engdyn noise

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Noise and Vibration Optimisation of aGasoline Engine

L. ChabotPSA Peugeot

Citroen

Oliver G.K.YatesRicardo Group

The vibratory and acoustic behaviour of the internal combustion engine is a highly complex one, consisting omany components that are subject to loads that vary greatly in magnitude and which operate at a wide rangof speeds. The interaction of these components and the excitation of resonant modes of vibration is a majoproblem for the powertrain engineer when optimising the noise and vibration characteristics of the powertrainFinite element analysis (FEA) and dynamic simulation programs such as ENGDYN have been developewhich provide the engine with simulation tools that enable what if comparative studies to be undertaken

Such tools have been extensively used by Ricardo and PSA over a number of years using the Ricardsoftware products described as part of this paper. This paper summarises a study that has been undertaketo assess and optimise the dynamic behaviour of a current production 1.6 litre gasoline engine with thobjective of reducing low frequency radiated noise from the cylinder block.

PEUGEOT

CITRON


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INTRODUCTION

The TU5JP engine is currently used for a numberof vehicle platforms produced by PSA, includingthe successful "Super mini" class car, the CitroenSaxo. The objective of the project was toinvestigate design modfications that would reducelow frequency radiated noise and vibration of thepowertrain, whilst being cost-effective toimplement. This would enable the engine toachieve a longer production life cycle and providefuture vehicle platforms with a more refinedpowertrain option.

The objective was to reduce the radiated noise ofthe TU5JP engine in the frequency range up to1kHz . The running engine speed range of interestwas from 2000 rev/min to 6500 rev/min at full load.The details of the engine are summarised below.

Type GasolineConfiguration 4 cylinder, in-lineBore 78.5 mmStroke 82.0 mmSwept Volume 1.6 litreFiring Order 1-3-4-2

TECHNICAL APPROACH

The technical approach detailed here isrepresentative of Ricardos standard methodsapplied to the analysis of the powertrain tooptimise its noise and vibration characteristics.The analysis approach uses a combination ofindustry standard proprietary software products(finite element mesh generation packages suchas IDEAS and FEMGEN and solvers such as

NASTRAN)and Ricardos own software products(ENGDYN and FEARCE).

A brief description of the Ricardo Softwareproducts used is given in Appendix A.

Ricardo used FEARCE for pre-processing toassemble the powertrain, with ENGDYN being

used for dynamic load prediction and forcedresponse and acoustic analysis.

a) Finite Element Model

The finite element models of the powertrainassembly were provided by PSA in IDEASuniversal file format. The block, head andtransmission were constructed of predominantlylow-order hexahedral elements, whilst theancillaries were modelled using a combination of

high order tetrahedral, shell and beam elementswherever appropriate. Lumped mass elementswere used to model components whose stiffnessdid not contribute to the local or global response inthe locations of interest. The completed finiteelement model of the powertrain is shown inFigure 1.

Figure 1 Finite Element Model

The FE models were translated into the Ricardostandard file format and were assembled and

joined together to form the complete powertrainusing FEARCE. For low frequency problems it is


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essential that the total modeled mass of thepwertrain and the distribution of that mass is

identical to the actual mass to ensure that the rigidbody modes and fundamental flexible modes ofthe powertrain are modeled correctly. A detailedmass breakdown of the modeled powertrainassembly was compared to supplied data for theactual components. Any subsequent discrepancyin mass was corrected by modifying the density ofthe relevant component or alternatively lumpedmass elements were added as appropriate

b) Modal Analysis

A free-free modal analysis was performed usingNASTRAN of the complete powertrain assembly.The resonant frequencies, mode shapes andstrain energy distributions were obtained forfrequencies up to 2.5 kHz.

c) Dynamic Load Calculation

The excitation loads acting on the engine werecalculated using ENGDYN. ENGDYN is used to

predict the time-domain response of the 3-dimensional vibration of the coupled crank trainand cylinder block system with non-linear oilfilmsat each of the main journal bearings. The massand stiffness matrices used for the dynamic modelof the crankshaft are derived by staticcondensation of the finite element model. Thesetechniques were first used by Hodgetts1Ref.[1,2,3,4] in the development of 3-dimensionalcrankshaft vibration methods, and subsequentlyapplied to various types of engine by RicardoRef.[5,6].

The mass and stiffness matrix reduction of thecrankshaft is performed internally within ENGDYN

1

Numbers in parentheses indicate references at theend of the article

by undertaking a series of FE analyses on thindividual crank webs to determine the we

stiffnesses and masses.

The reduction of the cylinder block finite elemenmodel to derive a reduced model for ENGDYwas performed using NASTRAN. The degrees freedom selected comprised all the criticafreedoms which represent the application oloading to the model. This included the bearingpiston top and bottom reversals and cylinder headFigure 2 shows the finite element models of thcrank train and cylinder block together with threduced models displayed in ENGDYN.

Figure 2 Display of Crankshaft and CylinderBlock Models in ENGDYN

The cylinder pressure data loading was provideby PSA from measurement data at critical enginspeeds. This was interpolated within ENGDYN tprovide cylinder pressure data for the analysis aintermediate engine speeds.

The oil film bearing model applied in ENGDYN fodynamic analysis is based on the non-lineamodel developed by Booker Ref.[6] and KikuchRef.[7]. For radial motion of the shaft in thbearing the oil film model is based on the mobilit


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method of Booker, using the short bearingapproximation (Ocvirk Solution). For tilt motion

(shaft axis tilting relative to bearing axis), theshort bearing approximation is also used todetermine the stiffness and damping of the oilfilm.

PSA supplied lubrication data including oil typeand oil supply temperatures to the main bearings.ENGDYN includes a thermal solution option thatpredicts the effect of the temperature rise in theoil at each main bearing due to the bearing actionThis option was used to derive the effective oiltemperature and therefore viscosity at each main

bearing.

Dynamic simulations were performed at enginespeeds from 2000 to 6500 rev/min at 100 rev/minintervals. The ENGDYN graphical user interfacewas used to plot and animate the dynamic resultsfor both the crankshaft and cylinder block.

d) Forced Response and Acoustic Analysis

The excitation loads calculated by ENGDYN are

used as input by the same program to perform amodal frequency response analysis to predict thevibratory response of the complete powertrain.

Radiated noise is calculated by ENGDYN fromsurface normal velocities using the Rayleighequation. Radiated sound power and radiationefficiency are calculated for each vibrating surfacewhich are approximated as a flat plate. Soundintensity is also calculated across the surface.

The acoustic wave equation:

2

2

2

2 ),(1),(t

tpc

tp = [1]

defines the velocity potential ),( tp which is

related to the time dependent particle velocity

),( tpu by:

),(),( tptpu = [2]

and c is the propagation velocity (p and t are thespatial & time variables.)

The Rayleigh integral equation Ref.[9] relates thevelocity potential at a point p to the normalvelocity of the plate:

dSqvr

ep

ikr

)(2

1)(

= [3]

The Rayleigh equation is valid for a flat plate in an

infinite baffle with the acoustic medium extendingin the positive normal direction. The point p canbe either in the acoustic medium, on the plate oron the baffle. The points q are points on thevibrating surface which when integrated representthe whole surface.

Two solutions to the Rayleigh equation areimplemented. The first uses a simple methodtreating each node as a piston in a baffleassuming that the vibration is the same over thesurface of the piston. This is normally acceptable

for models where the mesh density is sufficient togive a good representation of the mode shapesand the wavelength is large compared to theelement size. The second method solves thesame Rayleigh equation but to a greater accuracy(and therefore takes longer). The face setdefining the radiating surface is triangulated andthe pressure waves are integrated across eachtriangle with the integration order automaticallyadjusted according to the size of the element, thedistance from the element and the wavelength ofsound at the forcing frequency. Using its

minimum integration order this more rigorousmethod takes approximately three times longer tosolve than the simpler method.

The radiated sound power from a single vibratingsurface or a number of surfaces are presented


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using ENGDYN in the form of simple xy-plots oftotal sound power against engine speed and as

colour contour plots often referred to as CampbellDiagrams. In order to compare the CampbellDiagrams of two designs a colour contour plotshowing the difference in dB can also bepresented using ENGDYN. This plot enables thedifference in response between two designs to beeasily assessed.

BASELINE ASSESSMENT

Figure 3 shows the predicted sound power spectra

from both sides of the cylinder block for theexisting engine.

Figure 3 Sound Power Spectra for Base Engine

This figure shows radiated sound power levels to

kHz for engine speeds from 2000 to 6500 rev/mat full load. The significant frequencies arhighlighted, with the corresponding modes antheir descriptions summarised in Table 1. Thesmodes were identified by animating the vibratorresponse of the powertrain at the points indicateon the figure and comparing these operatinshapes with the previously calculated modshapes. These results were consistent with PSexperience of the engine based on experimentadata.

ModeNo.

Freq.(Hz)

Description

8 254 Driveshaft and powertrain out-of-planebending

9 260 First powertrain in-plane bendingmode

10 291 Compressor11 305 First powertrain out-of-plane bending

15 397 Driveshaft16 404 Powertrain torsion and starter motor

17 480 In and out of plane bending and sump24 606 Main Bearing 425 616 Block torsion and driveshaft bearing

bracket28 686 Axial mode of assembly and Main

Bearing 330 740 Main Bearing 3

31/36 763/ 813

Driveshaft bearing bracket and sump

Table 1 Significant Powertrain Modes

DESIGN OPTIMISATION STUDY

A number of design modifications were considereand were analysed under identical conditions tthat of the baseline in order to provide back-toback comparisons of the results. Two of thesmodifications are presented here.


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Modification 1 - Revised Crankcase Wall

Geometry

The lower crankcase was modified, as shownschematically in red in Figure 4, by curving the wallalong its length.

Figure 4 Revised Lower Crankcase Wall

Figure 5 Total Sound Power for Baseline andModification 1

Figure 5 shows the total radiated sound powerfrom the cylinder block whilst Figure 6 shows theCampbell diagrams for both the modified and

baseline designs and a difference diagramcomparing the two. These figures indicate

negligible improvement due to this modification.

Figure 6 Comparison of Baseline andModification 1

Modification 2 Additional Internal and ExternalRibs

Additional ribs were added internally to stiffen themain bearing bulkheads as shown in Figure

7together with additional ribs added to the outersurfaces of the cylinder block. Figure 8 shows theribs added on the driveshaft side of the engine inan attempt to stiffen the lower crankcase wall andto provide better support for the driveshaft bearingbracket.


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Figure 7 Internal Ribs at Main Bearings

Figure 8 External Ribs on Driveshaft Side ofEngine

Location MassIncrease

[gms]

Main bearing bulkheads 193

Rib running length of cylinder block ondriveshaft side

95

Attachment point of driveshaft bearingbracket

84

Attachment point of compressor bracket 104

Attachment point of alternator bracket 336

Table 2 Mass Increase due to Additional Ribs

The increase in mass due these changes was 81

gms. A breakdown of this total is summarised Table 2.

Figure 9 compares the density of modes to 1 kHbetween the baseline and the modified designThis shows that the natural frequencies o modebetween 540 and 980 Hz of the modified desighave increased when compared with the baseline

Figure 9 Modal Density Comparison

Figure 10 compares the total radiated sounpower at full load up to 1 kHz from the cylindeblock against engine speed for the two designsThis shows reductions of up to 2dB in total sounpower particularly between 3000 and 450rev/min. Inspection of the Campbell diagrams andifference plot of shows that between thesspeeds the noise levels between 600 and 800 Hhave been reduced.


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Figure 10 Total Sound Power for Baseline andModification 2

Figure 11 Comparison of Baseline andModification 2

An additional analysis was performed in which only

the internal ribs were added to the base engine.This demonstrated that these improvements weredue to these ribs and negligible benefit wasderived from the external ribbing. This result isshown in Figure 12.

Figure 12 Effect of Removing External Ribs

It was therefore recommended that only theinternal ribs be incorporated into the cylinder block

casting. The increase in mass due to thismodification was 193 gms, an increase of only0.5% to the existing cylinder block.

CONCLUSIONS

The dynamic behaviour of the TU5JP engine hasbeen assessed using proven simulation tools.Radiated noise from the cylinder block has beenpredicted up to 1 kHz using the Rayleigh methodand the critical modes of vibration of thepowertrain have been identified. The results of thebaseline assessment were consistent with PSAexperience of the engine based on experimentaldata.


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A number of production feasible designmodifications have been considered and

analysed. It has been demonstrated that bystiffening the main bearing bulkheads usingadditional ribs at the crankshaft centre-linereductions in radiated noise from the cylinderblock of up to 2dB can be realised with a minimalincrease in overall mass.

REFERENCES

1. I.Mech.E C99/71, 1971 Vibrations of aCrankshaft. D.Hodgetts, Advanced School ofAutomobile Engineering, Cranfield.

2. PhD Thesis (University of London) TheVibrations of Crankshafts. D.Hodgetts, CranfieldInstitute of Technology.

3. I.Mech.E C216/76,1976 The Whirl Modes ofa Crankshaft, D.Hodgetts, Cranfield Institute ofTechnology.

4. FISITA,1986 The Dynamic Response ofCrankshafts and Camshafts, D.Hodgetts,Cranfield Institute of Technology.

5. I.Mech.E AD Autotech Congress, 1987 TheMeasurement and Prediction of Flywheel Whirl,A.R.Heath, Ricardo.

6. I.Mech.E C14/87,1987 Computers in AnalysisTechniques for Reciprocating Engine Design,D.J.Lacy, Ricardo.

7. Transactions of the ASME, journal of BasicEngineering, Sept. 1965Dynamically LoadedJournal Bearings - Mobility Method Of Solution.J.F.Booker , Cornell University

8. Bulletin of the JSME, 1970 Analysis ofUnbalance Vibration of Rotating Shaft Systemwith Many Bearings and Disks. K.Kikuchi.

9. Acoustics: An Introduction to its PhysicalPrinciples and Applications, A.D.Pierce, McGraw-Hill 1981.

ACKNOWLEDGEMENTS

Ricardo would like to thank PSA Peugeot Citroefor the permission to publish this paper in the FiftRicardo Software International User Conference Detroit.

APPENDIX A

ENGDYN, Engine Dynamics Simulation Program

ENGDYN is a computer program for analysin

the dynamics of the engine and in particular thdynamics of the crank train and its interaction witthe cylinder block. ENGDYN provides a numbeof different solution techniques for predictinengine dynamics using models of varyindegrees of sophistication. The crank train ancylinder block models can either be defined arigid, compliant and dynamic. In its simplest formENGDYN can be used to perform a staticallydeterminate solution, whilst in its mossophisticated form it can be used to predict thtime-domain response of the 3-dimensiona

vibration of the coupled crank train and cylindeblock system with non-linear oilfilms at each othe main journal bearings. This flexibility enablethe user to generate an engine model and tperform a solution to meet his particular needs.

FEARCE, Finite Element Analysis RicardConsulting Engineers

FEARCE is a suite of programs and translatorwhich form a finite element analysis environmendedicated to powertrain analysis. The majo

function of FEARCE is to provide very advancepreprocessing, postprocessing and solvefunctionalit. It interfaces with most commercianon-proprietary products such as NASTRANIDEAS etc., as well as Ricardo Software productprimarily ENGDYN.


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