Hyundai Zeta Increased Power Connecting Rod Analysis November 15, 2006 G. Renzi

1. Hyundai Zeta Increased Power Connecting Rod Analysis November 15, 2006 G. Renzi

2. Discussion Topics • Engine Specifications • Material Property • Connecting Rod Design Specifications • FEA Assumptions and Loading Conditions • FEA Results • Discussion of Results

3. Zeta Engine Specifications/Assumptions

4. MaterialProperty PF Material: HS150

5. Zeta Connecting Rod Specifications

6. Zeta Connecting Rod Forge Design FORGE MODEL TOTAL WT = 602.9 grams CRANK WT = 456.8 grams PIN WT = 146.1 grams

7. Zeta Connecting Rod Machined Design ASSEMBLY MODEL TOTAL WT = 565.0 grams CRANK WT = 422.0 grams PIN WT = 143.0 grams MACHINED MODEL (w/o bolts & bushing) TOTAL WT = 494.8 grams CRANK WT = 360.1 grams PIN WT = 134.7 grams

8. Finite Element Analysis Model total nodes = 98,112 total elements = 60,895

9. Finite Element Analysis • The components included in the finite element model for a 1/4 of the connrod assembly are the rod, cap, bushing, bolt, wrist pin, crankshaft and crankshaft bearing shells. (All of these components are modeled as deformable structures.) The rod, cap, bushing and bolt meshes are constructed exclusively with 10-noded tetrahedral elements. The meshes for the remaining components consist of either 10-noded tetra, 15-noded prisms, or 20-noded hex elements. • 3-D, surface-to-surface contact elements exist at the following component interfaces: rod to cap wrist pin to bushing bushing to rod bolt to rod rod to upper bearing shell bolt to cap crank to upper bearing shell cap to lower bearing shell crank to lower bearing shell lower to upper bearing shells • All of the material constitutive properties specified for the connecting rod analyses are linear elastic.

10. Finite Element Analysis • All contact elements, except those between cap and rod and rod and bolt, exhibit standard surface-to-surface interaction behavior and function with a coefficient of friction equal to 0.1. Contact surface behavior between cap and rod is specified as “rough” and between rod and bolt is designated as “bonded”. The contact elements at both of these locations have infinite frictional resistance. • The entire loading history of the connecting rod is divided into multiple, sequential steps. The initial step is the pressfit of the wrist pin bushing in the rod’s small end bore. The next step is the assembly loading of the crank bearing. The requisite axial force in the fastener for bearing “crush” is established by an auto-pretensioning function built into the nonlinear solution algorithm. Lastly, the tensile and compressive forces for the max torque and overspeed load cases are applied as uniform pressures on the surface of the wrist pin.

11. Finite Element Analysis • Pressure loads on the wrist pin start and end at axial distances from the upper thrust surface of 1.50 mm and 10.36 mm, respectively (pressure applied on the red, wrist pin elements shown in the page 7 model plot). • Boundary conditions are applied to nodes on the crankshaft’s planar face. These restraints ground the rod assembly. In addition, b.c.’s are applied to nodes on the two planes of reflective symmetry. The nodes at each symmetric surface are assigned uni-directional b.c.’s that prevent their motion in a direction normal to the plane on which they lie. • The modified Goodman approach is used to evaluate the Signed von Mises Fatigue Safety Factors at different areas of concern. The modified Goodman procedure for this analysis uses, in part, two equations which are conditionally employed depending on whether the mean stress is positive or negative.

12. Finite Element Analysis The Fatigue Safety Factor equation used for positive mean stress is: FSF = 1 / {(smean / sult) + (samp / send)} The Fatigue Safety Factor equation used for negative mean stress is: FSF = (send / samp) where, samp = Amplitude Stress = (smax - smin) / 2 smean = Mean Stress = (smax + smin) / 2

13. Finite Element Analysis Load Calculations Based on Reciprocating and Rotational Mass at TDC and BDC

14. Fatigue Safety Factor Locations 1 12 2 3 4 5 6 8 10 7 9 11

15. Finite Element Analysis Results von Mises Stress Contours due to Bushing Pressfit Max Stress = 435.2 MPa

16. Finite Element Analysis Results Fatigue Factor Contours for Pin End Max Torque Overspeed FSF = 1.30 FSF = 1.64 FSF = 2.25 FSF = 1.48 FSF = 1.95 FSF = 1.14

17. Finite Element Analysis Results Fatigue Factor Contours for Shank Region Max Torque Overspeed FSF = 1.81 FSF = 3.13 FSF = 1.66 FSF = 2.69 FSF = 1.72 FSF = 2.64 FSF = 2.60 FSF = 1.58

18. Finite Element Analysis Results Fatigue Factor Contours for Crank End Max Torque Overspeed FSF = 1.59 FSF = 2.26 FSF = 2.5 FSF = 1.61

19. Finite Element Analysis Results Fatigue Factor Contours for Crank End at Bolt Spot Face Max Torque Overspeed FSF = 1.89 FSF = 2.37

20. Finite Element Analysis Results Failure Mode Safety Factor Table

21. Finite Element Analysis Results Joint Face Contact Pressure Bolt Pretension / Brg Assy Load Overspeed Pull Load 1 7 1 7 6 6 2 2 3 3 4 4 5 5

22. Finite Element Analysis Results Contact Pressure for Overspeed Pull Load (0 to 10 MPa Range)

23. I-Beam Yield Safety Factor X Y Center of Cylinder Bore Offset Load Condition Material: HS150 Note: Maximum Peak Cylinder Pressure of 85.9 bar at 5,000 RPM was considered. S.F. is acceptable. Centered Load Condition

24. Finite Element Analysis Results Axial Normal Stress Contours On Bolt Cross-Section Bolt Pretension / Brg Assy Load Overspeed Pull Load Stresses and Axial Force in Bolt

25. Finite Element Analysis Results Crank Bore Distortion

26. Discussion of Results • Two calculated Fatigue Safety Factors are less than Metaldyne’s allowable limit for gas engine connrods. The minimum FSF of 1.14 (1.02 -3) occurs between the 5 and 6 o’clock positions on the inner surface of the wrist pin collar. This minimum value of FSF develops when the rod undergoes max. torque loading. However, this low FSF result is an underestimate because of the finite element model’s wrist pin constraint and load arrangements. Specifically, the pin is under-constrained, and is loaded on its OD by pressures that are uniform (in both axial and circumferential directions). The location with the next lowest FSF exists within the pin end oil hole. Here the value of the Fatigue Factor when the connrod is subjected to max. inertia (i.e., overspeed) loads is 1.30 (1.18 -3). • A survey of the FEA-based Yield Safety Factors reveals that all are above Metaldyne’s allowable limit for gas engine connrods. • The contact pressures that develop at the rod-to-cap interface range in value from 0 to 320 MPa during application of the upward pull of the max. inertia loading condition (the most severe). The pressure distribution for this case (and therefore all others) is acceptable because only a small zone of zero contact pressure exists on the joint face. Contact pressure readings of zero MPa, which signify the onset of rod/cap separation, only occur in the one joint surface corner formed by the thrust face and crank bore edges.

27. Discussion of Results • The bushing pressfit produces a maximum von Mises stress in the part of 435 MPa. This max stress occurs at the 12 o’clock position on the rod’s small end bore, along the edge defined by the intersection of the cylindrical bore surface and the oil hole. This peak stress is below the material allowable. • The largest of the crank bore diametrical distortions develops for the tensile component of the max. inertia loading condition. This upward pull produces a vertical expansion and horizontal close-in of the crank bore equal to 78.8 m and -55.0 m, respectively. Both the vertical expansion and the horizontal close-in of the crank bore fall within the Metaldyne displacement limits. • Gross, shank-section, bolt stresses peak at 803 MPa while resisting the tensile force applied during the max. inertia loading condition. This peak stress does not exceed the material allowable. • The calculated 1.19 value for the minimum I-beam Yield Safety Factor is greater than Metaldyne’s recommended lower limit and is therefore acceptable.