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7/30/2019 Klarin Final

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2005 ABAQUS Users Conference 1

Acoustic Analysis of Power Units with an ABAQUSIntegrated Application Workflow

B. Klarin, B. Loibnegger, and Th. Resch

AVL List GmbH, Hans-List-Platz 1, 8020 Graz, Austria

Abstract: The design of low noise engines is still a challenging task due to contrary demands likeincreasing use of light weight materials and tighter limits of admissible stress and strain. To

shorten the entire design process with pre-optimized prototypes the simulation of engine and power unit vibration and structure-borne noise (surface velocity levels) becomes more and more a standard application. Nowadays the target is to better integrate the necessary combined simulation methodologies to reduce the overall analysis time (time of entire workflow). AVL EXCITE has been developed especially for acoustic application using an outstanding simulationtechnology, that enables results to be calculated very close to the absolute ones. Based on a

strategic co-operation with the company ABAQUS, AVL has implemented a seamless acousticanalysis, which is outlined in this paper. The complete workflow is demonstrated on an inline 4cylinder gasoline engine and compared with the traditional MSC.Nastran-oriented approach. Thisnew approach enables enhanced reliability due to automated data transfers, shorter project turn-around times and reduced costs in reaching the targets.

Keywords: Dynamics, Multi-Body Dynamics, Powertrain, Substructures, Vibration, DesignOptimization

1. Introduction

Virtual design and prototyping in the development of new combustion engines and power units play a critical role in todays automotive industry. Significant reduction of development time andcost can be achieved by high quality simulation results. At the same time, increasing power andspeed of the vehicles, combined with demands for light and compact design, and resultingcomplex geometry of engine parts, require detailed and time consuming calculations in the processof engine design.

To meet these challenging requirements, many sophisticated simulation tools have beendeveloped, which are applicable in both concept and detailed analysis phases of engine

development process.Results of numerical simulation of power units, as complex as they may be today, will no longer

be satisfactory unless they allow conclusions to be drawn with respect to the specific stages of the


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2 2005 ABAQUS Users Conference

development process. The requirements are to predict stress, durability, vibrations and specificvibro-acoustic phenomena. This is particularly true in regard to the cranktrain as the central part of the engine.

Due to increased efforts for detailed modeling of elastic multi-body systems and relevant non-linear body contacts, simulation models have become very complex and a high amount of computation time is necessary to calculate accurate results. Thus, another challenge is to deliver the results in the short time required for a fully integrated simulation solution in the differentstages of the development process.

Modern methods for simulation of engine dynamics consider the global movements, the coupledtorsional and bending vibrations of the crank train parts and the hydrodynamic influence of theslider bearings under running engine conditions. Specific results are normally produced in bothtime and frequency domains, including the detection of possible resonance (e.g. of a flywheel) andthe prediction of the strength of connected parts.

The solution procedure commonly used in the crankshaft dynamic analysis is based on acombination of the multi-body dynamics and the Finite Element Method (FEM). AVL haveintroduced the software EXCITE for this purpose, [1].

Due to the prediction requirements in the low frequency (mount vibrations) and the high frequencyranges (noise transfer), various detailed models are developed. Hence, requirement for highquality results on one hand and demands for less pre- and post-processing and calculation time onthe other generate conflicting demands for the engineers.

Therefore, many efforts have been made to automate the modeling process, e.g. for the complexdesign of the crankshaft. Furthermore, the contact models have been developed for slidingcontacts in order to obtain better results with reduced computing time [3, 8]. In addition, aneconomical bearing model has been developed, which is able to capture the physical behavior of

journal bearings with less computing effort.

2. General approaches

2.1 Multi-body system

In EXCITE, a multi-body system (MBS) approach has been implemented, which can be used tosimulate the dynamic behavior of crankshafts and engines. The total mechanical system of acranktrain consists of various parts (bodies) such as crankshaft and engine cylinder block, whichare coupled by non-linear contacts (joints) such as journal bearings. The mass/moment of inertiaand elasticity of these bodies are considered with the help of Finite Element Method (FEM).

In EXCITE (henceforth referred to as the engine simulation tool), two types of body motionhave been considered, namely the global motion (e.g., crankshaft rotation, connecting rod

movement) and local vibration motion.The vibration motion is governed by Newtons equation. For a linear elastic system, this can bewritten as,


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ext f pq K q Dq M +=++ *&&& (1)

where, q is the generalized displacement vector, and M , D , and K are mass, damping and stiffnessmatrices of the MBS, p* is the inertia forces due to large global motion of bodies, and the f ext isthe excitation forces. the f ext consists of two parts: (1) the external forces/moments such as gasforces, (2) non-linear constraint forces in joints such as hydrodynamic forces/moments in journal

bearings.

Details of the mathematical models for global motion are omitted here, but can be found in [2-3].

2.2 Hydrodynamic model of journal bearings with misalignment

In order to efficiently and correctly model the non-linear constraint forces and bending momentsin the oil film of journal bearings, an expedient hydrodynamic model has been developed toconsider the effect of journal misalignment in a rigorous way.

Figure 1 depicts a simplified geometry of a bearing with a misaligned journal. For a constant oilviscosity, the oil film pressure at a cross section of arbitrary axial coordinate ( z ) is determined bythe following Reynolds equation,

( ) ( )

( )[ ]'sin''cos'12

'cos'1'

'cos'1'

1

'2

332

avg C

z p

z p

R

+=

+

+

+

&&(2)

The boundary conditions are,

0)2/,'( = L p (3)

Based on the short bearing approximation [7] and the geometrical relationship shown in Figure 1,the analytical solution of Equation 3 has been obtained for oil film pressure,

+

=

+

= z

C LG

G z C L

z C G

G L

z C

p 102

10

22

2 341

234

2 (4)

where,

( ) 30 cossin H G avg && +=

&& 43211 H H H H G +++=

( ) cos1 += H

H 1, H 2, H 3 and H 4 are the derived coefficients attributable to journal misalignment.


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Figure 1. Bearing geometry with a misaligned journal

The enhanced HD model has been implemented as a so-called ENHD joint.. The implementedmodule is able to cope with both node to node coupling and node to surface connection betweentwo bodies. It can be used to model all types of journal bearings such as connecting rod big endand main bearings. It should be noted that in hydrodynamically lubricated journal bearings, the oilfilm thickness can drop to a level comparable to surface roughness. When this happens, there will

be asperity to asperity contacts between the journal and bearing surfaces. This effect has to beconsidered in order to correctly model the constraint forces/moments in these bearings and indeedto enhance the convergence behavior of the hydrodynamic solution. Hence, in the ENHD modulea statistical asperity contact model developed by Greenwood and Tripp has also been implementedin conjunction with the HD model described above.

Thus, the total pressure at a local position in the bearing is calculated by superimposing thecontact pressure ( pa) onto the hydrodynamic pressure ( p).

3. Acoustic Analysis Workflow

The target of acoustic simulation is to investigate complex engine dynamic behavior in the wholespeed range under different loading conditions in the most effective way during the EngineDevelopment Process (EDP). AVL has developed a straightforward procedure to perform acousticanalysis by using this simulation tool.

The main bodies for this class of engine dynamic analyses are the power unit, crankshaft andconnecting rods. The AVL MBS approach works with condensed models, so prior to the MBSanalysis it is necessary to use a general FE solver like ABAQUS, MSC.Nastran or similar toobtain reduced stiffness and mass matrices. Presented in this paper are the workflow (Figure 2)and results from acoustic analysis done with reduced matrices obtained from ABAQUS andMSC.Nastran. Based on a strategic co-operation with the company ABAQUS, AVL has

implemented an ABAQUS linear FE-solver inside the engine simulation tool a module calledEXCITE FEA (referred to as integrated ABAQUS solution).


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The starting point for generating the power unit, crankshaft and connecting rods FE models isCAD data. Usually for the power unit, the FE model is prepared by a pre-processing tool and usedfor further acoustic analysis. Basically two types of crankshaft models are supported: a full threedimensional FE model as condensed matrices, and a structured model, which consists of concentrated mass elements coupled by fully populated stiffness matrices. With the aim to speedup dynamic analysis AVL has developed a special tool to model a crankshaft structured model for which preparing only the CAD geometry is sufficient. All meshing and definition of stiffnessmatrices as well as mass distribution is done automatically inside the engine simulation tool by themodules AutoShaft and Shaft Modeler.

Figure 2. Acoustic analysis workflow.

Natural frequency analysis is an important part of every NVH Analysis. It has to be performed for each FE-model as model check step and as basic investigation about the dynamic behavior.Eigenmodes, eigenvectors and strain energy for each mode should be calculated and evaluated indetail for the frequency range of interest. At the intersections of main engine excitation orders,depending on engine type and curves representing natural frequencies, possible resonance can be

predicted. Based on the integrated ABAQUS solution this step can be directly performed

internally without the need for external FE-Solver.

Prior to the MBS analysis it is necessary to perform a condensation of the FE models and obtainreduced mass and stiffness matrices to be used in further dynamic calculations. AVL recommends


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a combined static and dynamic condensation. The static condensation is used for all those master nodes that are required to connect the different piece parts (e.g. main bearing nodes) and for master nodes where external loading are assigned. The dynamic behavior of the remainingstructure is evaluated by the dynamic reduction, in which the vibration behavior is illustrated bythe vectors for the requested frequency range. All separate bodies, connecting rods, crankshaft andmodel of the power unit are reduced separately. Using the integrated ABAQUS solution, which issimilar to the previous natural frequency analysis, condensation of different bodies can be

performed directly inside the MBS tool. Connecting rod reduced models can also be automaticallygenerated using the integrated tool.

Figure 3. AVL engine dynamic model.

AVLs engine simulation tool is a multi-body system in which linear bodies are connected withnonlinear connections-joints (see Figure 3) and it is the main tool in the presented acousticanalysis. To perform acoustic analysis it is necessary to excite the structure with external forcesand moments. Those forces and moments have to be determined by pre-calculations or bymeasurements. The main excitations which should be considered in power unit acoustic analysisare combustion, timing drive and piston slap forces. AVL has developed special tools for simulating dynamic excitation within the MBS model. Using these interfaces it is possible toautomatically assign excitation forces to the dynamic model, to update, modify and manipulatethem, and to perform a co-simulation.

Once dynamic calculations are performed, results for the retained nodal DOFs included incondensed model and results for joints are directly accessible. If structure borne noise for thecomplete power unit structure is evaluated, an additional calculation step called data recovery is


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performed. Data recovery can be done either with an external FE solver or by using embeddedfunctionality inside the engine simulation tool. The embedded solution makes those calculationsmuch faster and overall time needed for acoustic analysis much shorter. Results from datarecovery can be used in further air borne noise calculations.

4. Results Comparison

Models were run and compared between the integrated ABAQUS solution (based on version 6.4),and the MSC.Nastran 2001 solution for FE-model condensation. The intention of this was toincrease confidence in the ABAQUS integrated solution and to understand what parameters could

be modified to increase performance. Results of those comparisons are presented in this chapter.

Note that the investigated FEM model is a relatively coarse model. This model can be investigatedrelatively fast and with low system resources and it is sufficient for comparing solver capabilities,as required for the present paper, but it would not meet actual requirements for detailed acoustic

analysis of an engine. For a better overview, results for a single engine speed under full loadcondition are presented. A dynamic and acoustic complete analysis of a power unit is performedfor the full speed range and different loading conditions as full and part load as well as motoredcondition.

4.1 CPU and Hardware Requirements

The most demanding calculation in the presented benchmarking for FE-solver is the dynamiccondensation of the power unit. The original MSC.Nastran FE-model input deck is converted toABAQUS format using ABAQUS translator. Differences in number of nodes, elements and DOFsare results of different representation of beam and mass elements. The number of retained nodalDOFs and retained eigenmodes was the same in both cases and all calculations are done on thesame workstation. A comparison of time and disk space needed for condensation showed better

performance when the integrated ABAQUS solution is used. MSC.Nastran has an advantage inview of memory needed to perform the condensation task. In the case when the number of DOFsin the FE-model is larger, memory required for the integrated ABAQUS solution can easilyexceed local resources.

Table 1. Power unit FE-model condensation.Model ABAQUS MSC.NastranBody Powerunit Powerunit tTotal Number of Nodes in Uncondensed FE-Model 37583 37044Total Number of Elements in Uncondensed FE-Model 30983 30742Total Number of DOFs in Uncondensed FE-Model 151598 143658Number of Retained Nodal DOFs in Condensed FE-Model 950 950Number of Retained Eigenmodes in Condensed FE-Model 260 260CPU Time 19 min 20 s 34 min 29 sDisk Usage 3950.6 MB 4153.3 MBMemory Usage 341.52 MB 190.04 MB


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It has to be mentioned that the lower memory usage by MSC.Nastran is true in general for allinvestigated models and element types. The reduced disk usage and analysis time for ABAQUS,as shown in this example, depends very much on the element types used within the model.

Figure 4 shows a relative comparison of disk usage and analysis time for condensation for different types of elements in ABAQUS. For large FEM models the usage of C3D8I or C3D10Melements should be avoided due to the large increase of system resources for these types of elements. Instead of C3D8I element type C3D8R with hourglass enhancement are recommended.

Figure 4. Influence of element types on CPU time, memory and disk usage.

4.2 Natural frequency analysis

The natural frequency analysis is the first pre-step in any dynamic analysis. In the presented benchmarking, results from natural frequency analysis are compared in different steps of acousticanalysis. The first evaluation is done as a standard step in acoustic analysis workflow. As reducedmatrices from FE-model condensation are obtained from natural frequency solution, the same

comparison is done in this step. The engine simulation tool uses reduced stiffness and mass matrixand as a model preparation step is also able to perform natural frequency analysis for eachcondensed body. Results from those calculations are also compared with the solution when


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MSC.Nastran is used. In all three cases, the observed differences between the integrated ABAQUSsolution and MSC.Nastran are the same and have the same trend which is illustrated in Figure 5.Up to 500 Hz results from natural frequency analysis are approximately the same and fully satisfyrequirements for acoustic analysis. Above this frequency differences between calculatedeigenfrequencies and mode shapes become larger due to mode perturbations. The main reason for those perturbations is in the different representation of shell elements in MSC.Nastran andABAQUS. This effect can be clearly seen on power unit parts that are modeled only with shellelements such as oil pan, engine brackets, gearbox housing and manifolds. One of the focuses infurther investigations should be the development of best praxis methodology in acoustic analysis

based on ABAQUS solver and improving user experience and modeling guidelines especiallywhen shell and beam elements are used.

Figure 5. Power unit natural frequencies.

4.3 Bearings

Modeling the dynamic behavior of the crank train of engines is a very complex task due to thesuperimposition of its kinematics movement and deformation vibrations. In addition the non-linearity of the crank train bearings affects the forces inside bearings. Those forces are one of themain sources of power unit vibrations. Therefore, detailed bearing result evaluation is performedto check the quality of obtained results when different FE-solvers are used. The comparison shown


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in Figure 6 clearly points out that dynamic result for body matrices, derived from each of the twoinvestigated solvers, are similar and correct.

Figure 6. Forces and orbital path at main bearing 4.

4.4 Torsional Vibrations

One of the main noise sources in an IC engine are crankshaft torsional vibrations due to engine

speed irregularity. On one side they highly determine timing drive excitation and on the other theyare the main source of gear noise. The comparison between calculations based on the integratedABAQUS solution and MSC.Nastran condensed FE-models shows that crankshaft torsional

behavior is not influenced by FE solvers. Together with prior bearing analysis it can be concludedthat determination of power unit global stiffness, especially bearing supporting structure, is thesame when both tools are used. Also it can be concluded that power unit dynamic behavior inlower frequency range up to 500 Hz is represented in the same way in both FE-solvers.


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Figure 7. Torsional vibrations at crankshaft pulley and flywheel side.

4.5 Engine Mount Vibrations

To improve the noise and vibration behavior of a vehicle, the engine mount vibrations, whichexcite the vehicle frame at different frequencies, should be carefully analyzed. The vibrations of the frame and its components, which are excited by engine mount forces with higher frequencies,can increase the sound pressure in the vehicle compartment. The benchmarking presented in this

paper showed the highest differences in view of engine mount vibrations. In the lower frequencyrange up to 500 Hz, the response is almost similar for both FE solvers. Above this frequencyhigher differences occur and can be explained with differences in natural frequencies and modeshapes. Engine mount vibrations are mainly determined by the following parameters:

excitation coming from power unit global power unit global eigenmodes engine mount frequency dependent characteristics engine mount bracket local modes.

In the presented benchmarking the first three parameters can be neglected as the same excitationand the same engine mount frequency dependent characteristics have been used. Also global

power unit global eigenmodes can be neglected, as natural frequency analysis showed nodifferences for those eigenmodes. In the analysed engine mount, brackets are modeled using shellelements and local eigenmodes have higher difference in higher frequency range. In the presented

benchmarking in the MSC.Nastran model standard CQUAD4 and CTRIA3 elements are used and


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in EXCITE FEA C3D3R and C3D4R ABAQUS elements with enhanced hourglass control areused.

Figure 8. Accelerations at mounting brackets at 3000 rpm.

4.6 Structural Borne Noise

One evaluation criterion is the vibration levels representing the structure borne noise at discretestructure points. A more global statement about the structure borne noise emission can be made bysumming up the vibration levels in the acoustic relevant frequency range to integral levels. To

perform these acoustic evaluations a transformation of the results into the frequency domain isnecessary. This is done via FFT (= Fast Fourier Transformation).

Mostly velocity levels or acceleration levels on the engine surface are used to judge the quality of the design. To get a better overview of the power unit acoustic behavior, results from FFT analysisare usually transformed to the octave and 3 rd octave band levels are evaluated for characteristic

points on engine structure. The comparison between calculations based on the integratedABAQUS solution and MSC.Nastran shows that vibration levels are affected by FE-solver inoctave and 3 rd octave bands near to local resonances. For the analyzed engine, ABAQUS gives aslightly higher response in bands where local eigenmodes occur. In Figure 9 3 rd octave bands

velocity levels are compared at two locations of the engine in surface normal direction. Onecomparison point is located at the engine block structure (volumetric elements) and the second atthe oil pan (shell structure). Results on the engine structure show identical results, the results at the


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oil pan point out differences in the acoustically relevant frequency range above 250 Hz. Thereason for these differences are mainly found in the different element types for the shell elementsfor both FEM solvers.

Figure 9. Surface velocities at 3000 rpm.

Another evaluation criterion to judge power unit acoustic behavior is the evaluation of structural borne noise for complete radiating surface. To evaluate these results it is necessary to performadditional calculation steps and recover calculated vibrations from condensed FE-model havingonly retained nodal DOFs and retained eigenmodes back to the uncondensed FE-model. Thiscalculation step is known as direct data recovery and recovery can be done with external FE-solver or with direct data recovery inside the engine simulation tool. Using direct data recovery makesthose calculations much faster and overall time needed for acoustic analysis much shorter. To beable to use this functionality it is necessary to store the transformation matrix, having omitted a setof DOFs during condensation of the FE-model. Both ABAQUS and MSC.Nastran have this

possibility and are supported for direct data recovery. As an alternative solution to calculate entirestructural vibration results, a forced vibration analysis (for example mode based steady statedynamic) could be carried out using all forces of the engine body, calculated by the dynamicmulti-body analysis. The three different possibilities available to calculate surface levels areshown in Figure 10. A possibility of direct data recovery with ABAQUS is currently not available,although forced response analysis can be performed.


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Figure 10. Approaches and methods for calculation of entire structural vibration results.

Typical results of data recovery are normal velocity integral levels in different octave and 3 rd bands level evaluated for complete power unit structure or integrated in addition for characteristicradiating surface area. Figure 11 shows velocity level results for the 2kHz octave band (defined asintegral values for the frequency range between 1420 and 2840Hz, covering range of three 3 rd octave bands with center frequency at 1.6, 2 and 2.5 kHz in Figure 8. The comparison between

both versions is sufficient, taking different element formulation into account.

Figure 11. Integral velocity levels for 2 kHz octave at 3000 rpm.


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5. Conclusions

The presented paper outlines a workflow for the usage of ABAQUS linear FE-solver for dynamicand acoustic analysis including multi-body simulation by an example of a power unit under firedcondition. This workflow is fully integrated into the engine simulation tool AVL EXCITE. For thespecific FEM tasks modal analysis, matrix condensation and dynamic response analysis theABAQUS linear FE-solver is implemented into EXCITE as a module called EXCITE FEA. The

paper discusses the comparison of all workflow steps in view of system resources and maindynamic and acoustic results between the FE-solvers ABAQUS and MSC.Nastran. In addition theimpact of different ABAQUS element formulations on the system resources are presented.

It can be concluded that ABAQUS is a very competitive tool for dynamic and acoustic analysisand the integrated workflow ensures a compact and user friendly environment for the simulationof engine and power unit vibration and structure-borne noise. The system resources are highlydependent on the element formulation used in the models and should be further investigated. Afurther reduction of system resources, especially memory, would be required. Additional effortwould be oriented on development of the best practice methodology to use all advanced modelingtechniques available inside ABAQUS linear solver with the aim to increase accuracy of results of acoustic analysis.

6. References

1. Rasser, M., T. Resch and H.H. Priebsch. Calculation of Coupled Torsional, Bending andAxial Vibrations and Resulting Stresses in Crankshafts, MTZ Worldwide, pp.25-28, 2000.

2. Priebsch, H. H., J. Krasser, Simulation of Vibration and Structure Borne Noise of Engines A Combined Technique of FEM and Multi Body Dynamics, CAD-FEM Users Meeting,Bad NeuenahrAhrweiler, 1998.

3. Priebsch, H T. Resch, "Simulation and Optimisation of Engine Noise Predictive Input for the Development Process", 16th RIETER Automotive Conference 2003, Luzern.

4. H. Pramberger, L. Jun, "Engine Simulation for Vehicle Noise Reduction, SAE Conference2003, China.

5. Offner, G. et al, Quality and Validation of Cranktrain Vibration Prediction Effect of Hydrodynamic Journal Bearing Models, Proc. of Multi-body Dynamics Monitoring andSimulation Techniques III, pp. 255-271, 2004.

6. Parikyan T., T. Resch T and H.H. Priebsch, Structured Model of Crankshaft in theSimulation of Engine Dynamics with AVL EXCITE, ASME Fall Technical Conference,ICE-Vol. 37-3, Argonne, 2001.

7. Todorovic G., T. Parikyan, Automated Generation of Crankshaft Dynamic Model to ReduceEngine Development Time, SAE Paper Offer 03P-336, 2002.

8. Ma, M-T., et al, A Fast Approach to Model Hydrodynamic Behaviour of Journal Bearingsfor Analysis of Crankshaft and Engine Dynamics, Proc. of the 30 th Leeds-Lyon Symposiumon Tribology Transient Process in Tribology, pp. 313-327, 2003.

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