simpler way of obtaining stress components is superposition of static loading. The main idea of superposition is finding the basic loading positions, then applying unit load on each position according to dynamic loading of the crankshaft, and scaling and combining the stresses from each unit load. In this study both methods were used with 13 points over 720 degrees of crankshaft angle. The results from 6 different locations on the crankshaft showed identical stress components from the two methods.

Figure 6: FEA model of the crankshaft with fine mesh in fillet areas It should be noted that the analysis is based on dynamic loading, though each finite element analysis step is done in static equilibrium. The main advantage of this kind of analysis is more accurate estimation of the maximum and minimum loads. Any loading condition during the service life of the crankshaft can be obtained by scaling and combining the magnitude and direction of these two loads. Boundary conditions in the FE model were based on the engine configuration. The mounting of this specific crankshaft is on two different bearings which results in different constraints in the boundary conditions. One side of the crankshaft is fixed to the engine block by a ball bearing and the other side is rolling over a journal bearing. When under load, only 180 degrees of the bearing surfaces facing the load direction constraint the motion of the crankshaft. Therefore, a fixed semicircular surface as wide as the ball bearing width was used to model
Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Reserach Journal
Vol.2,Issue.IV/May; 2012

that section. This indicates that the surface can not move in either direction and can not rotate. The other side was modeled as a fixed thin semicircular ring which only holds the crankshaft centerline in its original position and acts as a pivot joint. In other words, the journal bearing is modeled in a way that allows the crankshaft to rotate about axis 1 as well as slide in direction 3 as occurs in a journal bearing. These defined boundary conditions are shown in Figure 7. Boundary conditions rotate with the direction of the load applied.

RESULTS AND DISCUSSION OF STRES ANALYSIS Some locations on the geometry were considered for depicting the stress history. These locations were selected according to the results of FE analysis, and as expected, all the selected elements are located on different parts of the fillet areas due to the high stress concentrations at these locations. Selected locations are labeled in Figure 8 and the von Mises stresses with sign for these elements are plotted in Figure 9. The critical loading situation is at the crank angle of 355 where the combustion exerts a large impact on the piston. At this time all stresses are at their highest level during stress time history in a cycle. As can be seen, location number 2 experiences the highest stress at this moment. Therefore, element number 2 was selected as the criticalelement. Figure 10 shows the maximum stress, meanstress, and stress range at the engine speed of 2000rpm at different locations. It can be seen that elementnumber 2 not only has the maximum von Mises stress,but it also carries the largest stress range and meanstress among other locations. This is important in fatigueanalysis since the range and mean stress have moreinfluence than the maximum stress. This is anotherreason for why having the stress history of criticalelements are more useful than static analysis of thecrankshaft

Figure 8: Locations on the crankshaft where the stress variation was traced over one complete cycle of the engine, and locations where strain gages were mounted

.
Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Reserach Journal
Vol.2,Issue.IV/May; 2012

Figure 9: von Mises stress history (considering sign of principal stress) at different locations at the engine speed of 2000 rpm

Figure 10: Comparison of maximum, minimum, mean, and range of stress at the engine speed of 2000 rpm at different locations on the crankshaft Figure 11 shows the effect of engine speed on minimum, maximum, mean and range of stress. This figure indicates the higher the engine speed, thelower the von Mises stress. It should, however, be noted that there are many other factors regarding service life of an engine. Other important factors when the engine speed increases are wear and lubrication. As these issues were not of concern in this study, further discussion is avoided.

Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Reserach Journal
Vol.2,Issue.IV/May; 2012

Figure 11: Variation of minimum stress, maximum stress, mean stress, and stress range at location 2 on the crankshaft as a function of engine speed The effect of torsional load was discussed in the load analysis section, and was pointed out that it has no effect on the stress range of the critical location. The von Mises stress at location number 2 shown in Figure 9remains the same with and without considering torsional load. This is due to the location of the critical point which is not influenced by torsion since it is located on the crankpin bearing. Other locations such as 1, 6, and 7 in Figure 8 experience the torsional load.

Figure 12 shows changes in minimum, maximum, mean, and range of von Mises stress at location 7 with considering torsion and without considering it during service life at two different engine speeds. It can be seen that the minimum von Mises stress does not change since the minimum happens at a time when the torsional load is zero. The effect of torsion is about 16 percent increase in the stress range at this location.Figure 12: Effect of considering torsion in stresses at location 7 at different engine speedsStress results obtained from the FE model were verified by experimental component test. Strain gages were mounted at four locations on the crankpin bearing. These locations are labeled as a, b, c, and d in Figure 8FEA results are also shown and compared with experimental results in this table. As can be seen, differences between FEA and strain gage results are less than 7 percent for different loadingconditions. This is an indication of the accuracy of theFE model used in this study. Table 2: Comparison of stress results from FEA and strain gages located at positions shown in Figure 8

Comparison of stresses at locations c and d resulting from loading in direction 1 in Figure 7 show symmetric stress values from FEA, experiment, and analytical method. The results from these three methods are close to each other. However, stresses obtained form FEA results and experiment show different stresses (i.e. non-symmetric) at locations a and b, resulting from loading in direction 2 in Figure

Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress (8), ISRJ Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)

“Dynamic Load and Stress Analysis of a Crankshaft”

Indian Streams Reserach Journal
Vol.2,Issue.IV/May; 2012

CONCLUSIONS The following conclusions could be drawn from this study: 1. Dynamic loading analysis of the crankshaft results in more realistic stresses whereas static analysis provides an overestimate results. Accurate stresses are critical input to fatigue analysis and optimization of the crankshaft. 2. There are two different load sources in an engine; inertia and combustion. These two load source cause both bending and torsional load on the crankshaft. 3. The maximum load occurs at the crank angle of 355 degrees for this specific engine. At this angle only bending load is applied to the crankshaft. 4. Considering torsional load in the overall dynamic loading conditions has no effect on von Mises stress at the critically stressed location. The effect of torsion on the stress range is also relatively small at other locations undergoing torsional load. Therefore, the crankshaft analysis could be simplified to applying only bending load. 5. Critical locations on the crankshaft geometry are all located on the fillet areas because of high stress gradients in these locations which result in high stress concentration factors. REFERENCES 1. Henry, J., Topolsky, J., and Abramczuk, M., 1992, “Crankshaft Durability Prediction – A New 3-D Approach,” SAE Technical Paper No. 920087, Society of Automotive Engineers 2. Guagliano, M., Terranova, A., and Vergani, L., 1993, “Theoretical and Experimental Study of the Stress Concentration Factor in Diesel Engine Crankshafts,” Journal of Mechanical Design, Vol. 115, pp. 47-52 3. Payar, E., Kainz, A., and Fiedler, G. A., 1995, “Fatigue Analysis of Crankshafts Using Nonlinear Transient Simulation Techniques,” SAE Technical Paper No. 950709, Society of Automotive Engineers 4. Prakash, V., Aprameyan, K., and Shrinivasa, U., 1998, “An FEM Based Approach to Crankshaft Dynamics and Life Estimation,” SAE Technical Paper No. 980565, Society of Automotive Engineers 5. Borges, A. C. C., Oliveira, L. C., and Neto, P. S., 2002, “Stress Distribution in a Crankshaft Crank Using a Geometrucally Restricted Finite Element Model”, SAE Technical Paper No. 2002-01-2183, Society of Automotive Engineers 6. Shenoy, P. S. and Fatemi, A., 2006, “Dynamic analysis of loads and stresses in connecting rods,” IMechE, Journal of Mechanical Engineering Science, Vol. 220, No. 5, pp. 615-624 7. Shenoy, P. S. and Fatemi, A., "Connecting Rod Optimization for Weight and Cost Reduction", SAE Paper No. 2005-01-0987, SAE 2005 Transactions: Journal of Materials and Manufacturing 8. Zoroufi, M. and Fatemi, A., "A Literature Review on Durability Evaluation of Crankshafts Including Comparisons of Competing Manufacturing Processes and Cost Analysis", 26th Forging Industry Technical Conference, Chicago, IL, November 2005

Please cite this Article as : Prof. N. G. Alvi , Dr. S. V. Deshmukh and Ram.R.Wayzode , “Dynamic Load and Stress Analysis of a Crankshaft” : Indian Streams Research Journal (MAY ; 2012)…...

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