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SAE TECHNICALPAPER SERIES

2002-01-0927

Analysis of Brake Caliper Seal-Groove Design

Okon D. Anwana, Hao Cai and H. T. ChangDelphi Automotive Systems

SAE 2002 World CongressDetroit, Michigan

March 4-7, 2002

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2002-01-0927

Analysis of Brake Caliper Seal-Groove Design Okon D. Anwana, Hao Cai and H. T. Chang

Delphi Automotive Systems

Copyright 2002 Society of Automotive Engineers, Inc.

ABSTRACT

It is well known that the design of the seal groove assembly in the brake caliper greatly influences the braking performance. The rubber seal performs the dual function of sealing the piston bore and retracting the caliper piston after a brake apply. However, the seal function is affected by the configuration of the seal groove, as well as the friction at the piston/seal and groove/seal interfaces. The material properties of the rubber seal are also important design parameters. Issues such as fluid displacement, piston retraction, piston sliding force, and brake drag are some of the critical brake performance parameters that must be considered in every caliper seal-groove design. Presently, the brake caliper seal groove design is still based on empirical rules established mainly from past experience and its performance is achieved through prototype testing. Indeed, an analytical model that offers some predictive estimate of the seal groove contributions to the braking performance is needed. This will enhance the optimization of the seal groove design, reduce the need for product prototyping and minimize costs. In this paper, we attempt to identify the critical design parameters in the seal/seal groove assembly and quantify their impact on the brake performance parameters.

1.0 INTRODUCTION

A typical disc brake system is comprised of a caliper housing, piston, seal, and shoe with linings (Fig. 1). During brake application, the apply system generates fluid pressure that pushes the piston towards the shoe with lining, forcing the shoe with lining to rub against the rotor, and generating the braking torque to stop the vehicle.

In addition to the braking output, disc brake engineers face the challenge of meeting the strict performance requirements established by the customer, such as fluid displacement and drag torque. Displacement is the additional volume of brake fluid needed for the caliper to

achieve a certain pressure during brake apply. Higher displacement means longer piston travel and consequently more brake pedal travel and slower response of the caliper.

Fig. 1 Schematic of the components of disc brake system.

Drag is the residual torque on the rotor after brake has been released. High drag not only affects lining life, it also increases fuel consumption and causes energy loss. Sometimes, design changes are needed late in the caliper product development cycle to meet the desired performance requirements (such as to achieve the matching retraction characteristics for the seal with system compliance). In such instance, the more timely and cost effective design option is to modify the seal groove design.

The rubber seal is designed to perform dual functions of sealing the piston bore during brake apply and retracting the caliper piston at brake release. The performance of the seal is influenced by the configuration of the seal groove, the seal material properties, as well as friction at the piston/seal and groove/seal interfaces. Issues such as fluid displacement, piston retraction, piston sliding force, and brake drag are some of the critical brake performance parameters that must be considered in every caliper seal-groove design [1, 2]. The need to increase piston retraction (thereby reducing drag) often conflicts with the requirement to reduce piston travel (i.e. reducing displacement) during brake application. This is one of the major challenges in the seal-groove design. Presently, brake engineers rely on prototype testing and

Shoe & Lining

Rotor

piston

Caliper housingBrake Fluid

Groove Seal

empirical rules for seal groove design [3, 7]. These empirical rules are usually limited to a few design parameters and cannot be reliably applied to predict the seal behaviour in a quantitative sense. Indeed, a quantitative measure that offers some predictive estimate of the seal groove contributions to the braking performance is needed.

There is very limited published information regarding the use of math-based approaches for the seal groove design. A number of studies on the rubber material properties for seal applications have been reported [2, 5, 6], as well as experimental estimates of friction effects in the caliper bore and piston [8]. Analysis of the boot seal using the finite element method has been documented [9]. One attempt at math-based design of the seal groove used the Taguchi approach for statistical evaluation of seal groove design parameters [11]. Additionally, the capabilities of commercial analysis software codes for rubber analysis have been reviewed that include estimates of the load-displacement relationship for a given seal groove design [4]. However, none of these studies went further to establish the critical link between the response of the rubber seal within the prescribed groove boundaries and the expected performance of the seal groove design.

In this paper, we attempt to quantify the critical design parameters in the seal/seal-groove assembly and evaluate their impact on some of the brake performance parameters. We note that the deformation of the seal in the seal groove assembly during brake application is governed by a set of design parameters that can be broadly classified as follows:

Material parameters time-dependent, non-linear material properties that govern the deformation of the rubber seal material, and consequently the braking performance.

Groove geometry geometric parameters that prescribe the kinematic boundaries of the seal deformation.

Friction parameters which defines the resistance to seal sliding at its interfaces with the caliper piston and seal groove.

Surface contact parameters - describes the non-linear contact behaviour of the seal at its interfaces with the seal groove as well as with the piston.

Temperature and environmental parameters time, temperature, and service conditions that affect the seal rubber material behaviour, and consequently the braking performance.

Given the complex interaction of these design parameters in the seal/seal groove assembly during

brake application, it would be impossible to adequately address all the design parameters within the limited pages of this paper. Since the seal groove geometry is most susceptible to design modification of all the design parameters, we take the incremental step of limiting the presentation in this paper to the design variations in the seal groove geometry. Studies of design variations in the seal material and friction will be presented in later publications. The understanding gained from this effort will not only enhance the seal-groove design optimization, but will also reduce product prototype testing and development lead-time. Note that product prototyping requires weeks of laboratory and field tests. The use analysis to examine and/or predict the effect of the seal groove geometry on the brake performance takes only a few hours. Thus the use of an analytical approach to reduce product prototyping can save weeks in product development cycle. .

2.0 SEAL GROOVE ANALYSIS

2.1 SEAL GROOVE ASSEMBLY

A seal groove assembly has three main components - rubber seal, piston and caliper groove. Fig. 2 shows the cross-section of the seal groove assembly components in the undeformed state, in which x denotes the radial direction and y denotes the axial direction.

Fig. 2 Seal groove components in the Undeformed State.

Various configurations of the seal groove exist in the industry, but the representative shape shown in Figure 2 captures the characteristic design features of a caliper seal groove. The characteristic features of the groove configuration include seal groove diameter, front angle, bottom angle, corner break, and groove width (Fig. 3). These features coupled with the piston diameter and seal dimensions uniquely define the seal groove assembly. By design, the seal outer diameter is larger than the groove outer diameter. Hence the rubber seal is squeezed between the groove and the caliper piston

BrakeApplyDirection

SealRubber

Seal GrooveContour

Piston

y

x

when assembled. Rubber squeeze in the seal groove assembly and its deformation during brake apply are critical parameters for evaluating seal performance and piston retraction. We now consider the analytical approach to estimating these parameters.

Fig. 3 Caliper seal-groove design features and dimensions

2.2 PROBLEM DEFINITION -THE SEAL GROOVE ANALYSIS

To study the effect of seal groove geometry on brake performance, we selected a seal of known material and configuration. A nomogram of the seal material properties had been established for a range of temperatures (-400F to 2300F), frequency and load rates [2]. Additionally, appropriate material constitutive relations were determined based on the results of the material tests. From friction tests performed at the rubber seal interface with the piston material surface, Coulomb friction laws were developed both for the dry and lubricated contact conditions. As expected, the coefficient of friction was lower under lubricated conditions than under dry contact conditions for the same loading conditions. Hence lubricated conditions in the seal groove assembly can be simulated analytically by using lower friction coefficients as determined from tests. Details on the friction study will be released in due time. The temperature was set at room temperature (under ambient conditions). By fixing the seal material, seal geometry, the lubricant, piston material and geometry, we implicitly standardize the material, friction, contact and environmental design parameters. Under such considerations, one can make judicious changes in seal groove geometry (i.e. in the geometry parameters) and monitor their effects on the braking performance under room temperature conditions. The modeling approach used for finite element analysis (FEA) of the seal groove assembly is described in the next section.

2.3 FEA MODEL

An axisymmetric FE model of the seal groove assembly components was developed in which the seal groove and piston were considered to be rigid. The rubber seal material was modeled as hyperelastic and viscoelastic. A hyperelastic material exhibits non-linear, large strain, and elastic deformation when loaded. Viscoelasticity describes the time-dependent as well as temperature-dependent changes in material properties under load. These material descriptions are consistent with the mechanical behavior of the seal rubber material observed from tests [2]. In this analysis, the second order Ogden hyperelastic model was used while the material viscoelastic behavior was described using the Prony series. Assuming Coulomb friction law, a static friction coefficient of 1 was used to model the friction contact between the rubber seal and the groove. A lower static friction coefficient of 2 was assumed at the piston/rubber interface to account for the effect of piston lubrication. To capture the pressure loading during a brake application, the fluid space in the caliper piston was modeled as fluid cavity with the hydrostatic pressure as the only degree of freedom. The braking load was applied in the form of pressure. But where analysis convergence was not feasible with pressure loading, the brake apply was implemented as a concentrated force on the piston. For simplicity, an infinitely-rigid caliper housing was assumed, hence the spring-back force from the caliper housing was not considered. Analysis was performed using the Abaqus Standard commercial finite element program i.e. the implicit integration scheme was employed in the analysis.

The analysis proceeded as follows. First, the seal groove assembly process was analyzed. During this process, the rubber seal was squeezed from its undeformed configuration into the prescribed boundaries of the groove and caliper piston (Fig. 4).

Fig. 4 Seal rubber deformation at seal groove assembly (Groove with 10deg front angle, 0.3mm corner break and 59.53mm groove diameter).

Groove InnerDiameter (Di)

GrooveWidth

FrontAngle

BottomAngle

CornerBreak

Groove OuterDiameter (Dg)

In the assembled position, the seal rubber is subjected to normal force Fn, a tangential force Ft, and moment. The magnitude of the normal force (Fn) at the piston/seal interface is a measure of rubber squeeze in the seal assembly. Piston sliding occurs when the tangential force (Ft) is exceeded during brake application.

The second step in the analysis involved the relaxation of the seal rubber in the assembled position assuming linear viscoelasticity. Following this, the brake apply sequence was simulated by applying the fluid pressure at a specified flow rate. Predictions on piston travel as well as rubber deformation during brake apply (Fig. 5) were obtained. Note that the extent to which the rubber seal extrudes into the corner break (at the front end) during brake application depends on the friction between the rubber seal and the piston as well as the front angle and corner break geometry.

Fig. 5 Seal rubber deformation at brake apply (Groove with 10deg front angle, 0.3mm corner break and 59.53mm groove diameter).

Thirdly, the brake release was analyzed through instantaneous release of applied pressure. Estimates of the piston retraction, relative to the assembled position were established. It is worth noting that, after brake release, the rubber retracts but not necessarily to its assembled position. Depending on the effects of friction and groove configuration, slippage between rubber seal and the piston can occur during the apply-and-release processes. By comparing the piston and rubber seal positions before, during, and after brake application, we can estimate the piston travel (or displacement) and retraction, which is a key contributor to brake drag.

Contour plots of the displacement and strain energy distribution in the seal groove assembly are shown in Figs. 6 and 7.

Fig. 6: Displacement Distribution in the seal assembly

Fig. 7 Strain Energy Distribution in the seal assembly

Similar results for rubber seal deformation at brake apply are shown in Figs. 8 and 9.

Fig 8 Rubber seal displacement at brake apply

Seal/Piston Apply End

Fn

Ft

Fig. 9 Strain Energy Distribution in the seal rubber at brake apply

We now examine the effect of design parameters on rubber squeeze, piston travel and retraction.

3.0 EVALUATION OF SEAL GROOVE DESIGN PARAMETERS

In evaluating the effect of changes in groove geometry on caliper seal performance, we specifically concentrated on the geometric features (such as groove front angle, corner break size, groove outer diameter) that most influence piston retraction using a fixed piston diameter and seal dimensions. Incremental changes in the groove configuration were made. Table 3.1 summarizes the range of design variables considered in this study. (The reader is referred to Fig. 3 for illustration of the seal-groove design variables).

Variations in seal groove width were also studied, but the groove width parameter was found to be an insignificant factor in piston retraction for the range of groove designs considered in this study. Prototype models were made and tested for each design configuration analyzed. Some of the test results are summarized in Section 4.0.

Table 3.1 Variations in Seal Groove Configuration (Fixed Housing Caliper Groove Design)

Front Angle (degrees) Corner Break (mm)

Groove Diameter (mm)

0.3 59.53

0.7 59.53

0.3 59.69

6.5

10.0

14.0

0.7 59.69

3.1 SEAL SQUEEZE FORCE MAGNITUDE

For comparative evaluation of the effect of seal groove parameters on rubber squeeze, we use a seal groove with 6.5 degrees front angle, 0.3mm corner break, and 59.53mm outer groove diameter as the base design. Let the rubber squeeze force be Fn for the base groove design. If a design change is made to the base groove design (for example, if the front angle is increased from 6.5 degrees to 10 degrees), the rubber squeeze force changes to a new value Fs. We can use the ratio of these forces as a measure of the change in the squeeze force associated with the design change.

Change in squeeze Force = Fs/Fn

Fig. 10 shows the variation in the seal squeeze with changes in the groove front angle, corner break and groove outer diameter as predicted by analysis.

Fig. 10 Variation in seal rubber squeeze force with changes in seal groove geometry.

In Fig. 10, the magnitude of the squeeze force in the base groove is shown as 100% (point A). If the front angle is changed from 6.5 degrees to 10 degrees (point C), the squeeze force is reduced to 86% level of what it was at the base design. If the front angle is increased to 14 degrees (point G), then seal squeeze force is reduced to 76% level. If the seal groove diameter is increased from 59.53mm to 59.69mm (point E), then the squeeze force is reduced to 80%. Changing both the groove diameter to 59.69mm and the front angle to 10 degrees (point J) reduce the rubber squeeze to 73%. Point L represents a change in both the groove diameter to 59.69mm and front angle to 14 degrees. Thus in moving vertically down from point A, we obtain estimates of the variations in the rubber squeeze force with changes in the groove front angle and/or groove outer diameter. When changes in the groove corner break dimension are included, we see further reduction in the rubber squeeze force as the points A, C, E, G, J, and L move respectively to points B, D, F, H, K and M.

Variations in Seal Squeeze Force with Changes in Groove Geometry Using the 6.5deg-0.3-59.63od Groove as Reference

60

65

70

75

80

85

90

95

100

0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7

Groove Corner Break Size (mm)

Cha

nge

in S

quee

ze F

orce

(%)

6.5deg-59.53od10deg-59.53od14deg-59.53od6.5deg-59.69od10deg-59.69od14deg-59.69od

A

BDF

C

E

GJ

L

HKM

3. 2 SEAL SQUEEZE FORCE DISTRIBUTION AND PISTON RETRACTION

Both the magnitude and distribution of the seal squeeze force vary with changes in the seal groove design. The distribution of this force along the piston/seal interface affects piston retraction. It was determined that as the distribution of the normal force is skewed away from the apply end of the seal groove i.e. away from the air end to the fluid end of the seal, piston retraction is improved (Fig. 11). This is attributable to the fact that the sliding force (Ft) is minimized with skewed distribution of the normal or squeeze force (Fn). Note that friction and rubber seal adhesion to the piston surface contribute to force Ft. As shown in Figure 11, the squeeze force in the groove with 6.5 degrees front angle, 0.3mm corner break size and 59.53mm groove diameter is fairly distributed along the seal width. When the corner break size was increased from 0.3mm to 0.7mm, a skewed distribution of the seal squeeze force was obtained and the piston retraction improved by 44% under the same apply load conditions. The 44% improvement in piston retraction was determined by prototype testing. Similar distribution in seal squeeze force and improvement in piston retraction were obtained when the front angle was increased from 6.5 degrees to 14 degrees. With this understanding, judicious changes to the seal groove geometry can be employed to achieve desired levels of piston retraction.

Fig.11 Squeeze force distribution along the seal/piston interface at seal assembly. (Seal groove with 4.17mm width and 59.53mm groove diameter).

4.0 EVALUATION OF SEAL GROOVE PERFORMANCE PARAMETERS

The major performance issues in caliper seal groove design include displacement and drag. Understanding the impact of rubber deformation, friction effects, piston

travel and retraction are critical for estimating the seal groove performance for a given seal groove design.

PISTON RETRACTION

The retraction of the piston after brake apply is quite vital for estimating drag in the seal caliper design. The ability to predict piston retraction analytically (and implicitly estimate drag and other performance parameters) provides several competitive advantages. Firstly, an estimate of the performance of the caliper seal-groove can be made early in the design when only the design variables are known, i.e. before prototypes are made. Secondly, in system-related performance issues such as drag, optimization of the brake caliper design can be implemented early and more efficiently. Furthermore, it is possible to redesign other components of the brake caliper that would otherwise be impossible to modify late in the product design cycle.

To validate the analysis predictions of retraction, a comparison was made between the piston retraction predicted by analysis and from test results under the same loading conditions. In this effort, we used the seal groove design with 6.5 degrees front angle, 0.3mm corner break, 4.17mm groove width and 59.53mm outer groove diameter as baseline design. Let the piston retraction for baseline groove design under specified loading be Rb. Then, the piston retraction Rc predicted for a given groove design under the same loading can be normalized as

It is worth noting that the piston retraction estimates were made after a few applications of the pressure load cycles. The analytical experience indicated that the piston retraction stabilized after three loadapply cycles. Fig. 12 summarizes the variations in piston retraction as a function of corner break.

In the plot shown in Fig. 12, the solid lines depict the analysis predictions of retraction for a groove with 59.69mm outer diameter, while the corresponding prototype test results are shown in dashed lines. Points A to D represent four configurations of the seal groove as follows:

Point A Seal groove configuration with 59.69mm outer diameter, 0.3mm corner break dimension, 10 degrees front angle.

0 0.2 0.4 0.6 0.8 1 1.2

Seal Width (Normalized)

Seal

Squ

eeze

For

ce (N

) 6.5 deg Front Angle - 0.3mm Corner Break

14 deg Front Angle - 0.3mm Corner Break

6.5 deg Front Angle - 0.7mm CornerBreak

Apply End

b

c

RR

NormalizedtractionPiston = Re

Fig. 12: Variation in Piston Retraction with Seal Groove Corner Break for Groove with 59.69mm outer diameter. (Reference: Groove with 59.53 OD, 0.3mm Corner Break, and 6.5 degrees Front Angle)

Point B Seal groove configuration with 59.69mm outer diameter, 0.3mm corner break dimension, 14 degrees front angle.

Point C Seal groove configuration with 59.69mm outer diameter, 0.7mm corner break dimension, 10 degrees front angle.

Point D Seal groove configuration with 59.69mm outer diameter, 0.7mm corner break dimension, 14 degrees front angle.

When the baseline groove design was changed to the groove configuration at point A, analysis predicted a 20.5% improvement in piston retraction. This prediction compares favourably with the 22.2% increase in retraction obtained from laboratory tests of the prototype model. If the corner break for this design change were 0.7mm instead of 0.3mm (Point C), analysis yielded a 62% improvement in piston retraction, versus 61.1% obtained from prototype tests. If the baseline groove design were changed to a groove configuration in point B, a 66.3% improvement in piston retraction was predicted by analysis. Test results showed a 72.2% actual improvement for this case. Similarly, if this latest design change included a corner break dimension of 0.7mm instead of 0.3mm (Point D), analysis predicted a 100% improvement in piston retraction, versus 94.4% obtained from prototype tests. Essentially, in moving up vertically from point A to B or from point C to D, we see the effect of changing the groove front angle from

10 degrees to 14 degrees on the piston retraction. The transition from point A to C or from point B to D shows the effects of increasing the corner break dimension from 0.3mm to 0.7mm on piston retraction. Overall, the piston retraction improved with increase in corner break dimension and/or groove front angle.

Figure 13 shows the effect of seal groove outer diameter on piston retraction. In this comparison, the corner break dimension is kept constant at 0.3 mm.

Fig.13: Variation in Piston Retraction with Groove Outer Diameter. (Reference: Groove with 59.53 OD, 0.3mm Corner Break, and 6.5 degrees Front Angle)

The lower pair of curves shows the predicted improvements in piston retraction for the 6.5 degrees front angle from analysis (solid curve) and prototype tests (dashed lines). The middle pair and upper pair of curves depict similar results for the 10 degrees and 14 degrees front angles respectively. These plots show that piston retraction is improved when the groove outer diameter is increased from 59.53mm to 59.69mm.

In Fig. 14, we compare the improvements in piston retraction predicted from analysis and laboratory prototype tests as a function of the groove front angle. To highlight and differentiate between these curves, the piston retraction is normalized in the form

where Rb is the piston retraction in the baseline groove design and Rc is the piston retraction for the current groove design under the same loading conditions.

Again, these results show improvements in piston retraction with increase in front angle. For grooves with 0.7mm corner break and front angles ranging from 10 degrees to 14 degrees, analysis predicted a higher rate

0

0.5

1

1.5

2

2.5

0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7

Groove Corner Break

Pist

on R

etra

ctio

n N

orm

aliz

ed

Analysis Result: Groove with 10deg Front AngleTest Result: Groove with 10deg Front AngleAnalysis Result: Groove with 14 deg Front angleTest Results: Groove with 14 deg Front Angle

D

B

A

C

1 Re =b

c

RR

Normalizedtraction

Variation in Piston Retraction with Seal Groove Diameter Ref: 6.5deg Front Angle, 0.3mm Corner Break, 59.53OD Diameter

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

59.52 59.54 59.56 59.58 59.6 59.62 59.64 59.66 59.68 59.7

Groove Outside Diameter (mm)

Pist

on R

etra

ctio

n N

orm

aliz

ed

Analysis Result - 0.3 Corner Brake, 6.5deg Front Angle Test Result - 0.3 Corner Break, 6.5deg Front Angle Analysis Results - 0.3 Corner Break, 14deg Front Angle Test Result - 0.3 Corner Break, 14deg Front Angle Analysis Result - 0.3 Corner Break and 10deg Front Angle Test Result - 0.3 Corner Break and 10deg Front Angle

of increase in piston retractions compared to results for similar grooves with front angles less than 10 degrees. However, test data from prototype models of the said groove configurations indicated a somewhat monotonic increase in retraction as the front angle was increased from 6.5 degrees to 14 degrees.

Fig. 14 Effects of front angle on piston retraction. (Seal groove with 59.69mm outer diameter).

From the results summarized in figures 12 to 14, it is evident that the basic trends in piston retraction for various groove configurations have been reproduced analytically. However, some discrepancies exist between piston retraction values predicted by analysis and comparable test results. We attribute the discrepancies in analysis predictions to two main sources. The first source of numerical error is from the limitations in the material hyperelastic and viscoelastic models. These material models were used to describe the seal material mechanical behaviour during brake application. Note that the coefficients for these material models were derived from the standard process of curve fitting the material test data. Clearly, approximations in the curve fitting process will affect the accuracy of the numerical predictions.

The other source of analysis error comes from the Coulomb friction model used in the analysis. In this friction model, assumptions were made regarding friction contacts at the piston/seal and groove/seal interfaces. For example, the values of friction coefficients were assumed to be constant during the analysis of the braking load sequence. Indeed, the nature of the frictional forces at the seal interfaces with the piston and the groove is presently not well understood. But preliminary results of the laboratory study suggest that the Coulomb friction model used in analysis may not be fully adequate for the range of seal groove configurations evaluated in this study. The actual state

of friction appears to be dynamic during the brake application process [2]. As noted earlier, details of the friction study will be released in due time.

There is ongoing effort to refine and expand the analytical model for improved estimate of the seal groove contributions to the braking performance.

5.0 CONCLUSIONS

In this paper, the authors have demonstrated the application of analysis in the caliper seal groove design. In this effort, the mechanics of rubber deformation within the prescribed boundaries of the seal groove were established. Predictions on the critical performance parameters of the seal groove design were made in terms of the geometric design parameters. The basic trends in piston retraction for various fixed-housing groove configurations have been reproduced analytically. Issues such as seal geometry, seal material creep, temperature, and time of exposure also affect seal deformation, and consequently the braking performance. These design parameters will be considered in the next stage of the analytical development.

Indeed, understanding the functional relationship between design and performance parameters is crucial for optimal seal groove design. Despite the assumptions made regarding some critical parameters such as friction, the results indicate the analytical approach developed herein can reasonably predict the seal retraction behavior for a given set or space of design parameters. The ability to predict piston retraction analytically (and implicitly estimate drag and other performance parameters) provides several competitive advantages, including the ability to optimize the performance of the caliper groove seal early in the design when only the design variables are known, i.e. before prototypes are made. Furthermore, with the understanding of retraction mechanism in the seal-groove, it is possible to modify other components of the brake caliper system that would otherwise be impossible to redesign late in the product design cycle. At this stage of development, the authors feel confident that the understanding gained from this effort will not only enhance the seal-groove design optimization, but will also reduce prototype testing and can shorten product development lead time by several weeks.

6.0 ACKNOWLEDGMENTS

The support of Brent K. Dunlap, the Analysis Manager at the Dayton Tech Center, Delphi Automotive Systems, is gratefully acknowledged. Our Special thanks also to the lab technicians in the wheel/brake caliper group at Dayton Tech Center, Delphi Automotive Systems, for providing the supporting test results.

6 7 8 9 10 11 12 13 14

Seal Groove Front Angle (deg)

Ret

ract

ion

Nor

mal

ized

Analysis Results - 0.3mm Corner BreakTest Results - 0.3mm Corner BreakAnalysis Results - 0.7mm Corner BreakTest Results - 0.7mm Corner Break

7.0 REFERENCES

1. Anwana, O. D. Analytical Design of the Seal Groove Assembly Case Study in the Performance Prediction, Delphi Automotive Internal Report (2001).

2. Anwana, O. D., Characterization of Rubber Material for Disk Brake Piston Seal, Delphi Automotive Systems Internal Report EWR--584-022 (2000).

3. Baptists, T. Brake Drag Torque Measurements, Delco Moraine - GM Report No. PG053026 (1988).

4. Chang, H., On a Numerical Study for Rubber Seals, SAE Transactions v.97, Paper No. SAE-880255 (1988).

5. Dinzgurg, B., Measurement of Rubber Elasticity and Correlation to Seal Life, SAE Congress, Detroit, MI, Paper No. SAE-970547 (1997).

6. Dinzgurg, B., The Selection of Elastomer Compounds through Correlation of Rubber Properties to Seal Life, SAE Congress, Detroit, MI, Paper No. SAE-2001-01-0686 (2001).

7. Hrbek, D., Piston Seal Characteristics on a Disc Brake Caliper with a Self Adjusting Park Brake

Mechanism, CPC, GM Report No. PG057117 (1991).

8. Lim, J., Brake Caliper Analysis of Piston Friction, CPE, GM Report No. PG051348 (1987).

9. Meada, N. and Matsuno, M., Analysis of Rubber Boot Seal using Finite Element Method, SAE Congress, Detroit, MI, Paper No. SAE-940289 (1994).

10. Moore, D., The Friction and Lubrication of Elastomers, Pergamon Press, 1973.

11. Shnaider, A., Authentic Involvement Design, Dept. of Mechanical Engineering, Monash University, MI, Final Report to PBR Automotive Ltd. (1996).

8.0 CONTACT

For further information, please contact Okon D. Anwana, Engineering Technical Center, M/C E-520, 1435 Cincinnati Street, Dayton, OH 45408. Telephone: 937-455-5864, Fax: 937-455-6798