Modern Tools and Methods for Low Noise Engine Development

Transcription

1 Modern Tools and Methods for Low Noise Engine Development C. Nussmann, C. Steffens 1, M. Atzler 2 1 FEV Motorentechnik GmbH Neuenhofstraße 181, Aachen, Germany marketing@fev.com 2 Institute for Combustion Engines, RWTH Aachen University Schinkelstraße 8, Aachen, Germany office@vka.rwth-aachen.de Abstract Modern diesel and gasoline engines pose wide challenges for the NVH development process. Lightweight design in combination with increased specific power contrasts with customers demands on excellent noise and vibration behavior. Meeting these challenges, FEV has established a wide range of CAE tools and methods to optimize noise and vibrations, even before hardware prototypes are available. Starting with the analysis of main engine components, the FEM based software tool FEV DIRA (Direct Impulse Response Analysis) is used to assess and improve structure born noise transfer. During the next development phase a virtual prototype of the powertrain is built up within the software FEV Virtual Engine. This MBS based software product uses flexible FEM structures to predict structure born noise and vibration during engine operation. To meet the specific NVH development targets, the degree of model depth can be increased by a stepwise integration of crank train, timing drive, valve train and gear box in the powertrain model. Subsequently the radiated noise is calculated using the software tool FEV FERS (Fast Evaluation of Radiated Sound Power). Integrating the presented CAE tools in the engine development process is the basis to meet the ambitious NVH targets of modern combustion engines. Within this publication the usage of the tools will be described and examples will be given. Additionally, the simulation accuracy could by proved by comparative measurements. 1 Introduction The development of today s powertrains is subject to a multitude of requirements. Diverse targets such as high performance, low emissions and fuel consumption, high comfort and quality as well as packaging need to be accomplished in shorter time-to-market and at reduced costs. Especially the noise and vibration behavior has become a key factor in customer quality perception, not only in luxury class vehicles. For an optimum NVH behavior it s crucial to focus on noise and vibration already in the concept phase of the powertrain development. Specific targets e.g. natural frequencies, vibration level and noise radiation must be defined already in the beginning of the development phase and continuously checked throughout the entire development process. The consequent use of virtual development methods focusing on powertrain induced vehicle NVH behavior is prerequisite for cost and time effective development avoiding NVH troubleshooting during vehicle integration. The usage of CAE tools and methods within the NVH development process will be described in this paper. In addition, the simulation accuracy could by proved by comparative measurements. 4337

2 4338 PROCEEDINGS OF ISMA2010 INCLUDING USD NVH Development Integrated in Powertrain Design 2.1 General Future vehicle development activities will be characterized by the trade-off between economy and driving comfort. All of the world s OEM s are currently working on new engine technologies, which are seen as innovations that tend to drive market shares. The driving forces behind these innovations are stringent legislative demands on exhaust quality and consumer expectations on fuel economy. Product price is defined by perceived quality and, in this respect, torque output and NVH are the major engineering targets. A primary goal for new engines as well as those that are currently under development is to be Best-in- Class concerning NVH behavior. However, new technologies such as direct injection engines generally increase the internal forces within an engine via higher combustion peak pressures and inertias from higher nominal speed. Such increases are generally an impediment for good NVH characteristics. Effective NVH development requires a systematic integration process. From the start of the process, at the concept definition stage, CAE calculations must be applied. NVH engineering is a discipline of interaction, whereby, all vehicle systems require individualized attention. The tools that are applied must have the capability to simulate the entire noise generation path beginning with combustion and ending with the radiated noise. Engine components can also be assessed and optimized with respect to their NVH behavior. According to FEV s development strategy NVH simulation methods are integral parts of the powertrain design process (Fig. 1). Calculation models, techniques and tasks are continuously adapted and upgraded. Starting with comparable simple models, the complexity of the models and the calculation procedures consecutively is increased with a full virtual powertrain at the end. Thus, in-time NVH input with respect to design freezes and consistent NVH support is guaranteed throughout the entire development process. Fig. 1. NVH development integrated in powertrain design

3 VEHICLE NOISE AND VIBRATION (NVH) 4339 After establishing of the powertrain concept based upon the consistent use of virtual methods detailed design decisions must be made in the subsequent design phase. The design is continuously examined for possible acoustic weaknesses especially with respect to critical acoustic interactions between the individual powertrain components. Besides structural single component optimization also NVH simulation of the complete powertrain has to be performed, therefore. The NVH simulation of the total powertrain orientates to test bench measurements and includes e.g. rev-up simulation which can be interpreted as a virtual NVH test bench run. The application of analytical, CAE-based and enhanced experimental investigation tools leads to a rapid and problem-focused development process. Coordinated cross-talk between experiment and simulation stimulates the development of advanced calculation procedures 2.2 Simulation Process To simulate the acoustic behavior of the powertrain, FEV uses a hybrid approach. Multi-Body Simulation (MBS), Finite Element Method (FEM) and airborne noise simulation can be combined for most accurate results. In Fig. 2 the entire calculation procedure is shown. Fig. 2. NVH simulation process The calculation of the excitation mechanisms relevant to the vibration behavior (e.g. crank and valve train) is carried out by MBS. For consideration of the component structural dynamics, the crankshaft and the housing structures are integrated as FEM flexible bodies in the MBS model. Thus, the dynamic effects due to structure flexibilities (e.g. crankshaft bending and torsion, brackets natural vibration modes etc.) can be calculated. In the subsequent Forced Response FEM calculation the FEV DIRA method is used. The FEM model is loaded with bearing reaction forces. According to FEV s strategy the FEM calculations are performed in time domain enabling entire speed rev-ups within one single calculation with audible noises as result. Thus, various acoustic layouts and optimizations can be investigated with regard to frequency behavior, overall level and sound quality. Aside from the evaluation of the surface velocity, the airborne noise radiation can be calculated depending on the application. The airborne noise calculation can be performed

4 4340 PROCEEDINGS OF ISMA2010 INCLUDING USD2010 with varying degrees of detail from simple approaches with empirically determined radiation loss up to sophisticated calculations by Boundary Element Method (BEM) or the new FEV approach FERS (Fast Estimation of Radiated Sound Power). This calculation methodology developed by FEV has been used and verified successfully [9]. 3 Tools and Methods 3.1 FEV Virtual Engine Generally, the most important NVH relevant excitations in a passenger car powertrain can be identified as follows: Cylinder pressure force due to the highly unsteady combustion, Main bearing reaction forces including gas and mass reaction forces of the cranktrain, as well as crankshaft vibration due to gyroscopic effects of the flywheel or the crankshaft torsion, Piston side forces including secondary motion, Camshaft bearing reaction forces including mass forces, opening and closing impacts and bearing impacts, and camshaft driven components like high pressure fuel pump, Timing drive forces Consideration of all of these important excitation mechanisms requires a very complex simulation model. However, in principle, such a procedure is conceivable but, depending on the design status, it can be difficult to acquire the necessary input data. Therefore, early in the design stage, assumptions and practical simplifications must be made when analyzing and optimizing powertrain components regarding NVH. Missing data can be taken from the large data base at FEV. Calculation of the excitation mechanisms presupposes a simulation model that considers the dynamic behavior of all of the rotating components (cranktrain, valvetrain, timing drive,) as well as the dynamic behavior of the housings. The method that is utilized for this purpose is based on a combination of FEM and MBS. The FEM models are used to simulate and to analyze the vibrational behavior of the housings, of the crank train / valve drive and its single components; the purpose of the MBS model is to simulate the whole body movement of all of the moving and rotating components as well as their nonlinear coupling. Figure 3 shows an example of such a MBS model. In this case, FEV Virtuel Engine was used. On the left one can see the rotating components while the housing is faded out and on the right side the housing structure is depicted. The crankshaft and the camshaft are implemented in the MBS model as flexible bodies as well as the complete housing structure, since their vibrational behavior significantly influences the dynamic behavior of the powertrain. Therefore FEM models are dynamically reduced using the modal condensation approach by Craig-Bampton [3]. In contrast, the con rod, piston, valves, etc. are considered as rigid bodies. Multi-Body Analysis yields translational and torsional displacements together with their time derivates for the entire powertrain including the vibrations of the flexible components. Furthermore the reaction forces in the bearings of the moving components will be determined. A Fourier analysis of the time signals is performed to obtain an acoustical assessment. Analyzing the bearing reaction forces enables the analysis and assessment of the excitation forces in the powertrain. Characteristics are evaluated and weak points can be detected. Further more the vibrations on the flexible structures can be used to assess the powertrain regarding noise quality issues like roughness. For overall noise level assessment of the powertrain these forces can be used as an input for the subsequent Forced Response Analysis in order to calculate surface velocities and the vibrations of the assembly mounting brackets. To investigate the dynamic behavior of the complete engine in the speed range of interest, a speed sweep calculation is performed. In this manner, the critical engine orders or resonant frequencies can be identified.

5 VEHICLE NOISE AND VIBRATION (NVH) 4341 Fig. 3 MBS Model of a powertrain Concept Phase In the early design phases principle and conceptual decision have to be made. In Europe, the 4-cylinderinline engine is still by far the most common type of engine. Due to the continuing downsizing trend fourcylinder in-line engines will be even more widely used in the near future. Through the constant increase in specific performance and the growing use of 4-cylinder engines in higher vehicle classes, the requirements on the acoustic behavior will increase further. h 2 FM, 2 = 4 λs mosz (2 π n) 2 (1) with m osz = oscillating mass (piston, conrod share), h = stroke, n = engine speed, λ s = pushrod ratio To counterbalance the resulting loss of comfort, a variety of today s 4-cylinder engines are equipped with balancing systems. Through balancing shafts that counter rotate at double crankshaft speed, the oscillating mass forces of the second engine order are compensated. Aside from oscillating mass forces, speed dependent tilting moments of the second engine order can raise difficulties in 4-cylinder engines. These can be compensated by an additional vertical offset of the balancing shafts. In the development of the investigated 4-clyinder engine, numerous balancing shaft variants were analyzed. In Figure 4 (top), the results for a vertical offset variant are shown. The second order of the tilting moment on the engine block in motored condition is depicted with and without vertical offset. It can be clearly seen that the tilting moment is almost eliminated through a vertical offset. Without vertical offset the increase of the tilting moment increases almost linearly with increasing engine speed. The clear advantages under motored conditions are obliterated by disadvantages during full throttle condition. At full load compensation effects between gas and mass forces normally occur: This effect is illustrated in Figure 4 (bottom). It shows the second order of the resulting tilting moment on the engine with and without vertical offset at full load. Through the full compensation of the mass moment, a significant deterioration can be noted up to about /min. The variant without vertical offset shows a minimum at about /min; in this point gas and mass forces balance out each other nearly completely. This effect leads to a strong load dependency of the engine sound of engines with a vertical balancing shaft offset, while engines without offset tend to lower noise levels at high way conditions. Therefore the vertical offset can be used to design the engine sound to the desired character in the field between sporty and luxury.

6 4342 PROCEEDINGS OF ISMA2010 INCLUDING USD2010 Torque [Nm] Torque [Nm] Motored Full Load Fig. 4 2 nd order oscillating torque with Height Offset w/o Height Offset 50% Height Offset with Height Offset w/o Height Offset 50% Height Offset Rotational Speed [1/min] Detailing Phase In the concept phase, the engine was planned through a consistent use of virtual methods. In the subsequent design phase, detailed design decisions must be made. In this phase, the design must be continuously monitored for possible acoustic weaknesses. Since components that were optimized separately can lead to unwanted acoustic interaction once they are assembled, acoustic calculations must be also performed on the total powertrain. Only in this way negative acoustic interactions of the components can be detected and eliminated at an early stage. Sum Level of bearing force Fig. 5 Reaction forces of the fuel pump drive Engine Speed [rpm] 10 db In Figure 5 an example of the optimization of the phasing angle of the high pressure fuel pump, which is driven by the camshaft, is given. A band pass of the summation of the camshaft bearing forces as a

7 VEHICLE NOISE AND VIBRATION (NVH) 4343 measure of the acoustic excitation - is depicted over engine speed. In this case the optimization of the phasing promises a reduction potential of the fuel pump related noise of more than 5 db. 3.2 Dynamic Impact Response Analysis (DIRA) The next step in the noise generation path is the evaluation of the structural transfer behavior and determination of the surface velocities. The excitation is transmitted as structure-borne noise to the exterior surfaces of the powertrain components where it is subsequently radiated as airborne noise. This airborne noise enters the passenger compartment at a muffled level. Conversely, the structure borne noise is induced into the vehicle body via the assembly mounts and then further transmitted to the passenger compartment. In the assessment or optimization of powertrain NVH, both values must be analyzed. To predict engine component NVH behavior or the behavior of the entire powertrain procedures must be developed. Today, Finite Element Analysis (FEM) is particularly suited to this purpose. Depending on the development status of the powertrain, high or low detailed models can be used. In early design stages, only simplified models of individual components can be assembled. Nevertheless, with the help of these simple models, it is possible to support the designers by providing basic information for their design decisions. The aim of this process is to produce an initial prototype with excellent NVH features. For instance, a single component model can facilitate the definition of the bottom end design of an engine block. During the design process, a full powertrain model gradually develops as the design progresses. At the end of the process, a full powertrain FEM model is available. In order to calculate the structural transfer behavior, Forced Response Calculations are performed Surface Velocity For this purpose, the excitation that was calculated with the help of MBS is applied to the FEM model and the structural response is then calculated, as a first step, as structure-borne noise velocities on the radiating surfaces. In this case the surface velocity is summarized over the relevant frequency range associated with a specific load and speed (Fig. 6, RHS). Engine Speed [rpm] Drehzahl [1/min] Surface Velocity Surface Velocity Level [db(a)] Fig. 6 Surface Velocity Level Another possibility to analyze the surface velocity behavior is displayed in Fig. 6, LHS. The level is shown as a function of frequency and engine speed. Significant resonances and engine orders can be observed. Therefore, this procedure is suitable to identify structural weaknesses. Dynamic structural optimization is used to minimize the transfer of vibration through the inner vibration path of the

8 4344 PROCEEDINGS OF ISMA2010 INCLUDING USD2010 powertrain s structure. Acoustically beneficial structural modifications target a high dynamic (acoustic) stiffness and simultaneously reduced weight in the engine block. In order to improve the structural transfer behavior, stiffening measures are implemented, such as additional ribs or localized wall thickening as well as contouring and the elimination of structure-borne transfer points. In this context, one major challenge is the NVH simulation of plastic components because the stiffness and damping behavior of plastic is depending on temperature, frequency, filler material and moisture content. Here, FEV uses a comprehensive simulation approach to predict the acoustic behavior of plastics accurately. The simulation is based on the finite element method (FEM) in conjunction with materialrelated modeling. For applications in the engine compartment, this means that the specific boundary conditions prevailing there must be taken into account. By integrating this simulation technique into the engine development process, the acoustics of plastic parts can be optimized at design stage. In order to describe the structure-borne acoustic behavior of plastics, both stiffness and damping properties must be considered. A suitable technique for measuring this is Dynamic Mechanical Analysis (DMA). However, since the DMA only permits measurements of up to approx. 100 Hz to be made, the frequency range relevant to an acoustic analysis of can only partly be covered. This can be overcome by using the time/temperature shift principle (TTS principle). One application of this method will be described in the following. Fig. 7 shows a subcomponent consisting of a plastic cylinder head cover, which is attached to a aluminum cylinder head and front cover by means of an elastomer seal and the equivalent stiffness of the bolts. The elastomer seal was modeled by means of solid elements according to the specifications described in [8]. This model was excited by a harmonic unit force acting at the roof of the combustion chamber. This corresponds to the boundary condition of a measurement which was performed simultaneously. Elastomerdichtung seal Valve Ventildeckel cover Cylinder Zylinderkopf head Fig. 7 FEMmodel of the cylinder head assembly Front Stirndeckel cover Fig. 8 gives a comparison of the measurements and calculated values for non-reinforced plastic at standard moisture content levels. The left-hand graph shows a measuring point in the area of the timing drive and the right-hand graph shows a measuring point at the rear end of the inlet side. A good correlation can be obtained in frequency and amplitude for both measuring positions over the entire frequency range. The pronounced damping effect of the assembly is accurately reflected in the calculation. Also, the significant resonance peak at 1250 Hz from the measuring point at the front end is accurately mapped in amplitude and frequency by the simulation. Outside the area of this distinct peak, there are some deviations in level between measurement and calculation. The consistently inconspicuous curve at the rear measuring point is accurately depicted in the simulation.

9 )] linear ( db [ [Geschwindigkeit]/[Kraft] linear ( db [ [Geschwindigkeit]/[Kraft] linear ( db [ [Geschwindigkeit]/[Kraft] )] linear ( db [ [Geschwindigkeit]/[Kraft] VEHICLE NOISE AND VIBRATION (NVH) 4345 Velocity/Force [db(1mm/s/n)] Measurement Simulation Fig. 8 Measured and simulated transfer functions of the cylinder head assembly Velocity/Force [db(1mm/s/n)] Measurement Simulation In the NVH development process, the effects of variations on the dynamic behavior of the components must be correctly predicted. In order to asses how well the dynamic simulation developed during this project does justice to this requirement, the effect of the choice of material is presented below. Here three variants under investigation are plotted graphically, each separated into measurement and calculation (see Fig. 9). The analysis is carried out at a measurement point at the rear end of the valve cover on the inlet side. At this measuring point, the non-reinforced material has a noise level of about 5dB higher in the 1200 Hz resonance, reflecting the results of the calculation. Similarly, the significant reduction in the amplitudes of the non-reinforced plastic up to 2500 Hz was correctly mapped. The slight difference in the dynamic behavior of the two reinforced plastics is similarly predicted by the calculation. All trend statements relevant to the acoustics in this example are also reflected in the simulation. Velocity/Force [db(1mm/s/n)] 10 Measurement )] 0 PA66 PA66-GF35-10 PA66-M Velocity/Force [db(1mm/s/n)] 10 Simulation )] 0 PA66 PA66-GF35-10 PA66-M Fig. 9 Effect of the filler material on measured and simulated transfer functions of the cylinder head assembly Summarizing, the discussed simulation procedures lead to accurate results and effectively facilitate the optimization and evaluation of the entire powertrain structure Engine Mount and Accessory Vibrations To assess the structural noise induction into the vehicle body, the powertrain mount vibration is analyzed in addition. These vibrations are dominated by global vibrations such as powertrain bending modes and/or more local vibrations of components that may be bolted to the engine such as accessories, engine mount

10 4346 PROCEEDINGS OF ISMA2010 INCLUDING USD2010 brackets, oil modules or ECU s (Engine Control Units). Because of the fact that the excitation forces are calculated with the entire powertrain, all excitation mechanisms that influence the mount vibrations are considered. The effects of free inertial forces and moments as well as gas force tilting moments, cylinder pressure fluctuations, and cylinder deactivation can be determined. In the FEM model, the assembly mounts are exactly modeled with regard to their dynamic and static characteristics (stiffness, damping). With this model, the engine mounts can be evaluated and configured regarding their position and structure Accessories In addition to bracket vibrations, another important NVH topic is the noise and vibration behavior of the auxiliaries. As an example, the radiated noise behavior of a cooled fluid alternator is discussed below. Fig. 10 shows the experimentally evaluated radiated noise. There is an obvious noise problem at around 1170 Hz that was analyzed using MBS and FEM calculations. In this context, it was essential to consider the fluid in the simulation model. A sound level amplification could be observed in comparison with the behavior without the fluid. The corresponding mode shape is displayed on the right side of the figure. The bending of the alternator leads to a deformation of the alternator housing. The excitation initiated by the rotor bending is transferred via the bearings and supporting walls to the outer surface. Based on this result, countermeasures were determined. Stiffening the rotor and the bearing walls as well as circular ribbing on the outer part of the housing successfully lead to a significant reduction of the peak at 1170 Hz. Noise Problem 90 [Lp] rpm Measurement of Sound Pressure Fig. 10. Example of auxiliary NVH optimization Mode f = 1167 Hz 3.3 Fast Estimation of Radiated Sound Power (FERS) FEV has developed a calculation procedure to predict the noise radiation based on previous surface velocity simulations. This approach is available as in-house tool named Fast Estimation of Radiated Sound Power (FERS) and can be used in conjunction with FEV DIRA results from single component structural analysis or full powertrain simulations. Compared to commercial acoustic software tools based on Boundary Element Method (BEM), this simplified approach is very effective regarding simulation time. By implementing the approach in FEV s NVH development process, the procedure becomes user friendly and pre- and post processing effort could be reduced. This becomes primarily important, when considering the short development times in the engine development process. As shown in Fig. 11 (LHS), the approach bases on the integration of single sub surfaces sound pressure share. By means of equation 2, the sound pressure at one sensor point is calculated out of the normal nodal velocities and the corresponding surface for each frequency step. Depending on their phase relation and

11 VEHICLE NOISE AND VIBRATION (NVH) 4347 their distance to the sensor point, attenuation or amplification of the sub surfaces sound pressure shares can appear. Fig. 11 Integration of sound pressure shares (LHS) [4] and envelope surface approach for sound power calculation (RHS) [5] To estimate the sound power of the analyzed component, the envelope surface approach is used. Fig. 11 (RHS) shows an example, using a half sphere as envelope surface. FEV uses 400 sensor points on the envelope surface in order to get accurate results for components with a highly pronounced radiation characteristic. The positioning of the sensor points is done automatically using the components geometry as input. Using equation 3, the sound power can be calculated by summing the intensities for each sub envelope surface. N jωρ0vn jk0 p( x, y, z) = e 2πr n= 1 n rn ΔS (2) with p = sound pressure, ω = circular frequency, ρ 0 = density of air, v n = structural velocity, r n = distance to sensor, k 0 = wave number, ΔS = sub surface P = I n ds S (3) with P =sound power, I = sound intensity, n = normal vector, S = control surface Level [db(a)] Surface Velocity 140 Sound Power dbrel (Surface Velocity) = 2.5e-8 m 70 4 /s² dbrel (Sound Power) = 1e-12 W Fig. 12 Estimation of radiated sound power Radiation Efficiency[dB] FEV FERS FEV FERS Maidanik Maidanik piston full radiator acoustic short circuit radiation

12 4348 PROCEEDINGS OF ISMA2010 INCLUDING USD2010 The effect of sound power prediction on the frequency spectrum of an example component is shown in Fig. 12 (LHS). Surface velocity level and sound power level can be compared, because the db reference values are derived from basic acoustic equations [7]. The thin wall structure with simple geometry has structural properties of aluminum with 5 mm wall thickness. In the lower frequency range, the sound power spectrum shows significant lower values than the surface velocity spectrum. Especially the first natural frequency at 100 Hz is no more detectable in the sound power spectrum. For frequencies above 2000 Hz, surface velocity and sound power mainly behave identical. The structure radiates the whole energy in the far sound field. The reason for this behavior can be explained by Maidanik s theory of sound radiation for plates [6]. In Fig. 12 (RHS) the radiation efficiency obtained by FEV FERS is compared with the theoretical behavior of a simple plate with comparable dimensions. In the frequency range below the first eigenmode, the structure behaves like a piston radiator; radiation efficiency rises over frequency with 20 db/decade. The amplitude in this region depends on the surface area of the component. Starting with the first natural frequency, the noise radiation is attenuated by acoustic short circuits. Especially at the resonance peaks, the sound radiation is decreased significantly by attenuating interference. For higher frequencies the behavior turns into full radiation; the efficiency is close to zero. The cut-off frequency depends on the component s structural properties; full radiation is achieved, when structural wavelength is higher than acoustic wavelength. Level [db(a)] Aluminium 5mm Steel 2mm Surface Velocity dbrel (Surface Velocity) = 2.5e-8 m 4 /s² Level [db(a)] Sound Power 140 dbrel (Sound Power) = 1e-12 W Aluminium 5mm Steel 2mm Fig. 13 Surface velocity and sound power of aluminum vs. steel component The need for noise radiation prediction becomes obvious, when different materials or wall thicknesses are compared during simulation activities. As an example, Fig. 13 (LHS) shows the surface velocity thirdoctave spectrum of the aluminum part with 5 mm wall thickness compared to the spectrum of the part with identical geometry and 2 mm thickness with steel properties. Despite of local differences, the general behavior and maximal values are very comparable. By contrast the sound power third-octave spectrum (Fig. 13 RHS) shows significant differences in shape and level between aluminum and steel. While the steel component behaves inconspicuous over the whole frequency range due to high radiation loss, the aluminum part shows a 10 db higher 315 Hz third-octave and significant higher values in the mid frequency range. As a result of this study, the aluminum part should be optimized by contouring or ribbing using the higher capabilities of the casting process. Ratings and recommendations based on sound power results can differ considerable from those based on surface velocity results, as this example for a thin walled structure shows.

13 VEHICLE NOISE AND VIBRATION (NVH) Virtual Vehicle Interior Noise Simulation (V-VINS) Besides the airborne and structure-borne noise radiation of the powertrain also the transfer behavior of the body plays a decisive role with regard to engine-induced interior vehicle noise. This means that conclusions as to the engine-induced interior noise cannot be drawn based on the calculations of the acoustic behavior of the powertrain alone. For an engine development that is targeted with regard to vehicle interior noise, FEV has developed a procedure that allows the simulation of the vehicle interior noise. This was done by combining the simulation of noise radiation with the experimental Vehicle Interior Noise Simulation (VINS) method [1]. VINS is an experimental transfer path analysis developed by FEV for the generation of synthetic interior noise. By combining experimental determined transfer properties of the vehicle body with powertrain structure-borne and airborne noise measurements, a synthetic interior noise is generated. For each noise path of the vehicle body transfer functions are determined experimentally. Through a phase-adjusted addition of all noise shares (gained from test bench measurements) with the vehicle body transfer functions, the engine induced interior noise can be entirely synthesized. In general, the interior noise can be divided into structure-borne and airborne noise. In many vehicles the structure-borne noise dominates the interior noise in the lower frequency range, whereas the airborne noise is dominant in the higher frequencies. Interior Noise based 2000 Interior Noise based 2000 on Measurement on Simulation Fig. 14 Vehicle interior noise simulation By combining VINS with the simulation of structure-borne and airborne noise radiation to a Virtual Vehicle Interior Noise Simulation (V-VINS) an assessment of the interior noise is possible even before the first prototype [2]. An example is given in Fig. 14. Thus, the engine-induced interior noise of a vehicle can be calculated and optimized prior to the engine s prototype phase thus saving development time. Furthermore, the engine can be designed and optimized for a best-in-class vehicle interior noise beginning in the concept phase. This method facilitates a targeted NVH engine development with a focus on excellent vehicle interior noise behavior.

14 4350 PROCEEDINGS OF ISMA2010 INCLUDING USD Summary Virtual NVH development tools are prerequisite for cost and time effective powertrain development with regard to excellent vehicle interior noise and vibration comfort representing one of the major market acceptance key factors. The use of virtual NVH prototypes reduces powertrain development time combined with significantly increased NVH quality of first hardware prototype. FEV uses modern CAE tools during the NVH development process to meet the challenges for accurate noise prediction in the virtual product development. Depending on the powertrain development stage, the degree of detailing can be adapted to find the best compromise between modelling effort and simulation accuracy. For this purpose the well proven methods FEV DIRA and FEV Virtual Engine are available. In a continuous improvement process the experience of many engine development projects flow back into these tools or spin off into new tools and methods. Especially the newly developed methodologies for plastic component simulation and airborne noise calculation with FEV FERS could improve simulation models prediction quality. Modern CAE tools allow a target oriented NVH simulation with regard even to vehicle interior noise. At the end, the consequent usage of all presented tools within a virtual development process provides an essential part in the realization of a low noise engine. References [1] N. Alt, N. Wiehagen and M.W. Schlitzer, Vehicle Interior Noise Simulation for Evaluating Prototype Powertrains in the Vehicle (Pt 1 & 2), Automobiltechnische Zeitschrift 5 und 6 (2001) [2] N. Alt, C. Steffens, C. Pilath and G. Eisele, Virtual Vehicle Interior Noise Simulation (VVINS), International Workshop on Virtual Product - Development of Automotive Powertrains, SPA, Belgien (2005) [3] R. Craig and M.C. Bampton, Coupling of Substructures for Dynamic Analysis, AIAA Journal, Vol. 6 (07/1968) [4] L. Cremer, M. Heckl, Körperschall Physikalische Grundlagen und technische Anwendungen 2. Auflage Springer-Verlag, Berlin (1996) [5] F. G. Kollmann, Maschinenakustik Grundlagen, Messtechnik, Berechnung, Beeinflussung, 2. Auflage Springer-Verlag, Darmstadt (2000) [6] G. Maidanik, Response of Ripped Panels to Reverberant Acoustic Fields, Journal of Acoustic Society 34 (1962) [7] S. Pischinger, Verbrennungskraftmaschinen, Vorlesungsumdruck, RWTH Aachen (2007) [8] S. Pischinger, C. Pilath, Entwicklung eines rechnergestützten Verfahrens zur akustischen Optimierung von Dichtungen zwischen Motorstruktur und Anbauteilen, Lehrstuhl für Verbrennungskraftmaschinen, RWTH Aachen, Abschlussbericht zum FVV Forschungsvorhaben Nr. 789 (2004) [9] C. Steffens, C. Nussmann and C. Pilath, Virtuelle Entwicklungsmethoden zur akustischen Optimierung von Antriebsaggregaten, 5. Symposium Motor- und Aggregate-Akustik, Magdeburg (2005)