The rear module of an hybrid bus was redesigned in a joint effort by Volvo andIDMEC in the framework of the european project HCV Hybrid Commercial Vehicle (GrantAgreement No 234019), replacing most of the metallic frame and other components with compositesandwich panels. The redesigned bus overall weight was reduced by approximately900kg. This paper presents the numerical and experimental modal analysis of a sandwichpanel that was produced using the same materials and production process used to produce themain prototype rear component

António A. Fernandes

J. Dias Rodrigues

Paulo C. Neves

Repositório Aberto da Universidade do Porto

18th International Conference on Composite StructuresLisbon, Portugal, 15-18 June 2015https://sites.google.com/site/18thiccs/MODAL ANALYSIS OF A COMPOSITE SANDWICH PANEL USED INTHE STRUCTURE OF AN HYBRID BUSPaulo C. Neves∗, J. Dias Rodrigues†, Antonio A. Fernandes∗∗ Instituto de Engenharia Mecaˆnica (IDMEC)Rua Dr. Roberto Frias, 4200-465 Porto, Portugale-mail: aaf@fe.up.pt† Faculty of Engineering, University of Porto (FEUP)Rua Dr. Roberto Frias, 4200-465 Porto, Portugalweb page: http://www.hcv-project.eu/overview.shtmlKey words: Vehicle, Sandwich panels, modal analysis.Summary. The rear module of an hybrid bus was redesigned in a joint effort by Volvo andIDMEC – in the framework of the european project HCV – Hybrid Commercial Vehicle (GrantAgreement No 234019)–, replacing most of the metallic frame and other components with com-posite sandwich panels. The redesigned bus overall weight was reduced by approximately900kg. This paper presents the numerical and experimental modal analysis of a sandwichpanel that was produced using the same materials and production process used to produce themain prototype rear component.1. INTRODUCTIONThe HCV project is an European Collaborative Project (Large Scale Integrating Project)with a duration of 48 month, from January 1st, 2010 to December 31st, 2013, approved underthe 7th Framework Programme. The HCV project aims to develop urban buses and deliveryvehicles with advanced second generation of energy efficient hybrid electric power-trains inline with objectives in topic SST 2008.3.1.5 - Urban buses and delivery vehicles using secondgeneration hybrid electric technology. The final result will be the demonstration of a passengerbus and a distribution truck with this advanced technology.An 18-ton passenger bus demonstrator and a 6-ton distribution truck demonstrator will befully developed in the project. A low weight body will be demonstrated for the 18-ton bus andsome weight optimisation will be made on the distribution truck. The hybrid electric conceptsin the HCV project aims to further integrate components and to improve the hybrid systemwith new gearboxes, electrification of auxiliaries and advanced control system. A total of 12demonstrators will be exposed within the project. New advanced technologies will be testedand made public across Europe from east to west and north to south in a variation of climatesand seasons.Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesThe rear module of an hybrid bus was redesigned, produced and assembled into a prototype,replacing most of the metallic frame and other components including panels with compositesandwich panels. The chassis was modified to fit the sandwich panels and to allow furtherweight reduction. Figure 1 shows the final configuration of the sandwich panels.Figure 1. Sandwich panels in the rear module.Other parts of the structure were also redesigned by Volvo, using metal and foam based sand-wich structures and other materials, although this particular part of the work is not addressedin this paper. For the rear module, four independent composite sandwich panels were producedby vacuum infusion: the largest rearmost panel, two panels which received two seats each anda floor panel just in front of the rear door. The largest sandwich panel replaces all transversalstructural elements in the rear module and supports totally three seating passengers and severalequipments and systems (exhaust system and all equipments at roof level, which also introducethermal and vibration loading) and to a great extent the longitudinal beams where the engine ismounted and the cooler.Every proposal of modified bus must satisfy specified axel loads, which were verified sub-jecting the whole structure to unit gravity in the vertical direction.The overall weight reduction was approximately 900kg, 80% above the initial objective.This reduction was confirmed by weighting the prototype.It was calculated using the numerical model that, without inserts, the main rear sandwichcomponent weights 62.3kg, and the floor panel and the two podesters, together, weight 25.8kg.effective reduction of weight of the structure was thus estimated to be 308kg (taking only therear module modifications into account). Figure 2 shows the prototype main sandwich panelsoon after being demoulded.One important achievement is that the panels are produced in one single process and aredirectly assembled into a more simplified structure. The accuracy in the production of the sand-wich panels proved to be adequate for a direct assembly without the need of forced adjustments.It also represents the abolition of a number of components which were integrated into a singleone and the consequent reduction of assembly operations and time consumed.Last but not least, despite the experience of the workforce involved in the production ofthe panel, margin still exists for process optimization. When adopted for industrial production,several improvements and more detailed planning will contribute to an increased efficiency.2Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesFigure 2. Main sandwich panel finishedNevertheless, the implemented prototype process proved competitive given the mentioned inte-gration and assembly simplification.The technology and experience are currently being used by Volvo and IDMEC in the for-mulation of an alternative bus concept in HCV, where conventional steel frames are replaced bystructural sandwich to an even larger extent. Furthermore, the number of passengers seating inthe new rear sandwich component is increased: in the current stage, 9 passenger seats are con-sidered. The design is also complicated by the fact that the energy cells for a potential variantwith electric drive should be possible on the roof, close to the rear door, and several suspensioncomponents are directly mounted onto the composite sandwich.Prior to the production of the sandwich components, a plate was produced for testing. Al-though the materials used were the same, the lay-up is of the sample sandwich is different.Tests were performed by IDMEC to measure the properties of the real laminates to assess theveracity of the numerical model used in the design of the rear module. The experimental modalanalysis tests were performed at the Laboratory of Vibrations of the Departement of MechanicalEngineering, FEUP. The complete report of the experiments is presented in this paper.The plate is 522mm long and 142mm wide. The lay-up was assessed by microscopy andthe thickness of the individual plies was averaged out from several measurements at differentlocations. The lay-up considered representative of the real sandwich and used in the numericalmodels is:0.35mm of carbon fibre fabric at 0/90◦ (0◦ is the longitudinal direction of the plate)0.30mm of unidirectional carbon fibre at 0◦0.35mm of carbon fibre fabric at 0/90◦29mm of PVC foam core (40kg/m3)0.40mm of carbon fibre fabric at 0/90◦0.30mm of unidirectional carbon fibre at 0◦0.40mm of carbon fibre fabric at 0/90◦3Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesThe differences in ply thickness on both skins were considered to be due to slight differencesin volume content of fibres in the laminates. The material properties used in the numericalmodel were not changed to account for this difference.The composite sandwich plate received from Volvo for testing is shown in Figure 3.Figure 3. Plate cut out from rear sandwich componentAn undamped model was used.2. PRELIMINARY MEASUREMENTSPreliminary measurements were made to determine the natural frequencies and mode shapesin free-free boundary condition. To materialize this condition, the plate was supported by twosoft foam pads which were expected to significantly reduce the upper limit value of the fre-quency range containing the rigid body modes of the assembly, while minimizing the externalground noise effects.To obtain some frequency response functions of the plate an ICP accelerometer (Bru¨el &Kjaer 4507) was attached to one of the faces of the plate using special purpose mounting beewaxand an instrumented impact hammer (Bru¨el & Kjaer 8202) was applied to provide the excitationin several locations of the same skin.Both transducers were connected to a dynamic signal analyser for signals conditioning andanalysis using a Bru¨el & Kjaer 2035 spectral analyzer. The signal analysis was performed insidethe spectral analyzer, providing the set of frequency response functions of type accelerance fora coarse measuring mesh considered in this study. The set-up is shown in Figure 4.Figure 4. Preliminary measurements set-up4Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesIn parallel, the natural frequencies and mode shapes were extracted from a numerical modelusing Abaqus. The model consists of a flat rectangular shell which was modelled using thesame element types (S4), material properties (for the carbon fabric) and section lay-up(exceptfor ply thicknesses, which were modified to correspond to the case of the tested sandwich) asused when modelling the rear sandwich in the numerical model of the complete bus used in thedesign phase. The layers, the thicknesses and the material properties are not the same in thetested plate and in the sandwich components produced for the prototype bus.These preliminary measurements and studies were concluded with the preparation of thenext measurements: the nodal lines of the different modes of vibration were mapped and over-lapped in a single map to allow the correct positioning of the accelerometer. This map of nodallines is shown overlapped by the mesh of the plate in Figure 5. This procedure allowed for theidentification of mesh nodes where the accelerometer could be positioned and the respectivecoordinates were taken to be used in the measurements. It should be noticed that the dark ar-eas correspond to areas with small normalised diplacement (forcedly including the nodal lines)which were avoided for the positioning of the accelerometer.Figure 5. Nodal lines map overlapped by the mesh of the plateFigure 6 shows the magnitude of the frequency response function (FRF) determined directlyon the excitation node – one of the nodes identified using the nodal lines map.5Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesFigure 6. direct FRF obtained with the model3. EXPERIMENTAL MODAL ANALYSIS3.1 Experiment using an instrumented impact hammerAn experimental modal analysis was performed on the same plate. Using the data from theprevious analysis a measurement mesh with 55 points was defined as illustrated in Figure 7.Figure 7. Grid marked on the plateLike in the preliminary measurements, to materialize free-free boundary condition, the platewas supported by two soft foam pads. An ICP accelerometer (Bru¨el & Kjaer 4507) was attachedto the face of the plate opposite to the measurement mesh using special purpose mounting6Paulo C. Neves, J. Dias Rodrigues, and Antonio A. Fernandesbeewax, in a location just opposite to the mesh node 3, marked with a red circle in Figure 7, andan instrumented impact hammer (Bru¨el & Kjaer 8202) was applied to provide the excitation ateach point of the measurement grid.Again, both transducers were connected to a dynamic signal analyser for signals condi-tioning and analysis using a Bru¨el & Kjaer 2035 spectral analyzer. The signal analysis wasperformed inside the spectral analyzer, providing the set of frequency response functions oftype accelerance for the entire measuring mesh considered in this study. The measurementswere made in the frequency range 0–1600Hz. The set-up is identical to the one shown in Figure4.From the measured FRFs, in a first stage, a multi-degree of freedom (MDOF) technique,which uses the ”least squares complex exponential” time domain algorithm, eas applied in orderto identify the natural frequencies and damping ratios values for each mode.In a second stage, the residues were identified with a ”least squares frequency domain”technique and the mode shapes were obtained. The modal parameters were identified by usingthe modal analysis software LMS CADA-PC.3.2 Experiment using a shakerAn second experimental modal analysis was performed on the same plate. The same mea-surement mesh with 55 points illustrated in Figure 7 was used.To materialize free-free boundary condition, the plate was suspended using soft nylon ropes.A shaker (?????) was applied to provide the excitation at point just opposite to the mesh node3, marked with a red circle in Figure 7, and an ICP accelerometer (Bru¨el & Kjaer 4507) wassequentially attached to every node of the measurement mesh using special purpose mountingbeewax.The set-up is shown in Figure 8.Figure 8. Comparison of natural frequenciesAgain, both transducers were connected to a dynamic signal analyser for signals condi-7Paulo C. Neves, J. Dias Rodrigues, and Antonio A. Fernandestioning and analysis using a Bru¨el & Kjaer 2035 spectral analyzer. The signal analysis wasperformed inside the spectral analyzer, providing the set of frequency response functions oftype accelerance for the entire measuring mesh considered in this study. The measurementswere made in the frequency range 0–1600Hz. The set-up is identical to the one shown in Figure4.From the measured FRFs, in a first stage, a multi-degree of freedom (MDOF) technique,which uses the ”least squares complex exponential” time domain algorithm, eas applied in orderto identify the natural frequencies and damping ratios values for each mode.In a second stage, the residues were identified with a ”least squares frequency domain”technique and the mode shapes were obtained. The modal parameters were identified by usingthe modal analysis software LMS CADA-PC.4. RESULTS AND DISCUSSIONFigures 9 and 10 respectively represent the Bode’s diagram and the Nyquist’s diagram ofthe driving point accelerance (A33).Figure 9. Bode’s diagram of the driving point accelerance (A33)The graphic in Figure 11 represents the magnitude of the 55 frequency response functionsof the type accelerance obtained with the linear average of five independent measurements foreach point during the experiment using an instrumented impact hammer, where special care wastaken to ensure a good signal/ noise ratio and a good coherence level for the entire frequencyanalysis band (0:1600 Hz).8Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesFigure 10. Nyquist’s diagram of driving point’s FRF A33 (w)Figure 11. Measured frequency response functions9Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesTable 1 presents the natural frequencies and modal damping ratios identified during theexperiment using an instrumented impact hammer.Table 1. Identified natural frequencies and modal damping ratiosMode ω (Hz) Damping ratio (%)1 5032 5233 8154 10305 11496 13477 1436The graphic in Figure 12 shows the comparison of the natural frequencies extracted in thepreliminary measurements, in the experimental modal analysis with hammer and from the nu-merical model.Figure 12. Comparison of natural frequenciesFigures 13 through 19 show the comparison between experimental mode shapes (above)andthe numerical ones (below). Note that the first six modes extracted with the numerical modelwere not presented since they are rigid body modes and that the plots of the results from ex-perimental analysis do not include the areas of the plate outside the grid limits (approximately10mm wide margins around the plate).10Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesFigure 13. Comparison – mode 1Figure 14. Comparison – mode 211Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesFigure 15. Comparison – mode 3Figure 16. Comparison – mode 412Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesFigure 17. Comparison – mode 5Figure 18. Comparison – mode 613Paulo C. Neves, J. Dias Rodrigues, and Antonio A. FernandesFigure 19. Comparison – mode 75. CONCLUSIONSThe numerical models were able to extract modal parameters that were very approximate tothose effectively identified by experimental modal analysis of the sandwich plate. In fact, thefirst five mode shapes are very approximate. The sixth and, even more so, the seventh modeexhibit more relevant differences despite the fact that some similarity can still be identified.Although it was possible to observe an almost constant proportional relation between thenatural frequencies extracted with the numerical models and those experimentally identified, thenecessary factors to correct the densities or the stiffness properties are to high to be admissible.The modelling strategy and material properties used in the bus structure design was assessedwhen used in the modal analysis of the sample plate produced using the same materials andproduction process. The results obtained in this assessment do not indicate any limitation whenusing this approach to model the bus structure.References[1] Grant Agreement No 234019: Annex 1 – Description of Work; November 18th, 200914

Modal analysis of a composite sandwich panel used in the structure of an hybrid bus

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