2012-36-0195

Optimization of new plastic bracket NVH characteristics using CAE

Reinaldo dos Santos

Ford Motor Company

Masoud Saadat Ford Motor Company

Santosh Neriya Ford Motor Company

David Popejoy Ford Motor Company

Valter E. Beal SENAI CIMATEC

Copyright © 2012 SAE International

ABSTRACT

NVH requirements are critical in new driveline developments.

Failure modes due to resonances must be carefully analyzed and potential root causes must have adequate

countermeasures. One of the most common root causes is the

modal alignment. This work shows the steps to design and

optimize a new plastic bracket for an automotive half shaft

bearing. This bracket replaces a very stiff bracket, made of

cast iron. The initial design of plastic bracket was not stiff

enough to bring natural frequency of the system above engine

second order excitation at maximum speed. The complete

power pack was modeled and NVH CAE analysis was

performed. The CAE outputs included Driving Point

Response, Frequency Response Function and Modal analysis.

The boundary conditions were discussed deep in detail to make sure the models represented actual system. After some

iteration, weaker areas were identified and the design was

changed, increasing stiffness and shifting some low frequency

modes to higher frequencies. The remaining mode below

engine second order could not be changed adequately, so a

different strategy needed to be taken. An elastomeric isolator

was added between bearing and bracket, in order to dampen

the vibrations. The material chosen was EPDM, due to its

damping coefficient and high temperature resistance. The

model was submitted to a new analysis, when the stiffness of

the isolator could be determined in order to match the resonant frequency. This isolator reduced the transmissibility of the

vibration through bracket and the amplitude of the vibration

was decreased to an acceptable level with this strategy.

INTRODUCTION

There is a big demand for automobiles with smaller fuel

consumption and emissions levels. The North American

Legislation CAFE (Corporate Average Fuel Economy) has rigid targets for fuel consumption and emissions. It

determines that vehicles, which have a fuel consumption 28.8

miles per gallon in 2010, must present a fuel consumption as

low as 34.1 miles per gallon until 2016 (CHEAH et al., 2010).

There are different strategies to meet these targets.

Alternatives on how to improve the efficiency of powertrains,

use alternative energy sources, improve aerodynamics and

reduce size and mass of vehicles have been investigated in

many researches. Simulations show that a mass reduction of

10% can bring a fuel consumption reduction of 6.7% in

passenger vehicles and 7.6% in pickups in North America

(HEYWOOD, 2010).

The mass reduction can be achieved through a combination of

material substitution, redesign of vehicles, and components

and also its size reduction. The conventional materials used in

vehicle construction can be replaced, mainly, by high strength

steel, aluminum, composites and thermoplastics (CHEAH et

al., 2007).

The usage of thermoplastics in automotive applications is

currently very common. Most of modern vehicles have around

100 – 150kg of plastics per unit (MARK, 2004). However,

most of these components are used as trim parts. With the

development of new engineering plastics, more resistant to high temperatures, chemical attack and mechanical

solicitations, its usage as structural components have been

increasing. One of the most used polymers in this kind of

application is the Polyamide. The addition of fillers and

additives made the polyamide a good substitute for metallic

parts in many applications.

Even though polymers have many advantages, the working

conditions in powertrains are very severe for this kind of

material. High temperatures, high intensity and cyclic loads,

and vibration excitations are examples of usual conditions in

powertrains regular usage. Polymers can degrade under high

temperatures and become fragile under low temperatures.

They are less resistant to peak efforts and fatigue than most

metals and their fatigue mechanisms are more difficult to predict. The stiffness of polymers is also much reduced

comparing to metals.

Some of the potential failure modes of these applications are

related to NVH. Vibration excitations can be amplified due to

resonances. This can degrade the occupant comfort and in the

extreme case can lead to a catastrophic structural failure.

These factors must be taken in account when developing a

polymeric component, in order to avoid undesired results.

This work describes the development of a powertrain bracket

in polyamide. The function of the part is to hold a bearing of

half shaft system. Currently this part is made of cast iron and

its mass is approximately 1.4kg. The current part is very stiff and resistant to the applied loads. A holder (steel strap)

attached to the bracket holds the bearing that supports the

linkshaft. This assembly is bolted to the engine block. The

proposed part is made of PA 6.6 GF (Polyamide 6.6 glass fiber

reinforced). It has a plastic body with an integrated holder,

closed with a bolt and a nut. Between the holder and the

bearing there is an EPDM (Ethylene Propylene Diene

Monomer) ring. The function of this ring is to isolate

vibrations and compensate thermal expansion. The ring in

contact with the bracket is made of the same material (PA 6.6

GF). The ring attached to the bearing is made of steel. The intermediate layer is made of EPDM. This material was

chosen due to its temperature resistance and damping factor.

TARGET SETTING

Once the concept was defined, the durability, NVH and

temperature targets were set. Concerning durability, the RLD

(road load data) was acquired in the routes of the vehicle

durability tests. To measure the actuating forces in the bracket

accurately a device was built. It replaces the bracket and

acquires data through a load cell. Another important

measurement taken was the working and peak temperatures.

These temperatures vary according usage conditions, but for

simulation purposes the values considered as the worst

conditions were:

• Working temperature: 100oC

• Peak temperature: 140oC

The natural frequency target of the system was set as 260Hz

minimum. This definition considered the maximum engine

2nd. order excitation, with a 30% safety coefficient. The part

must be stiff enough to keep the natural frequency of the

system above this target, avoiding resonances. This target is

easily achieved with the current bracket. For the plastic

bracket, due to its lower stiffness, a different strategy was

adopted. The noise factor (resonance due to low stiffness) was

compensated with the tuning of the EPDM ring. Its stiffness

was adjusted to the critical frequencies, avoiding resonance

effects.

OPTIMIZATION WITH CAE

Once the targets were set, the design was submitted to a

fatigue simulation. A key assumption was made that the

bracket material was homogeneous and isotropic, due to

difficulties to predict the fiber orientation (due to mold

filling), as well as limitations of the fatigue software. The

simulation calculated the damage for one force unit (1N) in

each direction (x, y, z). Then, accumulated damage for all the

loads applied along durability tests, scaling each load case

according to the measured road load data at each time step, is

calculated. The first results showed a poor fatigue life for the part due to stress concentration in some areas, as shown in red

in the figure 1.

Figure 1 – Fatigue CAE results

The design was improved using topology optimization in the

software Optistruct. The analysis showed the areas with higher

stress concentration, where the part should be reinforced (fig.

2).

Figure 2 – Regions to be improved shown in the topological

study

Some ribs were added in the weaker areas. The fatigue life

was increased, but there were still areas with low fatigue life.

The results can be seen in the figure 3.

Figure 3 – Fatigue CAE results for the part with added ribs

After this work, the design was submitted to a NVH CAE in

order to understand its behavior in terms of natural

frequencies. The whole powertrain was modeled and the main

components were checked for resonances. Multiple iterations

were performed to determine the modes as well as the

response of the linkshaft bracket under unit load excitation.

The Driving Point Response analysis was also performed,

where the bracket response is measured at the excitation

location (fig. 4). Different materials are considered before identifying the optimal material for the linkshaft bracket. A

preliminary simulation was done, considering extreme and

intermediate combinations of material resistance and

temperature, to help understanding the system. The material

characteristics were taken from Rhodia Technyl Product

Datasheets (RHODIA, 2010). The combinations are listed

below:

Material working temperature

• PA 6.6 40% GF (A 218 V40) 140oC

• PA 6.6 50% GF (A118 LV50) 100oC

• PA 6.6 60% GF (AFX 218 V60) 23oC

The characteristics of the materials considered for the simulations are listed in the table 1.

Figure 4 - System FE model and the Driving Point Response

locations

The work was intended to characterize the effect of both the

bracket and EPDM ring materials on the natural frequency of

the system, using the same bracket geometry. At first, the

EPDM ring was considered as being made of the same material as the bracket (PA.6.6 GF).

Figure 5 – Comparison of resonant modes with different

materials

Table 1 - Different component properties used for the FE model

Material

Temperature 23oC 100

oC 140

oC 23

oC 100

oC 140

oC 23

oC 100

oC 140

oC

Tensile modulus [Mpa] 11400 6207 5165 13000 7474 6221 16000 9700 8000

Tensile Strength at break [Mpa] 189.3 103 89 203.6 111 94 195 106 90

Tensile Elongation at break 4 6.75 8.76 4 5.99 6.86 3.1 4.5 5.2

Poisson’s ratio 0.35 0.35 0.35 0.35 0.35 0.35 0.35 0.35 0.35

Density 1.36 1.25 1.16 1.38 1.31 1.26 1.69 1.55 1.45

Flexural modulus [Mpa] 12230 6049 5352 16320 7644 7511 13800 6400 6200

A 218 V40 A 118 LV50 AFX 218 V60

This simulation showed a small effect in the natural

frequencies. The modes and frequencies were similar for the 3

materials (fig. 5). As a reference, similar analysis was

performed considering the steel as the material.

The Driving Point Response (DPR) analysis was performed by

applying unit excitation in X, Y, and Z and measuring the acceleration in X, Y, and Z respectively. This analysis showed

the resonant modes and their frequencies. The graphs show

peaks around 70Hz, below the target of 260Hz (fig. 6).

Figure 6– DPR analysis results

The areas with higher displacement and stress concentration

(Von Mises) were identified (fig. 7). This helped to reinforce

the part in the weaker areas

Fi

Figure 7 – Displacement (a) and Stress (b) plots

The part was reinforced with some ribs, showing small

improvement. The part was then redesigned in order to

eliminate some low frequency modes and increase the natural

frequency of the remaining ones sitting below the target. The

region of the holder was reinforced with some triangle shaped

ribs and the base of the bracket was made larger, using the

space available The part was submitted again to fatigue CAE

and after improvements in the regions with high stress concentration, the analysis showed a better stress distribution,

with fatigue life in the worst stress concentration areas above

target (higher than 5.4 fatigue life cycles). The figures 8 and 9

show the stress distribution in the final version.

Figure 8 – Fatigue CAE final result

Figure 9 - Fatigue CAE final result

The new design was submitted to a new NVH simulation. The

modes in this version were shifted to frequencies around 100

and 120 Hz. The part was reinforced with new ribs and

analyzed again. The new results showed a reduction in the

amplitudes, with negligible changes in the frequencies. Since

the changes in the design haven’t shown the needed results, a

different strategy was adopted. Instead of making the design

more robust, it was changed to compensate the noise factor

related to stiffness. The EPDM ring was then tuned. Its

stiffness was set to in a way its oscilation could dampen the mode with higher influences in the system. For the resonant

mode at 120Hz the stiffness were determined through CAE.

Given the EPDM material and damping characteristics, the

dynamic stiffness curves of the EPDM, as shown in fig. 10

and 11, are determined as follows:

• The outer layer of the EPDM ring is fixed and unit

load is applied at the center of the EPDM ring in

different directions.

• Displacement of the application load is measured and

inverted to get the dynamic stiffness of the EPDM

from 0-500 Hz. Since the response of the analysis is the displacement of the part for the applied load and

by definition stiffness is defined as

Force/Displacement, in this case one unit force, the

stiffness of the rubber would be the unit divided by

Displacement.

The values from the curves at the resonant frequency of 120

Hz, calculated from the CAE, are shown below and provided

to the supplier, for the design of EPDM ring.

• Radial Stiffness (Y & Z-Dir) = 2065 N/mm @ 120 Hz

• Axial Stiffness (X-Dir) = 1052 N/mm @ 120 Hz

Figure 10 - Radial stiffness of the EPDM ring

Figure 11 - Axial stiffness of the EPDM ring

With radial and axial Stiffness shown curves it was possible to

design a ring to achieve the design targets. The figure 12

shows a comparison of the Driving Point Responses of the

first proposal (baseline), the 1st iteration and the final design.

It shows the differences in terms of amplitude of vibration and

frequency for the three designs after bracket changes and

tuning the EPDM isolator. The proposed design with the presence of EPDM isolator shifts the response peaks between

250 and 300 Hz to a frequency above 300 Hz. It also reduces

the amplitude of the responses around 120 Hz.

Figure 12 – DPR curves for initial, 1st. iteration and final

bracket proposals

The figure 13 shows the modal analysis of the initial

(baseline) proposal and final one. The reduction of amplitudes

can be seen in the images. The modes in 103.5 and 137.5Hz

had smaller amplitudes and have not shown big influence in

the system response.

Figure 13 - Modal Analysis Baseline (a) vs. New Proposed

Design (b)

SUMMARY/CONCLUSIONS

This study has shown a way to design a new plastic part

replacing an existing one made of cast iron. The optimization

process was described in order to provide a better

understanding of how the available tools can be used to

achieve useful results for new applications. As expected, the

plastic part design was not able to achieve natural frequency

targets due its smaller stiffness. The package limitations also contributed to limit the improvement of the component.

According to the simulation results, the strategy of

compensate noise factor, tuning the EPDM isolator to the

frequencies below the target was efficient to reduce vibration

amplitudes. The next step of this development are building

prototypes of the plastic bracket and running a DOE (Design

of Experiments) to confirm EPDM ring tuning. This DOE

would consist in a series of physical modal analysis

measurements with rings with different stiffness. This can be

used to determine statistically the optimum value for EPDM

ring stiffness to dampen the resonant frequencies.

REFERENCES

1. CHEAH, Lynette et al. Factor of Two: Halving the Fuel Consumption of New U.S. Automobiles by 2035 Cambridge: Laboratory for Energy and Environment

Massachusetts Institute of Technology, 2007.

2. CHEAH, Lynette et al. Meeting U.S. passenger vehicle fuel economy standards in 2016 and beyond. Burlington:

Elsevier, 2010

3. HEYWOOD, John B. Assessing the Fuel Consumption and GHG of Future In-Use Vehicles PEA-AIT International Conference on Energy and Sustainable Development: Issues and Strategies (ESD 2010) The

Empress Hotel, Chiang Mai, Thailand. 2-4 June 2010.

4. MARK, Herman F. Encyclopedia of Polymer Science & Technology. 3rd ed. Hoboken: John Wiley & Sons, Inc.,

2004.

5. RHODIA, Relatório Técnico FS: 2010-120, São Bernardo

do Campo: Rhodia, 2010.

6. RHODIA, TECHNYL A 118L V50 - A FT 051 - FICHA TÉCNICA - VERSÃO 01, São Bernardo do Campo:

Rhodia, 2000.

7. RHODIA, TECHNYL A 218 V40 Product Datasheet – A FT 110- 2010, São Bernardo do Campo: Rhodia, 2010.

CONTACT INFORMATION

Reinaldo dos Santos

rsant138@ford.com

Av. Henry Ford, 2000

42810-225 - Camaçari - BA - Brazil

Masoud Saadat

msaadat1@ford.com

Powertrain NVH Research & Development

Advanced Engineering Center, Ford Motor Co.

2400 Village Rd, Dearborn, MI 48124 - USA

Santosh Neriya

sneriya@ford.com

Powertrain NVH Research & Development

Advanced Engineering Center, Ford Motor Co.

2400 Village Rd, Dearborn, MI 48124 - USA

David Popejoy

dpopejoy@ford.com

ATNPC, Ford Motor Company

35500 Plymouth Rd., MD 246

Livonia, MI 48150 - USA

Valter E. Beal

valter.beal@fieb.org.br

SENAI CIMATEC

Av. Orlando Gomes, 1845

Salvador – BA - Brazil

ACKNOWLEDGMENTS

The authors gratefully acknowledge Alexandre Morbeck and

Rhodia for the support with technical information of

polyamides; Roberto Morinaga, for the support with NVH

knowledge and incentive and Bin Juang and Jershi Chen for

the CAE resources provided for this work.

DEFINITIONS/ABBREVIATIONS

CAD - Computer Aided Design

CAE - Computer Aided Engineering

CAFE - Corporate Average Fuel Economy

DOE - Design of Experiments

DPR - Driving Point Response

EPDM - Ethylene propylene Diene monomer

FE – Finite element

NVH – Noise, Vibration and Harshness

PA 6.6 - Polyamide 6.6

PA 6.6 GF - Polyamide 6.6 glass fiber reinforced

RLD - Road load data