cae of vehicle frontal crash analysis - jetir · 2018. 6. 19. · biw (body-in-white) and seat...

© 2018 JETIR March 2018, Volume 5, Issue 3 www.jetir.org (ISSN-2349-5162)

JETIR1803109 Journal of Emerging Technologies and Innovative Research (JETIR) www.jetir.org 596

CAE OF VEHICLE FRONTAL CRASH ANALYSIS 1Mr.Lijo Sebastian

1 Asst. Manager in Designs – FAYAT Group, AMIE mech.

Abstract: This paper introduces the research course of vehicle crash safety performance. A vehicle frontal crash FEA model is

established according to CMVDR294. Vehicle frontal crash simulation calculation is conducted by LSDYNA. Energy absorption area and

transmission route of impact force in frontal crash are introduced in details. Effects of longitudinal beams deformation on vehicle

crashworthiness are discussed deeply through simulation of the longitudinal beams design plan. Finally, tests are made to verify feasibility

of the longitudinal beams design plan and reliability of the vehicle frontal crash model. As result, references are provided for vehicle

longitudinal beams design and improvement for the future and moreover this paper used the Hyper Mesh and LS-DYNA software to

establish a dummy finite element simulation model. The noise and vibration of a poor automotive seat aggravate the interior cabin noise

and discomfort. The automotive seat structural noise and vibration is caused by the transmission of the power train or road vibration into

the seat. The characterization of seat structural dynamics behavior in early design phase assists to effectively improve the NVH quality of

the seat. The seat nonlinear buzz, squeak, and rattle (BSR) noise are the major issues which are directly linked to the NVH quality of the

vehicle. For this purpose, a practical CAE (computer-aided engineering) concept modeling method is introduced and developed for full

BIW (body-in-white) and seat separately. Here, the seat concept model is employed to allow designing the seat structure modifications as

well as examining the effects of the modifications on the rattle noise. Comparisons of the results of the simulation and experiment validate

the developed seat CAE model. Three modifications are proposed to optimize the dynamics of the seat structure to prevent the seat rattle

noise. These modifications are designed to shift or decrease the seat torsion resonance and vibration level, respectively. The results verified

that by modification the seat structural dynamics, the nonlinear events such as rattle noise and in general BSR noise can be reduced or

controlled accordingly.

Index Terms – CAE, CFD analysis using LSDYNA, Hyper Mesh, ANSYS CFX, MATLAB, improve the NVH quality, BIW, flow

simulation, seat structural dynamics, torsion resonance and vibration level …

I. INTRODUCTION

Squeak and rattle in the automobile industry are terminologies to describe short duration transient noises that are generated by the

relative motion or impacts between vehicle parts. In other words, in general, BSR noise is a high frequency audible phenomenon resulting

from two distinct forms of noise: (a) caused by friction between elements under forced road excitation (buzzes, squeaks) and (b) caused by

loose or overly flexible elements with the potential for impact with other elements (rattles). The mechanisms involved in generating squeak

and rattle noise are mainly nonlinear. This nonlinearity and complexity make it very difficult to simulate in a CAE analysis. Therefore, there

are no predictive CAE tools available in the industry thus far. Most CAE methods focus on preventive tools. Some researches in the squeak

and rattle preventive methods [1, 2] have been established at Ford in past years. But here, it is shown that there is a strong link between the

structural dynamics and the rattle noise so that the rattle noise can be predicted and controlled from the structural analysis in early design

phase. The vehicle seat BSR (buzz, squeak, and rattle) noise is one of the major issues which are directly linked to the NVH (noise, vibration,

and harshness) quality of the vehicle. Predicting and improving the seat BSR noise in early design phase is still challenging. This is mainly due

to the complexity, nonlinearity, and uncertainty of the impact mechanism at joints contributed to the rattle. Controlling BSR is becoming

essential with the trend toward using lightweight materials combined with the increase in number of the seat sub-components such as

electronic gadgets [3]. FEA theory and its application were born in early 1960s. At that time computer simulation research on vehicle crash

had been carried out in foreign countries.

However, the research was restricted by the development of computer hardware technology and algorithm theory, actual breakthrough

started in 1986 when LSDYNA succeeded in simulating vehicle large deformation for the first time [2]. Ever since then, computer simulation

technique based on dynamic explicit nonlinear finite element technology has seen its widespread applications abroad.

1.2 CAE SIMULATION AND CONCEPT MODELING

In order to create a concept model, the first step is to decide about the layout and functional components of the structure [10]. Generally,

in concept modeling three major parts should be simplified and modeled.

1.2.1 BEAM-LIKE STRUCTURES

The most primary parts are beam-like structures or main members that define the frame body of the vehicle. Beam-like components are

those members that have small cross sections in comparison with their lengths. Main members of the BIW such as doorsill, pillars, and the

frame of seat structure are modeled using beam elements. Since the shape of the cross sections is considered similar to the corresponding

sections at the physical model, calculated properties of the beam elements such as bending

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Advanced beam model Real cross section Concept beam model

Fig. 1.1 Detailed models (left) and concept model of a selected beam (right)

And torsional moments of inertia reflect good approximation in accordance with real members. According to a market survey, squeaks and

rattles are the third most important customer concern in cars after 3 months of ownership [4]. Furthermore, upcoming electric cars will

highlight the importance of the BSR issues [5]. BSR is generally caused by loose or overly flexible elements under excitation. Modern

advances in the vehicle noise and vibration control engineering have reduced the transmission of the vibration or noise from different sources

such as powertrain or road into the passenger cabin [5]. Predicting and controlling BSR in the early design phase is important to be

investigated. As for rattle simulation, it includes complicated periodical nonlinear impact and needs a special CAE model to directly simulate

these phenomena [6]. The cross section of 1D elements is created approximately like real members by accounting the effect of all shell

elements (outer skin and inner reinforcements) involving beam-like structures. Concept and advanced model of a beam-like structure is shown

in Fig. 1.1.

BEAM-LIKE STRUCTURE CROSS SECTION PROPERTIES

In order to compute the cross section properties, at first we need centroid coordinate. Then moment of inertia about the centroid axes will

be calculated. This moment will be transformed to the global centroid coordinate set using coordinate axes rotation laws. Finally the

transmission law is employed to compute each moment of inertia separately.

Centroid coordinates:

Specify local coordinate axes and their angles to global coordinate axes:

Cross section area calculation and ith section moment of inertia:

Fig. 1.2 Selected beam cross Section and needed geometrical data to Calculate its properties

The mentioned procedures are conducted in MATLAB software. In this section these properties are derived with parametric formula for

arbitrary cross section depicted in Fig. 1.2.




∑

∑

∑

∑

Fig. 1.3 Beam-like structures modeling using standard profiles in the library of finite element software

`

The contribution of section moment of inertia about centroid:

∑

∑

∑

STANDARD CROSS SECTIONS DEFICIENCY

Although previous researches didn’t mention to the selection procedure of the beamlike cross sections, it is not easy to have a standard

section that satisfies all the properties of original one. In other words, it’s only possible for members like Fig. 1.3 that has a standard cross

section. Usually we need to neglect one cross section property (product moment of inertia); because of symmetrical geometry this property is

equal to zero. For a special case and make it clear to you, using a rectangular profile with thickness t (Fig. 1.4), three principal properties of a

cross section (I; J; A) will be calculated. For this reason the Eq. (8.11) must have real answers (t_b, h).




Fig. 1.4 Sample standard cross section

Though the calculation of the moment of inertia in accordance with the above equation is possible, considering the torsional moment

of inertia in this equation makes it more complicated. For beams with arbitrary cross sections, calculation of the moment of inertia in

accordance with Eqs. (8.12) and (8.13) will be complicated. Furthermore, if you can compute that, it is rare to find standard cross section that

satisfies all calculated properties. One reason is that torsional moment of inertia (and warping function) heavily depends on the shape of cross

section. Above materials show deficiencies of standard cross sections and make clear advantages of arbitrary cross sections. In Eq. (8.12) ϕ (y,

z) is warping function that can be computed based on the elasticity theory from the differential equation (8.13). nz and ny are normal unit

vector of the beam cross section.

{

[ ]

∫ (

)

(

) (

)

Fig. 1.5 Steps of reducing detail model joint to concept condensed joint in order to connect beam elements

1.2.2 JOINTS MODELING

Joints are in secondary order of importance and they connect beam-like structure to each other. Panels are other major components that

perfect the layout of the structure. In NVH concept modeling analysis,

Preparing the concept joints is a critical issue. Usually, model reduction methods (static or dynamic) are exploited to reduce the FE model

computation time and cost. According to the characteristics of the model and analysis requirements, an appropriate model reduction method

[11] such as Guyan, CMS, and SEREP has to be selected. Then the large FE model is condensed in a few degrees of freedom by the method

specified transformation matrix (TC) according to Eqs. (8.14)–(8.16), where subscripts A is representative of all DOFs, subscripts i stands for

the DOFs that are going to be omitted, and subscripts b are those that will be kept after model reduction [10]. Since the concept joints connect

the concept beam-like structures (1D beam elements) to each other in their terminal nodes, the detailed joints are condensed into a reduced

description of the stiffness and mass matrices at the boundary nodes.




Here, both Guyan and CMS1 model reduction are tested for reducing & generating concept joints; the results for both have a little

discrepancy in low frequency range. Therefore, as the aim of the concept modeling is focusing on global modes of the structure that are usually

below 100 Hz, Guyan method is used to prepare concept joints. Neglecting the inertia effects in calculating reduced model, Guyan method is

also called static condensation. Equation (8.17) is the general form of the static finite element model, in which the internal and boundary DOFs

are reordered to make the calculation process straightforward. Then by a little manipulation on block rows of this equation, the Guyan

condensation matrix (TG) can be extracted according to Eq. (8.18). The procedures of converting advanced joint model to concept joints are

illustrated in Fig. 1.5.

{ } {

} [ ]{ }

[ ] [ ] [ ] [ ]

[ ] [ ] [ ] [ ]

[

] {

} {

}

[ ] [[

]

[ ]]

Fig. 1.6 the detailed or advanced (left) and the concept model (right) for the front windshield

Table 1.1 the effects of the important panels on the BIW torsional stiffness (W/O denotes “without”)

1.2.3 PANELS MODELING

Although most of the panels have a decisive role on the automotive structural dynamics at high frequencies (>100 Hz), efficacy of the

special panels such as the roof and the platform tunnel in vehicle body cannot be neglected. In other words, BIW structural mode shapes and

resonance frequencies are directly linked to such panels [13]. Table 1.1 investigates the influence of the roof panel, front windshield, and rear

windshield in the first torsion mode of BIW. The concept panels are simply created using shell elements (2D elements) by some principal

nodes of the real panel provided that the shape of the panel is approximately kept. Figure 1.6 depicts the concept panels of the front windshield

of the BIW. Finally the attachment of panels and beams is conducted by interpolation and rigid elements (e.g., RBE2 and RBE3 in NASTRAN

[14]).

II. 2.1 ESTABLISHMENT OF CAE MODELING AND ANALYSIS

Finite element analysis is almost utilized in analyzing the vibratory dynamics of these systems, though the full nonlinearity of the problem

with the transient impacts is not accounted for the analysis. In the above analysis the resonant frequencies and mode shapes from the linear

finite element model are used to extract the local dynamics of the components in terms of a simple multi-degree of freedom spring mass

model. Then, characterization of the seat dynamics is the based method for predicting and controlling of the BSR noise. It showed that the

passenger had serious injuries mainly in the head, chest and legs in the frontal crash. In view of the seriousness of passenger’s injury in vehicle

frontal crash, corresponding statutes related to vehicle frontal crash have been formulated in Asia, America and Europe. Therefore, simulation

analysis of vehicle frontal crash is very significant to improve vehicle crashworthiness in frontal crash (Fig. 2). Vehicle frontal crash model has

large quantities of nodes and units, so it has certain high requirements for software and hardware. This modeling is completed on work station

by using Hyper mesh, which is common preprocessing software in automobile industry. For more accurate and more effective modeling, a LS-

DYNA file [6] may be set up for each crash respectively according to the subsystems of the vehicle, the suffix of the file is key. Subsystems of

the model include: BIW system, four doors and two hoods system, chassis system,

W/O Roof panel

W/O Front

windshield

W/O Rear

windshield Final concept model

Torsion resonant

Frequency 36.21 Hz 39.99 Hz 44.04 Hz 48.82 Hz




Fig. 2a the frontal crash simulation (below)

Fig. 2b the 100 % frontal crash simulation

For computer simulation research of vehicle frontal crash, except that the key objects to be considered in the vehicle simulation are not

totally the same as those in front crash. Besides front longitudinal beams and dash panel, the front tyre, the front door frame, the steering wheel

and the pillar on the side crash side are also the key objects for consideration and research (Fig. 3).

The side impact simulation. The key point to be considered in side crash simulation is the peripheral structure, including side door, door

frame, pillar, roof and floor on the crashed side, passenger seat and so on. The expected objective is that the deformations of the pillar and the

door are small. Therefore, the side impact simulation is equally important for the research of vehicle crashworthiness (Fig. 4). Since crash

accidents endanger human safety, definite and stringent requirements for vehicle crashworthiness are specified in automobile codes or

standards and Snow ingress analysis based on models from

Thermal analysis shown in (fig .5)




Fig. 3 The 40 % ODB crash simulation

Fig. 4a the side impact simulation




Fig. 4b Impact in meshed model (Left), 4c actual and real side impact (Right)

Fig. 5 Snow ingress analysis based on models from thermal analysis

Crash simulation was carried out for this type of vehicle according to Regulations for Protective Design of Passengers in Frontal Crash

(CMVDR294), which is officially promulgated for execution in our country. The model was stricken against a rigid barrier (as shown in Fig.

6) at a speed of 50 km/h. The bode deformation after the crash is as shown in Fig. 7. Through simulation calculation, the total energy variation

tendency of frontal crash model is shown in Fig. 8. After the crash is performed, the general

Fig. 6 The total vehicle frontal crash FEA model configuration




Fig. 7 The vehicle frontal crash simulation result

System deformation energy (internal energy) and dynamic energy are about 47 % respectively, the remaining 6 % of energy are crash

interfacial energy and hourglass energy in value calculation, of which the hourglass energy only covers 2.5 % of the total energy. Therefore, it

can be seen from the figure and the data that this simulation is believable. The response characteristics of the total energy may be used to

evaluate the general crashworthiness of the body structure.

Fig. 8 The system energy curves (below)




Fig. 9 Division of crash energy absorption areas

2.2. ENERGY ABSORPTION AREA OF VEHICLE BODY AND FORCE AND MODELING OF OTHER PARTS

Transmission Route in Frontal Crash At present, the method popularly adopted for vehicle body design is to divide the body into

three areas: front, middle and rear energy absorption areas [7], as shown in Fig. 9. In which, the front energy absorption area is mainly made

up of front bumper beam, bumper beam buffer block, engine hood front end and crash box. Such components are made of high performance

plates, which can absorb the energy produced in the impact as much as possible through their deformation, and also continue to split the

energy to the left and the right by using the force borne by the structure and transfer the energy rearward. The middle energy absorption area is

mainly composed of upper and lowers longitudinal beams of the body, fender, engine hood rear part and auxiliary frame. Such structures

absorb most energy produced in the crash through reasonable bending deformation. The rear energy absorption area is composed mainly of the

driver’s cab that is both strong and rigid. In design of rear energy absorption area, deformation that may cause injury to the passengers must be

avoided as much as possible so as to guarantee the passengers’ safety by reducing the intrusion into the front floor and the crash speed as well

as the acceleration of frontal crash.

Fig. 10 Division of crash force transmission (below)

Fig. 10 Division of crash force transmission




Fig. 11 Effect of base pressure on tail gate side

CRASH FORCE TRANSMISSION ROUTES: The front, middle and rear energy absorption areas are the basis of multiply force routes for

transmission in frontal crash. Setting of multiply force transmission routes in frontal crash for frontal crash may effectively absorb energy and

transmit crash force. Multiply force transmission routes in frontal crash are generally divided into three layers as shown in Fig. 10, which are

typical force transmission routes in the course of frontal crash. Effect of base pressure on tail gate given in fig. 11. It can be seen from the

above figure that the upper layer of frontal crash force transmission routes is composed of such components as engine compartment, upper

longitudinal beams and front damper mounting hood, which absorb some of the crash energy from the front area, and disperse the rest of

energy to Pillar A, front wall and reinforcing beam. The middle layer

Fig. 12 The velocity plane and pressure distribution




Fig. 13 Contribution of the individual parts of the underbody to Cd*A. and The effect of underbody deflectors and body plates and overall

stress distribution reduce dirt deposition in rear rims (below)




Fig. 14a (Left), 14b (Right) wheel mounting and suspension pressure distribution

Fig. 14c, 14d meshing of related parts

No deflector 79 of 1045, 7.5 % No deflector, 6 % of released nr

Deflector 38 of 1045, 3.6 % Deflector 4.3 % of released nr

Fig. 15 hits on rear suspension parts and wheels




Fig. 16 The right longitudinal beams design plan 1 (The above is the original plan

Of the Longitudinal beams)

2.2.1. INFLUENCE OF LONGITUDINAL BEAMS ON THE CRASH

It is known from the above discussion that the longitudinal beams of the body plays a vital role in frontal crash. Since one of the engines

mounted in this type of vehicle has a large volume, the shape of the right longitudinal beams has to be amended according to the proposal of

the general arrangement. The most direct scheme is to reduce Y section in the middle of the right longitudinal beams. Design plan (1) is as

shown in Fig. 16. Through simulation calculation, it is discovered that the middle part of the right longitudinal beams has seriously bent before

the energy absorption box is entirely squashed. As shown in Fig. 17, this type of state is not favorable to energy absorption in the early stage,

which may cause more serious injury to the passenger.

Fig. 17 Plan 1 the right longitudinal beams deformation (below)




Semi-enclosed

Reinforcing part

Fig. 18 The right longitudinal beams design plan 2 (the above is plan 1)

On the basis of Plan 1, Plan 2 is a transition design by using a smooth curve in the variable cross section, which is bound to affect the rigidity

of the middle part of the right longitudinal beams. Therefore, a semi-enclosed reinforcing part is added in the middle part to reinforce the

rigidity at this place, as shown in Fig. 18. Through simulation calculation, the right crash box is squashed completely, but the middle part of

the longitudinal beams is not bent. This state will make the energy transmit directly to the root of

Fig. 19 The right longitudinal beams deformation of plan 2 (below)

Fig. 20 The right longitudinal beams design plan.3 (the above is plan 2)




Fig. 21 The right longitudinal beams deformation of plan 3

The right longitudinal beams and the front floor, and the intrusion of the right fire bulkhead are increased obviously, as shown in Fig. 19. Plan

3 is made on the basis of Plan 2, in which the semi-enclosed reinforcing part in the middle of the right longitudinal beams is cancelled. In order

to reinforce the bending rigidity at this position, two 20 mm stiffeners are added in the weak position in the middle of the right longitudinal

beams, as shown in Fig. 20. Through simulation calculation, the middle of the right longitudinal beams is bent seriously after the crash box at

the front end of the right longitudinal beams is completely squashed. This type of design can ensure the bending rigidity of the right

longitudinal beams and also most energy produced in the crash is consumed in the area in front of the cab. So this type of deformation is

relatively idealistic in frontal crash, as shown in Fig. 21.

III. 3.2. DESIGN VERIFICATION

In order to verify the design feasibility, the design of Plan 3 is used in a real vehicle crash test. Seen from the crash result as shown in Fig.

22, the front energy absorption area and the middle energy absorption area of the vehicle have suffered from more serious deformation, which

absorbs most of the energy. Behind Pillar A, the area deformation is smaller, the intrusion of the fire bulkhead is smaller, the front floor does

not have essential deformation, the door can be opened normally, and the fuel tank does not have any leakage. In this test, only the

crashworthiness of the vehicle body is verified, so no safety airbag is installed in the test vehicle. After the real vehicle crash, two places are

found to have obvious fractures, one is the connecting position of the engine suspension to the engine body, and the other is the lower housing

of the gearbox, where serious cracks occur as shown in Fig. 23. However, the higher simulation result may be because of the different contact

mode algorithm, the higher contact rigidity used as well as one-directional mass point instead of the mass of vehicle body accessories and

interior decorations. As a whole, the fitness of the wave forms and peak values of the simulation curve and the test curve is higher. After the

failure mode of the corresponding

Fig. 22 Comparison between test result and simulation result (below)




Fig. 23 The failure parts after testing

Fracture position is adjusted in the crash model; the simulation calculation is carried out for the second time. An acceleration curve

corresponding to the time course is output and compared with the acceleration signal collected in the test, as shown in Fig. 24. the left side

acceleration peak appears near 45 ms while the right side acceleration peak appears near 60 ms. It is measured that the variations of the two

acceleration curves are basically the same, but there are certain differences between the peak value and the corresponding time when it

appears, the simulation value is a bit higher than the test result. According to the modeling experience, the difference between the test and the

simulation time is caused by the instable factors existing in the model, e.g. negative value of contact energy and contact failure. Similarly,

when floor deformation value is acquired from the front floor of the cab, the size of the deformation may directly affect the leg injury value of

the passenger, so to control the intrusion of this area is utterly important, as shown in Fig. 25. The accelerations of simulation value and test

value are controlled within 2 g. The intrusion difference at corresponding positions is controlled within 4 mm, which are used to verify the

reliability of this simulation and the feasibility of body design. See Table 2 for specific values.

Fig. 24a Comparison between acceleration curves of left B pillars (solid line is

Simulation value, dotted line is test value)




Fig. 24b Comparison between acceleration curves of right B pillars (solid line is

Simulation value, dotted line is test value)

Table 2 Summary of acceleration and intrusion

Fig. 25 the measuring points of floor intrusion (below)

IV. 4. CONCLUSION AND PROSPECT

In this paper, with a Class A vehicle of a certain type as the research object, the situation of vehicle crash safety research and

development both home and abroad has been introduced, the frontal crash FEA model of this type of vehicle has been established, simulation

calculation of the model has been made, frontal crash energy absorption areas have been divided, transmission routes of frontal crash force

have been analyzed. Through simulation calculation of vehicle body right longitudinal beams design plan, the longitudinal beams

B Pillar lower acceleration Left side front floor intrusion (mm) Right side front floor intrusion (mm)

Left Right Point x 1 Point x 4 Point x 8 Point P1 Point P4 Point P6

CAE Value 37.29 40.31 10.11 13.1 13.56 15.16 13.54 12.15

Test Value 36.41 39.56 8.69 9.02 9.63 13.31 11.33 10.01




deformation influence on the whole vehicle in the course of crash has been deeply discussed. The optimum design plan of the right

longitudinal beams has been used for real vehicle frontal crash test to verify the feasibility of this longitudinal beams design and the

reliability of the frontal crash model. A method is presented for the seat concept modeling to predict the dynamic behavior of the seat. In the

early vehicle design phase, without access to any detailed data, it is required to develop a concept CAE model based on the initial design

information and predecessor seat data. In this study, the members of the seat structure are modeled and approximated using beam elements.

Also, for shortening the solution time, the Guyan reduction is used to condense detailed joints properties in their boundary nodes to connect

concept beams.

Dynamic comparison between model and experiment is performed to confirm the validity of the developed NVH concept model. The

value of resonance frequencies and MAC showed that these concept models reflect good enough correlation with the experimental test in

low frequency range (below 60 Hz). Seat structure has a large contribution on the noise generation because of the structural vibration and

generally low frequency forces (below 40 Hz) translate from the road to the seat through the tires and suspensions. Therefore, for BSR noise

reduction, structural mode must be controlled. As a solution, the first seat resonance frequency (31.9 Hz) must be shifted up or its vibration

level be reduced. Three different methods were proposed.

These methods are: increasing the seatback stiffness, sensitivity analysis & using mass damper. As the experimental test shows, these

solutions can effectively reduce seat rattle noise from the structure. A noise measurement technique was used to automatically detect the

potential source regions of BSR on the seat subject to sinusoidal signal excitation with 5 Hz discrepancy between 10 and 80 Hz. After

measuring the sound pressure level, the main source of seat rattle is identified. When the input frequency changed from 25 or 35 Hz to 31

Hz, the first seat structural mode was excited and it caused to produce rattle noise.

The most important characteristics of the identified source mechanisms included relative movements such as impact induced phenomena,

and slip–stick between two parts in the frequency range from 300 to 1,200 Hz. Rattles were found to be more dominant than squeaks. The

research method and the simulation results in this paper have provided certain examples for body longitudinal beams design in future and

certain references for optimization of related frontal crash structures.

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cae of vehicle frontal crash analysis - jetir · 2018. 6. 19. · biw (body-in-white) and seat...

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