Journal of Mechanical Engineering 2013 6

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59 (2013) 6

Strojniški vestnik Journal of Mechanical Engineering

Since 1955

Contents

Papers

358

Tomaž Petrun, Jože Flašker, Marko Kegl: A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations

367

Pingfa Feng, Chenglong Zhang, Zhijun Wu, Jianfu Zhang: Effect of Scratch Velocity on Deformation Features of C-plane Sapphire during Nanoscratching

375

Sedat Karabay, Kasım Baynal, Cengiz İğdeli: Detecting Groan Sources in Drum Brakes of Commercial Vehicles by TVA-FMEA: A Case Study

387

Peng Gao, Shaoze Yan, Liyang Xie, Jianing Wu: Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths

400

Tomaž Berlec, Janez Kušar, Lidija Rihar, Marko Starbek: Selecting the Most Adaptable Work Equipment

409

Iztok Palčič, Marc Pons, Andrea Bikfalvi, Josep Llach, Borut Buchmeister: Analysing Energy and Material Saving Technologies’ Adoption and Adopters

418

Yilmaz Kucuk: Modeling Nonlinear Viscoelastic Nanoindentation of PVAc at Different Unloading Rates

Journal of Mechanical Engineering - Strojniški vestnik

Vytautas Ostasevicius, Rimvydas Gaidys, Rolanas Dauksevicius,

Sandra Mikuckyte: 351 Study of Vibration Milling for Improving Surface Finish of Difficult-to-Cut Materials

6 year 2013 volume 59 no.

Strojniški vestnik – Journal of Mechanical Engineering (SV-JME) Aim and Scope The international journal publishes original and (mini)review articles covering the concepts of materials science, mechanics, kinematics, thermodynamics, energy and environment, mechatronics and robotics, fluid mechanics, tribology, cybernetics, industrial engineering and structural analysis.

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Editor in Chief Vincenc Butala University of Ljubljana Faculty of Mechanical Engineering, Slovenia Technical Editor Pika Škraba University of Ljubljana Faculty of Mechanical Engineering, Slovenia Editorial Office University of Ljubljana (UL) Faculty of Mechanical Engineering SV-JME Aškerčeva 6, SI-1000 Ljubljana, Slovenia Phone: 386-(0)1-4771 137 Fax: 386-(0)1-2518 567 E-mail: info@sv-jme.eu, http://www.sv-jme.eu Print DZS, printed in 450 copies Founders and Publishers University of Ljubljana (UL) Faculty of Mechanical Engineering, Slovenia University of Maribor (UM) Faculty of Mechanical Engineering, Slovenia Association of Mechanical Engineers of Slovenia

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59 (2013) 6

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Strojniški vestnik Journal of Mechanical Engineering

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Journal of Mechanical Engineering - Strojniški vestnik

mož Mrvar, Jožef Medved, Janez Grum: alysis of Laser Coating Ceramic Components TiB2 and TiC y EN AW-6082-T651

no. 6 2013 volume 59

year

Cover: Evolution in programming software made robots suitable not only for pick and place operations, but for multiaxis machining as well. With reasonably appointed milling limitations robots can be used in every branch of industry. Some materials (steel, wood, foam and plastics) which were machined for research purposes in laboratories are presented in the cover.

Image Courtesy: Laboratory of Computer Aided Design (LECAD), Laboratory for cutting (LABOD), Faculty of Mechanical Engineering, University of Ljubljana

International Editorial Board Koshi Adachi, Graduate School of Engineering,Tohoku University, Japan Bikramjit Basu, Indian Institute of Technology, Kanpur, India Anton Bergant, Litostroj Power, Slovenia Franci Čuš, UM, Faculty of Mech. Engineering, Slovenia Narendra B. Dahotre, University of Tennessee, Knoxville, USA Matija Fajdiga, UL, Faculty of Mech. Engineering, Slovenia Imre Felde, Obuda University, Faculty of Informatics, Hungary Jože Flašker, UM, Faculty of Mech. Engineering, Slovenia Bernard Franković, Faculty of Engineering Rijeka, Croatia Janez Grum, UL, Faculty of Mech. Engineering, Slovenia Imre Horvath, Delft University of Technology, Netherlands Julius Kaplunov, Brunel University, West London, UK Milan Kljajin, J.J. Strossmayer University of Osijek, Croatia Janez Kopač, UL, Faculty of Mech. Engineering, Slovenia Franc Kosel, UL, Faculty of Mech. Engineering, Slovenia Thomas Lübben, University of Bremen, Germany Janez Možina, UL, Faculty of Mech. Engineering, Slovenia Miroslav Plančak, University of Novi Sad, Serbia Brian Prasad, California Institute of Technology, Pasadena, USA Bernd Sauer, University of Kaiserlautern, Germany Brane Širok, UL, Faculty of Mech. Engineering, Slovenia Leopold Škerget, UM, Faculty of Mech. Engineering, Slovenia George E. Totten, Portland State University, USA Nikos C. Tsourveloudis, Technical University of Crete, Greece Toma Udiljak, University of Zagreb, Croatia Arkady Voloshin, Lehigh University, Bethlehem, USA President of Publishing Council Jože Duhovnik UL, Faculty of Mechanical Engineering, Slovenia General information Strojniški vestnik – Journal of Mechanical Engineering is published in 11 issues per year (July and August is a double issue).

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6 Contents

Contents Strojniški vestnik - Journal of Mechanical Engineering volume 59, (2013), number 6 Ljubljana, June 2013 ISSN 0039-2480 Published monthly

Papers Vytautas Ostasevicius, Rimvydas Gaidys, Rolanas Dauksevicius, Sandra Mikuckyte: Study of Vibration Milling for Improving Surface Finish of Difficult-to-Cut Materials Tomaž Petrun, Jože Flašker, Marko Kegl: A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations Pingfa Feng, Chenglong Zhang, Zhijun Wu, Jianfu Zhang: Effect of Scratch Velocity on Deformation Features of C-plane Sapphire during Nanoscratching Sedat Karabay, Kasım Baynal, Cengiz İğdeli: Detecting Groan Sources in Drum Brakes of Commercial Vehicles by TVA-FMEA: A Case Study Peng Gao, Shaoze Yan, Liyang Xie, Jianing Wu: Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths Tomaž Berlec, Janez Kušar, Lidija Rihar, Marko Starbek: Selecting the Most Adaptable Work Equipment Iztok Palčič, Marc Pons, Andrea Bikfalvi, Josep Llach, Borut Buchmeister: Analysing Energy and Material Saving Technologies’ Adoption and Adopters Yilmaz Kucuk: Modeling Nonlinear Viscoelastic Nanoindentation of PVAc at Different Unloading Rates

351 358 367 375 387 400 409 418

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 351-357 © 2013 Journal of Mechanical Engineering. All rights reserved.

DOI:10.5545/sv-jme.2012.856

Original Scientific Paper

Received for review: 2012-11-08 Received revised form: 2013-03-05 Accepted for publication: 2013-03-15

Study of Vibration Milling for Improving Surface Finish of Difficult-to-Cut Materials Ostasevicius, V. – Gaidys, R. – Dauksevicius, R. – Mikuckyte, S. Vytautas Ostasevicius – Rimvydas Gaidys – Rolanas Dauksevicius* – Sandra Mikuckyte 1 Kaunas

University of Technology, Faculty of Mechanical Engineering and Mechatronics, Institute for Hi-Tech Development, Lithuania

This work studies the influence of high-frequency excitation of a cutting tool during end milling of workpieces made of difficult-to-cut metallic alloys. It is demonstrated that high-frequency vibrations superimposed onto the continuous movement of the tool lead to milling process stabilization with superior surface finish in comparison to conventional machining. A finite element model of the vibration milling tool was built and verified experimentally. The model treats the tool as an elastic pre-twisted structure (mill cutter) characterised by its natural vibration modes. The resonance frequencies of the axial vibration mode of cutters of two different lengths were predicted numerically and subsequently used for excitation of the vibration milling tool during cutting experiments. Qualitative and quantitative characterization of the surface quality of the machined stainless steel and titanium alloys was performed. Measurement results have confirmed that excitation of a specific tool mode is a prerequisite for achieving maximal efficiency of the vibration milling process. Statistical analysis of the collected roughness measurement data identified factors that most significantly contribute to the improved surface finish of the workpieces.

Keywords: vibration cutting, finite element model, pre-twisted cantilever, axial mode, roughness

0 INTRODUCTION Controlling vibration phenomena in production machines is one of the approaches for improving their efficiency. This also applies to cutting tool vibrations generated during machining, when the magnitude of the vibrations directly influences workpiece surface quality. Continuous efforts to enhance cutting performance have revealed that machining quality may be improved if a tool is assisted with high-frequency vibrations. During the resulting vibration cutting process [1], the tool periodically loses contact with the chip leading to a reduction in machining forces, friction, and temperature in the cutting zone and the formation of thinner chips, as well as simultaneously preventing generation of micro-cracks on the cutting edge and workpiece surface. As a consequence, this improves cutting stability, surface finish, and tool life when compared to conventional machining [2].

The authors of this paper previously reported on an approach for the reduction of workpiece surface roughness by exciting higher-order transverse modes in the vibration turning tool [3]. Twardowski et al. [4] analysed various factors affecting surface roughness after end milling of hardened steel in high-speed milling. The work also involved analysis of surface profile charts in terms of vibrations and cutting force components. Chen et al. [5] studied the effect of surface roughness of end mills on cutting performance in the case of high-speed machining. A novel optimization design approach to the processing parameters for high-speed cutting has been proposed. Taylan et al.

[6] considered the wear performance of CBN and

TiN-coated CBN cutting tools during face milling of 61 HRC hardened 90MnCrV8 tool steel workpieces.

Kubiak et al. [7] evaluated the effect of topographical parameters on wettability and spreading phenomenon by using statistical covariance analysis. Anisotropic surfaces were prepared by abrasive polishing on aluminum, titanium, copper alloys, and steel. Imani and Moosavi [8] have proposed a new force model of torsional-axial and transverse vibration for drilling, which was validated experimentally. Simulation of torsional-axial vibrations is accomplished using Bayly‘s model, which is based on the fact that a twist drill lengthens when it »untwists«.

A number of studies have been conducted to monitor vibrations during the milling operation. The results of Zhang and Chen [9] indicate the feasibility of using cutting vibration amplitudes and frequency peaks in two directions, X and Y, for the monitoring of tool condition during end milling operations.

Seguy et al. [10] examined the relationship between chatter instability and surface roughness evolution for thin wall milling. An explicit numerical model was developed for analysis of modal interactions. It takes into account the coupling mode, the modal shape, and the fact that a tool may have a ploughing effect.

Grazeviciute et al. [11] presented measurement results for surface roughness and chatter frequency in the case of the vibration milling of aluminum AlMgSi0.5 (6060) and plastic PA6. The results indicate that the surface of he aluminum alloy after vibration-assisted milling is smoother than after conventional milling. However, no difference was detected in the case of milling plastic material.

*Corr. Author’s Address: Kaunas University of Technology, Studentu 65, Kaunas, Lithuania, rolanasd@centras.lt

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 351-357

Promising results obtained by the authors (us) during research work on vibration turning [3] prompted us to focus on the milling operation. The research methodology employed in this work is based on the combined numerical-experimental approach used in [3], which is extended here to the case of vibration milling. The research objective of this paper is the vibration milling of difficult-to-cut alloys with the aim of determining the influence of tool excitation conditions on the surface finish of workpieces through the application of qualitative and quantitative surface characterization methods.

resulting in a lower number of degrees of freedom (DOFs), which, in turn, reduced the computational time of simulations.

1 NUMERICAL ANALYSIS OF THE VIBRATION MILLING TOOL 1.1 Structure of the Developed Tool Prototype Fig. 1a provides a structural diagram of a prototype of the vibration milling tool developed at the Kaunas University of Technology. A ring-shaped piezoelectric transducer 8 is embedded in the vibration tool assembly consisting of components 2 to 9. The transducer, powered by collector rings 4, is used to excite high-frequency vibrations in the cutting edge of the mill cutter 10. A horn 9 with chuck is fitted onto the end of the assembly in order to augment cuttertip vibration amplitude, which may reach up to 20 μm in this case. The vibration milling tool operates in a resonance mode: its length is equal to the integral number of half-wavelengths (Fig. 1b). The tool is mounted in a standard Weldon holder 1 at a nodal point, thereby preventing vibration energy losses by dissipation into the machine tool body.

1.2 Finite Element Modelling 1.2.1 Assumptions Laboratory testing of the vibration milling tool has demonstrated that cutter dynamics has a negligible effect on the axial vibrations generated by the tool.

This finding justified our model reduction approach, which assumed that the actual excitation provided by the piezoelectric transducer may be represented as an equivalent base excitation, which is imposed on the cutter at the place where it is clamped in the chuck.

It should be noted that the vibration tool constitutes a linear dynamic system, therefore its vibrational characteristics may be established by analyzing a numerical model of a single cutter with the boundary conditions that are equivalent to those of the actual vibration tool. Therefore, a structurally complex transducer assembly was discarded in the model 352

a) b) Fig. 1. a) Structural diagram of the vibration milling tool; and b) schematics of its operation principle: 1 – standard holder (Weldon) DIN 6359, 2 – cylinder, 3 – textolite cylinder, 4 – collector rings, 5 – nut, 6 – bolt, 7 – collet, 8 – piezoceramic rings, 9 – horn, 10 – cutter

1.2.2 Model Formulation SolidWorks software was used for preparation of the 3D models of a mill cutter shaped as a pre-twisted cantilever (Fig. 2a). After exporting to FEM software ANSYS, the models were imposed with the relevant boundary conditions (Figs. 2 b and c). The main properties of FE models: cutter length l = 96 mm and l = 74 mm, diameter d = 10 mm, density ρ = 8000 kg/ m3, Young’s modulus E = 207 GPa, Poisson’s ratio n = 0.3. The models were meshed with tetrahedral finite elements SOLID92 characterised by 3 DOFs per node.

A Cartesian coordinate system was adopted for modal analysis, while for harmonic and transient simulations a cylindrical system was chosen with the following DOFs: displacements z along the cutter rotational axis (axial direction), displacements r orthogonal to the axis (transverse direction), and rotations φ about the rotational axis (torsional direction).

It was essential to accurately reproduce the actual conditions of elastic cutter clamping in the chuck since they may have a tangible influence on cutter dynamics due to variable clamping force, temperature, etc. To this end, a sub-model consisting of elastic links was inserted into the cutter model. Spring elements COMBIN14 were selected for modelling the contact interaction in the zones where the cutter is mounted in a chuck (the cutter was secured with three bolts located around the shank at 120° angles). The cutter

Ostasevicius, V. – Gaidys, R. – Dauksevicius, R. – Mikuckyte, S.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 351-357

a) b) c) Fig. 2. a) SolidWorks model of a mill cutter with marked zones where boundary conditions are set; ANSYS finite element model of b) 74 mm; and c) 96 mm length end mill with designated zones that are meshed with spring elements COMBIN14

was fixed elastically with longitudinal spring elements kr, kz and torsional elements kφ (kr, kz stiffness in transverse and axial directions, respectively, kφ - torsional stiffness). These link elements were placed at each node belonging to the three zones located on the shank (Fig. 2a). One node of a spring element was connected to the respective zone node, while the other element node was either constrained (during modal analysis) or was imposed with a base excitation in the axial direction (during harmonic and transient analyses). The developed FE model consists of about 6000 elements SOLID92 and 250 elements COMBIN14 with a total of ca. 30000 DOFs. The dynamics of the cutter are described by the equation of motion in a block form by considering that the base motion law is known and is defined by the nodal displacement vector {UK(t)}:

where {UNrel} is a component of the relative displacement with respect to the moving base displacement {UNk}.

[ M NK ] {UN } + [ M KK ] {UK }

Fig. 3. Experimental frequency response of the cutter in the axial direction and the corresponding numerical responses

[ M NN ]  [ M KN ] [C ] +  NN [CKN ]

[ K ] +  NN [ K KN ]

    [CNK ]  U N  + [CKK ]  U K    [ K NK ] {U N } =  {0}  , (1) [ K KK ] {U K } {R}

{ } { }

where {UN(t)}, {UK(t)} - nodal displacement vectors representing displacements of free nodes and baseexcited nodes, respectively; [M], [K], [C] - mass, stiffness and damping matrices respectively; {R} vector representing reaction forces of the base-excited nodes.

The displacement vector of the unconstrained nodes is expressed as:

Vectors {UK} and {UNk} correspond to rigidbody displacements exerting no internal elastic forces within the structure. A proportional damping approach is adopted in the form [C] = α[M] + β[K], where α and β denote Rayleigh damping coefficients. The final equation in a matrix form is as follows:  , [ M NN ] U Nrel + [CNN ] U Nrel + [ K NN ]{U Nrel } =  M  −1

where  M  = [ M NN ][ K NN ] [ K NK ] − [ M NK ] , (3)  

{

}

{

}

where the left-hand side of the equation contains matrices of the structure constrained in the nodes of applied base excitation, while the right-hand side denotes a vector of inertial forces acting on each node as a result of the base excitation.

{UN} = {UNrel} + {UNk}, (2) Study of Vibration Milling for Improving Surface Finish of Difficult-to-Cut Materials

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 351-357

1.2.3 Experimental Verification The model was verified experimentally in order to confirm that it is able to accurately predict the dynamic characteristics of the cutter that is excited by the vibration milling tool. The level of coincidence between the measured and simulated frequency responses was used to characterise model accuracy.

the model was adjusted by varying the stiffness of the spring elements until an acceptably close agreement between the numerical and experimental resonant frequencies was obtained. Fig. 3 illustrates a comparison between the simulated and measured frequency responses of the cutter in the axial direction (the experimental curve represents the cutter response to a harmonic excitation with a constant relative amplitude).

a)

Fig. 4. Scheme of experimental set-up for vibration milling: 1 machine construction, 2 desk, 3 workpiece, 4 vibration tool holder DIN 6359, 5 horn, 6 piezoceramic rings, 7 cutter, 8 collector rings, 9 generator, 10 amplifier, and 11 multimeter

The vibrational response of the cutter is largely predetermined by its boundary conditions. Therefore

b) Fig. 5. Photos of surfaces of machined workpieces (50× magnification): a) after conventional milling and b) after vibration milling

The responses in Fig. 3 reveal that the major measured and simulated resonance peaks coincide fairly well. The accuracy of the FE model is evaluated using the relative error, which is calculated using numerical and experimental values of the resonant frequencies. The relative error was determined to be less than 2%, which confirms the accuracy of the developed FE model.

Resonance frequencies of the axial vibration mode of mill cutters of two different lengths were calculated using the FE model, yielding 14.05 kHz for the 74 mm cutter and 18.4 kHz for the 96 mm cutter.

Table 1. Surface finish results of milled workpieces using a 74 mm length cutter Materials and milling type Stainless steel (no vibrations) Stainless steel (with vibrations) Titanium (no vibrations) Titanium (with vibrations)

1 0.61 0.48 0.60 0.49

Numbers of measurement [μm] 2 3 4 5 0.56 0.75 0.87 0.53 0.39 0.67 0.49 0.43 0.60 0.56 0.53 0.47 0.70 0.46 0.50 0.47

6 0.58 0.60 0.56 0.49

Average [μm] 0.65 0.51 0.55 0.52

Difference [%]

6 1.11 0.88 0.85 0.82

Average [μm] 1.16 0.90 0.94 0.83

Difference [%]

21.54 6.33

Table 2. Surface finish results of milled workpieces using a 96 mm length cutter Materials and milling type Stainless steel (no vibrations) Stainless steel (with vibrations) Titanium (no vibrations) Titanium (with vibrations)

354

1 1.17 1.00 0.76 0.83

Numbers of measurement [μm] 2 3 4 5 1.14 1.15 1.17 1.21 0.86 0.91 0.78 0.95 1.30 0.88 0.86 1.01 0.85 0.85 0.80 0.80

Ostasevicius, V. – Gaidys, R. – Dauksevicius, R. – Mikuckyte, S.

22.59 12.54

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2 CHARACTERISATION OF WORKPIECE SURFACE QUALITY 2.1 Evaluation of Surface Roughness A series of conventional and vibration-assisted milling experiments were performed in order to determine the effectiveness of the vibration milling tool with respect to improvement in surface quality. Fig. 4 provides a schematic representation of the measurement setup.

Experiments were carried in CNC milling center DMU 35M with workpieces made of stainless steel (1.4301) and titanium (GOST 22178-1976) without the use of cooling-lubricating fluids.

Two end mills of different length 74 mm (length of working part is 22 mm) and 96 mm (length of working part is 48 mm) but of the same diameter (10 mm) were used for the vibration milling experiments.

The cutters were manufactured by “ASP Arno” from tungsten carbide and are coated with titanium aluminum nitride (TiAlN). This coating is suitable for high temperature and high speed machining of difficult-to-cut metallic alloys with minimal use of

lubricant. Therefore this coating was chosen for our milling experiments that were carried out at high temperature without lubricating fluids. The thickness of the TiAlN coating is in the range of 2 to 4 μm and the oxidation temperatures are between 480 and 900 °C. The hardness is typically 2800 HV. Each cutter has its axial resonance at different excitation frequencies due to its difference in length (14.05 and 18.4 kHz). These frequencies were used for tool excitation in the axial direction during vibration milling experiments, which were conducted using the following regimes: milling depth ap = 1 mm, feed rate vf = 66.66 mm/min, milling speed np = 1000 rev/min.

Each material was cut with both mill cutters for 30 mm with conventional milling and then switched to 30 mm cutting with vibration assistance.

The surface of the workpieces was analysed qualitatively by means of a JEOL JSM-IC25S scanning electron microscope. The obtained images shown in Fig. 5 provide a visual proof that the surfaces of the workpieces machined with vibration milling are smoother with respect to the

a)

b) Fig. 6. Normal probability plots when 96 mm length cutter is used and the machined material is: a) stainless steel, b) titanium

a)

b) Fig. 7. Normal probability plots when 74 mm length cutter is used and the machined material is: a) stainless steel, b) titanium Study of Vibration Milling for Improving Surface Finish of Difficult-to-Cut Materials

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 351-357

conventional process. The image in Fig. 5b indicates no microcracks in the case of the vibration machined workpiece. A closer inspection of SEM images reveals that workpiece surface after conventional milling (Fig. 5a) is characterised by a ploughing effect due to cutting process instability. The surface in Fig. 5b, on the contrary, shows no signs of ploughing and is characterised by a regular finely meshed structure.

Quantitative results of surface quality were obtained using MITUTOYO SURFTEST SJ-201.

The values of the cutoff and evaluation length were 0.8 and 2.4 mm, respectively. Measurement data indicate that the values for surface roughness in the case of vibration milling are approximately one roughness grade number lower with respect to the conventional process (according to DIN EN ISO 1302). Surface roughness data for six machined workpieces was collected during this experimental study. Measurement results demonstrate that difficultto-cut materials machined using the vibration milling process are characterised by improved surface finish in comparison to the conventional process (Tables 1 and 2). When the tool was driven with an excitation frequency of 18.4 kHz (96 mm cutter), the best results for the surface finish were obtained in the case of stainless steel milling (Table 2). This result is attributed to better machinability of stainless steel 309 in comparison to titanium alloy Ti 6-4.

2.2 Statistical Analysis Roughness measurement data was subjected to statistical treatment in order to more thoroughly characterise the effectiveness of the vibration milling process.

As a first step, the assumption that the samples come from normal distributions was tested. A normal probability plot provides a quick idea (Figs. 6 and 7).

Both data scatters approximately follow straight lines through the first and third quartiles of the samples, indicating approximately normal distributions. A

shift in the mean from stainless steel and titanium machined using conventional milling to stainless steel and titanium machined with vibrations is evident.

A hypothesis test is used to quantify the test of normality. Since each sample is relatively small, a Lilliefors test is used. The test performs the default null hypothesis that the sample comes from a distribution in the normal family. The results returned by each test indicate a failure to reject the null hypothesis that the samples are normally distributed. This failure may reflect normality in the population or it may reflect a lack of strong evidence against the null hypothesis due to the small sample size.

A variance analysis was performed to estimate the relative impact of resonance frequency, applied vibrations, and material type on surface quality (Table 3). The p-value for the resonance frequency effect is 0.0159, which is highly significant. This indicates that surface roughness is largely dependent on the dynamic characteristics of the mill cutter. In turn, this implies that in order to increase the positive influence of the vibration-assisted cutting process, the tool should be superimposed with vibrations of a frequency that corresponds to the axial resonance frequency of the cutter. As a consequence, there is amplification of axial vibration amplitudes with simultaneous intensification of the twisting motion of the cutting tip due to coupling of the axial and torsional deflections inherent to the helix-shaped mill cutter, which represents a case of parametric vibrations. The p-value for the vibration milling effect is 0.0471, which is significant as well.

This indicates that the surface quality also depends on the machining means. In other words, statistically, the dynamic characteristics and machining means have the greatest influence on surface quality.

3 CONCLUSIONS This paper has presented an approach to reducing surface roughness of milled workpieces made from difficult-to-cut metallic alloys via excitation of the first

Table 3. Results of variance analysis Source Resonance frequency Vibrations Material Resonance frequency×Vibrations Resonance frequency×Material Vibrations×Material Error Total

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Sum Sq. 0.32 0.03645 0.01805 0.005 0.005 0.00845 0.0002 0.39315

d.f.

1 1 1 1 1 1 1 7

Mean Sq. 0.32 0.03645 0.01805 0.005 0.005 0.00845 0.0002

Ostasevicius, V. – Gaidys, R. – Dauksevicius, R. – Mikuckyte, S.

F 1600 182.25 90.25 25 25 42.25

Prob > F 0.0159 0.0471 0.0668 0.1257 0.1257 0.0972

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axial mode of the vibration milling tool. A combined application of numerical and experimental analysis has confirmed the validity of the proposed approach.

An experimentally verified finite element model of the vibration milling tool was built on the basis of the actual tool prototype. Complex tool structure in the numerical model was reduced to a single pretwisted cantilever, which was imposed with boundary conditions (clamping and excitation conditions) that accurately reproduce those of the actual vibration tool. The model was used to determine the resonance frequency of the axial vibration mode of cutters of two different lengths. Milling experiments demonstrated that excitation of the axial mode in the vibration milling tool leads to an appreciable reduction in the surface roughness of stainless steel and titanium workpieces. The application of qualitative and quantitative characterization methods revealed better surface quality in comparison to the conventional milling process: surfaces having one roughness grade lower finish were obtained. A statistical analysis of the collected roughness data allowed us to establish that the dynamic characteristics (excitation frequency) of the tool and machining method (with or without the assistance of high-frequency vibrations) have the largest effect on surface quality. The reported research results demonstrate that it is crucial to dynamically tailor the excitation frequency of the vibration cutting tool in order to generate the required vibration mode in the mill cutter and thereby achieve the most pronounced improvement in surface finish in difficultto-cut materials.

The proposed approach allows efficient machining of high-strength alloys and could significantly facilitate the treatment of hard and brittle materials such as ceramics, glass, and composite materials. The reported vibration milling experiments were successfully performed under dry machining conditions, which demonstrates that assisting cutting with high-frequency vibrations could benefit the implementation of a minimum quantity lubrication method into industrial manufacturing processes.

4 ACKNOWLEDGMENTS This research work was funded by a grant from the Lithuanian Agency of Science, Innovation and Technology for the implementation of EUROSTARS project No. E!7288.

5 REFERENCES [1] Kumabe, J. (1979). Vibration Cutting. Jikkyo Publishing, Tokyo.

[2] Brehl, D.E., Dow, T.A. (2008). Review of vibrationassisted machining.

Precision Engineering, vol. 32, no. 3, p. 153-172, DOI:10.1016/j. precisioneng.2007.08.003.

[3] Ostasevicius, V., Gaidys, R., Rimkeviciene, J., Dauksevicius, R. (2010). An approach based on tool mode control for surface roughness reduction in high-frequency vibration cutting. Journal of Sound and Vibration, vol. 329, no. 23, p. 4866-4879, DOI:10.1016/j.jsv.2010.05.028.

[4] Twardowski, P., Wojciechowski, S., Wieczorowski, M., Mathia, T. (2011). Surface roughness analysis of hardened steel after high-speed milling. Scanning, vol.

33, no. 5, p. 386-395, DOI:10.1002/sca.20274.

[5] Chen, C-H., Wang, Y-C., Lee, B-Y. (2013). The effect of surface roughness of end-mills on optimal cutting performance for high-speed machining. Strojniški vestnik - Journal of Mechanical Engineering, vol. 59, no. 2, p. 124-134, DOI:10.5545/sv-jme.2012.677.

[6] Taylan, F., Colak, O., Kayacan, M.C. (2011).

Investigation of TiN coated CBN and CBN cutting tool performance in hard milling application. Strojniški vestnik - Journal of Mechanical Engineering, vol. 57, no. 5, p. 417-424, DOI:10.5545/sv-jme.2010.059.

[7] Kubiak, K., Wilson, M., Mathia, T., Carras S. (2011).

Dynamics of contact line motion during the wetting of rough surfaces and correlation with topographical surface parameters. Scanning, vol. 33, no. 2, p. 370377, DOI:10.1002/sca.20289.

[8] Imani, B., Moosavi, S. (2009). Time domain simulation of torsional-axial and lateral vibration in drilling operation. Proceedings of International Conference on Application and Design in Mechanical Engineering, Penang, Malaysia, p. 9G1-7.

[9] Zhang, J.Z, Chen, J.C. (2008). Tool condition monitoring in an end-milling operation based on the vibration signal collected through a microcontrollerbased data acquisition system. International Journal of Advance Manufacturing Technology, vol. 39, no. 1-2, p.

118-128, DOI:10.1007/s00170-007-1186-6.

[10] Seguy, S., Dessein, G., Arnaud, L. (2008). Surface roughness variation of thin wall milling, related to modal interactions. International Journal of Machine Tools & Manufacture, vol. 48, no. 3-4, p. 261-274, DOI:10.1016/j.ijmachtools.2007.09.005.

[11] Grazeviciute, J., Skiedraite, I., Jurenas, V., Bubulis, A., Ostasevicius, V. (2008). Applications of high frequency vibrations for surface milling. Mechanika, vol. 1, p. 4649.

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 358-366 © 2013 Journal of Mechanical Engineering. All rights reserved.

DOI:10.5545/sv-jme.2012.920

Original Scientific Paper

Received for review: 2012-12-18 Received revised form: 2013-02-28 Accepted for publication: 2013-03-29

A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations Petrun, T. – Flašker, J. – Kegl, M. Tomaž Petrun1,* – Jože Flašker2 – Marko Kegl2

2 University

1 AVL-AST d.o.o., Slovenia of Maribor, Faculty of Mechanical Engineering, Slovenia

This paper deals with a theoretical and numerical study of various viscosity terms in the modified elasto-plastic friction model and their influence on the resulting friction force-torque, transmitted through the contact of the friction clutch. Various simple viscous definitions for fluids considering shear rate dependent viscosity were investigated. The Carreau fluid model was chosen as a basis for the research, since it can describe Newtonian, dilatant, and pseudo plastic fluids. In addition to theoretical investigations, numerical simulations under realistic friction clutch operation conditions were carried out. The results were compared to the results of a validation case for a dry friction clutch simulation using the modified elasto-plastic friction model. This research showed significant differences between various viscosity definitions and revealed the drawbacks of such approach, were simple viscous models were used.

In addition to viscosity, calculating the heat generated due to friction and its influence on the contact temperatures are discussed briefly.

The basic theory and equations are given along with the directions for future work. The requirements for an accurate temperature calculation in the friction contact are outlined.

Keywords: friction model, friction clutch simulation, simple viscosity models, Carreau fluid

0 INTRODUCTION In the development process of a complete vehicle powertrain, each part of the whole system must be considered accurately. To be able to do this, numerical software tools with adequate models of single parts, assemblies, and connections between bodies are needed.

The friction clutch is an important part of the vehicle powertrain. Therefore, modelling and simulation of a fully functional friction clutch is vital for accurate powertrain response simulation. In order to extend and upgrade the functionality of the commercial software code for numerical multi-body simulations, AVL EXCITE, such a model with special requirements has been developed and validated experimentally, Petrun et al. [1]. The development and validation was carried out for dry friction clutch applications. In the automotive industry, especially in passenger cars, the most common type of clutch is the dry friction clutch due to its good drivability, comfort, and its simple and relatively cheap construction.

However, wet clutches are also used, although they are more common in high power engine applications, where performance is more important than comfort.

In these applications the fluid is used to achieve the desired tribological characteristics of the contact pair (usually metal to metal) and also to cool the friction clutch.

358

This research work focuses on additional viscous terms for the modified elasto-plastic (EP) friction model, proposed by Petrun et al. [1]. The development and experimental validation of the EP Friction Model was focused only on dry friction clutch applications where very good agreement for various commercial friction materials was achieved. However, in order to account properly for wet friction clutch applications, one must also consider the viscous forces in the open and synchronization phase, due to the viscosity of the fluid present.

In the literature, Bukovnik et al. [2], Fajdiga et al. [3], many complex viscous models have been proposed, where the dynamic viscosity η is a function of the shear rate γ , temperature T and pressure p, for example: Vogel, Barus, Cross, Rodermund, Kuss, Roelands, etc. Furthermore, the resulting viscous force in these models also depends on the real geometry and patterns of the contact pair, lubricant flow conditions, pressure distribution in the contact, etc. To get accurate results with these models, complex multimaterial simulations, including flexible bodies and an adequate fluid model, are needed to calculate the pressure and gap height distribution in the contact.

The friction model discussed in this work is intended to be implemented in the target software AVL EXITE. This fact imposes some special requirements, outlined as follows. The friction model must connect two bodies in the multi-body system as a joint, Fig. 1. The model must be able to calculate the exact

*Corr. Author’s Address: AVL-AST d.o.o., Trg Leona Štuklja 5, 2000 Maribor, Slovenia, tomaz.petrun@yahoo.com

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 358-366

amount of the transmitted torque through the friction contact. Available inputs for the friction model are the relative sliding velocity of the bodies in contact.

The model should account for important friction induced dynamics. Currently, the friction model is still 1D. Therefore, the investigated viscous models should also be 1D and only velocity dependent. For this reason, only simple shear rate (velocity and gap height) dependent viscous models will be investigated in this work.

calculation are given in order to indicate the direction of future work.

1 THE MODIFIED EP FRICTION MODEL The modified elasto-plastic friction model for dynamic friction clutch simulations in a multi-body system, proposed by Petrun et al. [1], is a member of the Single State Variable (SSV) friction model family, Dupont et al. [4] and [5]. These models describe the friction contact as a contact of two brushes, where bristles interact, Fig. 2.

Fig. 2. Friction contact description by a SSV friction model

The resulting friction force FfEP of the modified EP friction model is defined as follows: F fEP = (σ 0 z + σ 1 z ) FN , (1)

Fig. 1. Example of a multi-body system in the target software for friction model implementation

The main focus of this research is a theoretical and numerical investigation of the influence of various viscosity models on the behaviour of the friction clutch and the contribution on the resulting friction force.

In particular, the influence on the synchronization process and the so-called drag force in the open phase are of main interest. This work focuses on four different simple viscosity models, which could be used for the application at hand. Furthermore, a basic theory for the calculation of heat generation due to friction will be briefly discussed at the end. Namely, the consequences of the heat generated due to friction are temperature changes at the friction contact. Since temperature has a notable influence on the tribological characteristics of the contact and on friction induced dynamics, this is a very important topic, [1]. The main requirements and demands for an accurate temperature

where σ0 represents the stiffens of the brush, σ1 is the damping coefficient of the brush, z is the bristle deflection, also called the pre-displacement, z is the first time derivate of z, FN is the normal force, and F the external force, see Fig. 2. Note that the state variable z is calculated from the differential equation for z :

z =

 dz sign(v)  = v 1 − α ( z , v) z  , (2) dt g (v )  

where, v is the relative sliding velocity, α(z,v) is a function controlling the friction contact state, and g(v) is a function describing the tribological characteristics of the friction contact, e.g. the Stribeck curve. For more details please see [1], [4] and [5].

This modified elasto-plastic friction model is capable of continually providing accurate results of the actual transmitted friction torque through the friction contact in all operational phases of a friction clutch. Furthermore, this model accounts for realistic tribological parameters and friction induced dynamics

A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations

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and was successfully validated against experimental measurements.

The model setup and experimental validation were carried out for dry friction clutch applications, where very good agreement was achieved. For wet friction clutch applications, additional terms for viscosity must be added. An advantage of this model is that additional terms can easily be embedded as additive terms into the expression for the resulting friction force [1], [4] and [5]. For our purpose the extended friction model can be written as follows:

F fEP + = (σ 0 z + σ 1 z ) FN + Fv , (3)

where, Fv represents the resulting viscous force defined by an adequate viscosity model.

2 SIMPLE VISCOSITY MODELS Four different simple viscosity models will be investigated. In these models the dynamic viscosity η is considered either constant or shear rate dependent.

For all presented models the same contact area AC is taken into account and the relative velocity v is averaged at the mean radius of the friction clutch. The mean radius is defined as, [6]:

Rm =

2  Ro3 − Ri3  3  Ro2 − Ri2

  , (4) 

where Ro is the outer radius and Ri is the inner radius of the friction clutch contact area.

The simplest viscosity force definition is as follows, [1], [4] and [5]:

Fv = η ( γ ) AC

dv . (7) dh

The shear rate is:

γ =

dv , (8) dh

where the dynamic viscosity is defined as:

(

2 η ( γ ) = η0 1 + ( λγ )

)

n −1 2

. (9)

As one can see, this viscosity definition is not linear anymore. Here η0 is the dynamic viscosity at a given temperature and zero shear rate, the factor λ is a fluid (material) constant, and n is a flow behaviour index, which defines the type of the fluid, Fig. 3, as follows: < 1  n =1 >1 

pseudo plastic fluid Newtonian fliud . (10) diilatant fluid

The viscosity of pseudo plastic fluids decreases with increasing shear rate (for example; toothpaste).

This type of fluid is also known as a shear thinning fluid. Newtonian fluids have a constant dynamic viscosity regardless of the changing shear rate (for example, water, gasses, some oils, etc.). Dilatant fluids react to increasing shear rate with increasing viscosity. They are also called shear thickening fluids.

This type of fluid is less common.

Fv = σ v v, (5)

where σv represents a constant viscosity function which can be defined as:

η AC . (6) h Here η represents the dynamic viscosity of the fluid, AC is the contact area, and h is the gap height.

All these parameters are considered to be constant; the only variable is the relative velocity v. The viscosity function σv can be temperature dependent. This viscosity model depends linearly on velocity.

The next 3 viscosity models are special cases of the so-called Carreau fluid definition. Here, the dynamic viscosity η (γ ) is shear rate dependent and the resulting viscous force is defined as follows, [7] to [10]:

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σv =

Fig. 3. The Carreau fluid model – shear rate dependent viscosity for various flow behaviour index values

In the case of the flow behaviour index n = 1, the definition for viscosity becomes:

η ( γ ) = η0 . (11)

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 358-366

In this case, the resulting viscous force becomes quite similar to the simple linear definition, i.e. Eq. (5).

In this research work, the following 3 values of the flow behaviour index were chosen: n1 = 0.9, n2 = 1 and n3 = 1.1, shown in Fig. 3. These values were chosen just to represent the differences between various fluid types. One can use any values for the flow behaviour index. The parameter λ is in all cases set to λ = 1.

3 HEAT GENERATED DUE TO FRICTION It is commonly known that friction generates heat, which is typically dissipated. The heat generated causes warming up of the friction contact materials.

This temperature change changes the tribological characteristics of the friction contact pair. The change in tribological characteristics has a direct influence on the amount of transmitted torque in the slip phase and on the load capacity in the stick phase. It can also have a significant influence on friction induced dynamics, [1]. The influence is even stronger for wet friction clutch applications due to a change in the viscosity of the fluid.

The amount of heat generated depends on the resulting friction force Ff and relative velocity v, [11] to [14], as follows.

Qslip = F f v. (12)

This heat is generated during sliding friction. A small amount of heat is also generated when the bodies in a friction contact stick. The proposed modified EP friction model is also dissipative for the stick phase, Dupont et al. [4] and [5]. The amount of generated heat in the stick phase is defined for the modified EP friction model as:

Qstick = σ 1 z. (13)

The heat generated in the stick phase is typically negligible, compared to the heat generated in the sliding phase of a friction clutch.

Qstick << Qslip = Q. (14)

The generated heat Q raised the temperature of the bodies in contact, of nearby bodies, and the surrounding environment. This temperature change depends on the thermal capacity of each body, the contact area, and the heat flux distribution between bodies. For simplification reasons the heat flux is

assumed to be limited only to bodies in the friction contact. The total heat flux is defined as: Q q= . (15)

AC The heat flux into each body depends on the thermal conductivity of the body material and the heat transfer coefficient. Since the friction contact of an automotive friction clutch usually consists of two totally different materials with different thermal characteristics, the heat flux into each body must be defined separately. The heat flux into body 1 can be defined as:

q1 = xq, (16) where, x represents the portion of the total heat flux q flowing into body 1. The heat flux q2 into body 2 is then defined as:

q2 = (1 − x)q, (17)

where

q = q1 + q2 . (18)

When the friction contact is lubricated as in the case of a wet friction clutch, submerged in a fluid, the total heat flux must be divided into three parts. In a lubricated, submerged friction contact, a portion of the heat q3 is conducted into the lubrication - cooling fluid. The total heat flux is then:

q = q1 + q2 + q3 . (19)

To get accurate results for heat flux distribution, temperature change, and local temperature distribution, a full 3D thermal finite element (FE) and finite volume fluid dynamics (CFD) simulation is required, [15]. Since, in this research stage, the friction model used is rather simple and implemented as 1D for now, an accurate thermal calculation is not possible. Besides this, the target numerical code for multi-body system simulations does not currently allow the calculation of temperatures within the involved bodies. Therefore, since temperature is an important influencing factor, it will be addressed in more detail in future research work.

4 VALIDATION MEASUREMENTS In order to validate the developed modified EP friction model, a special test bed was build. The test bed consists of two shafts. Each shaft is attached at

A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations

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Fig. 4. Test bed assembly for friction model validation

one end to a torque sensor and an electro motor and at the other end to a steel disc, as shown in Fig. 4.

One electro motor is used as a drive (engine side; ES) and the other one as a brake (brake side; BS).

On one disc, an additional and replaceable disk is mounted. In that way, different replaceable discs with various commercial friction materials attached can be mounted. The friction contact occurs by bringing into contact the outer surface of both discs with the help of a hydraulic axial bearing. This friction contact represents a simplified friction clutch, shown in Table 1. No conventional automotive friction clutch was used since the conventional friction clutch is a complex multi-body system where parts are connected using various connection types (rivet connections, springs, dampers, shaft couplings, etc.). The properties of these connections are hard to predict since they depend on manufacturing procedure, tolerances, etc.

In the simplified friction clutch, all parts are rigidly connected and the only unknowns are the tribological parameters of the contact.

Table 1. Simplified clutch parameters Ro Ri μC μS vS

0.095 [m] 0.060 [m] 0.350 [/] 0.420 [/] 0.500 [m/s]

This test bed with a simplified friction clutch enables the acquisition of angular velocities on both sides of the test bed, transmitted torques on both shafts, and the pressure (normal force) in the actuation system using a LabView interface. With the help of a LabView program and an electro motor controller, 362

realistic conditions, similar to those in a vehicle powertrain, can be created and measured.

The tribological parameters were calculated for each measurement case from the measured values for the relative sliding velocity, the normal force, and the transmitted torque. For this research work, the values for the tribological parameters listed in Table 1 were used.

5 RESULTS In this section, the results for three different friction clutch synchronization simulation cases are presented.

The first case is an artificial one in order to show the difference between various viscosity models where various flow behaviour index values were used.

For simplification reasons, the gap was considered constant in this case. In the second case, a variable gap height was considered and compared to the first case.

The third case is a comparison to a real validation case presented in [1]. Compared to the original dry validation case, various viscosity terms were added to the friction model equation. Again, for simplification reasons the same tribological parameters were used as in the validation case. This approach gives a good quantitative comparison of the contributions of viscosity for wet friction clutch applications.

All cases represent a synchronization of two rotating bodies, as was the case during the validation measurements on the test bed. At the beginning, the engine side body rotates with a prescribed constant angular velocity and the break side body rests at zero rpm. After synchronization, both bodies rotate with the prescribed angular velocity of the engine side.

For all simulations of a given case (using various viscosity models), the same initial and boundary

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 358-366

conditions were used. The contact area was also the same for all simulations.

5.1 Differences between Various Friction Definitions In this example, the differences between various viscosity definitions are presented. All models have the same initial viscosity value η0 at zero shear rate.

The value for η0 is artificial, but is nevertheless within the range of realistic values of oils used for wet friction clutches. For the Carreau fluid, 3 different

flow behaviour index n values were used to represent the differences between various values, see Fig. 3.

In Fig. 5, one can see that the synchronization process is quite different for various viscosity definitions. The synchronization starts immediately due to the viscous force and proceeds rapidly after two seconds, when normal force is applied and the friction force prevails. No load is applied to the break side body. The synchronization duration depends on the viscosity definition. For shear thickening fluids – dilatant fluids, the process finishes much faster than

Fig. 5. Synchronization of two rotating bodies with various viscosity definitions (BS –brake side, ES – engine side)

Fig. 6. Transmitted torque through the friction contact during the synchronization – comparison of various viscosity definitions A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations

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Fig. 7. Comparison of viscous torque for constant and variable gap height for the Newtonian fluid definition

Fig. 8. Comparison of angular velocities for dry validation case and additional viscosity term for wet applications for various viscosity definitions (BS –brake side, ES – engine side)

for Newtonian and shear thinning fluids due to the increasing viscosity with increasing shear rate.

One can also see that the results for the simple viscous definition and for Newtonian fluids are quite similar.

In Fig. 6, the torques, transmitted through the friction contact, are plotted. As in Fig. 5, it can be seen that the viscosity definition significantly influences the synchronization process.

As already mentioned, the only input variable in this case is the relative sliding velocity of the bodies in contact, since the artificial gap is considered constant. Next, the results for a variable gap height are presented.

364

5.2 The Influence of a Variable Gap Height In this example, the same load was used as in case 1.

The gap height is normal force dependent and is, at the beginning, at an initial value of 1 mm. After the normal force is applied, the gap is proportionally reduced to a value of 0.1 mm. These values of gap heights were chosen for simplification reasons. In real cases, the end gap is 0 mm.

In Fig. 7, the resulting viscous torque is plotted for the Newtonian fluid. It can be seen that the difference between the model with a constant gap and the model with a variable gap is minor for the cases calculated here. When compared to the total torque transmitted through the friction contact, the influence of the

Petrun, T. – Flašker, J. – Kegl, M.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 358-366

Fig. 9. Comparison of transmitted torque through the friction contact for the dry validation case and the additional viscosity term for wet applications for various viscosity definitions

variable gap height can be practically neglected since no visible difference in velocities and synchronization times was observed on the velocity plots. This is also true for all other viscosity models used in this research work.

For the conditions considered so far, one can say that the viscosity of fluids (for the fluid models used here) is mainly velocity dependent, since, when looking at the equation for the shear rate (Eq. 8), the velocity change is much higher than the gap height change.

5.3 Comparison to a Real Validation Case To investigate the influence of the additional viscosity term in the modified EP friction model, a comparison to a real validation case was carried out. This case was used to validate the modified EP friction model for dry friction clutch applications, where very good agreement of measured and simulated results was achieved, [1]. To demonstrate the influence of various viscosity models, only part of the validation case was used, the open phase, the synchronization, and a small part of the locked phase.

In addition to the simulation results for the dry friction clutch, the results for all four viscous models are plotted. To be able to compare those results, the tribological parameters were assumed to be the same for dry and wet simulations, see Table 1.

In Fig. 8, it can be seen that viscosity influences the response only when the surfaces of the friction

clutch are sliding relative to each other. As soon as the velocities of the bodies in contact synchronize (locked phase), the viscous part of the torque becomes zero and the system behaves as if it were dry. The same can be seen in Fig. 9.

6 CONCLUSIONS In this research work, a theoretical and numerical investigation of simple viscosity models in a friction contact was carried out. The viscosity terms were added to a modified EP friction model equation, which was validated for dry friction clutch simulations and will be implemented into the commercial numerical tool for multi-body analyses AVL EXCITE. As expected, the results confirm that the viscosity definition of the fluid may have a significant influence on the resulting friction force/torque and consequently on the friction clutch synchronization process. Therefore, such simulations can show only quantitative trends for various viscosity models. For qualitative, realistic results, an experimental validation would be necessary. These simple models probably do not provide highly accurate results since the real 3D geometry of the contact is not taken into account.

For the cases presented here, it was also discovered that the shear rate is mainly velocity dependent since high velocity changes are present and the gap change is minimal.

This research also shows that the modified EP friction model enables simple addition of various force

A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations

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terms to the expression for the resulting friction force/ torque, transmitted through the friction contact. These additional terms have no influence on the solution process of the main EP friction model equation. They also do not change the structure of the main equation.

The only difference is in the resulting friction force/ torque, transmitted through the friction contact and, of course, in the contributions of each term (friction force, viscous force, etc.).

In order to introduce at least an approximate temperature dependency, the underlying equations for friction generated heat are relatively simple. However, the hard part is to determine the heat flux into each body and the actual contact temperature, which decisively influences the tribology characteristics of the contact pair. To achieve this, a full 3D FE model and simulation would be needed. When using a wet clutch one would even need a multi-material model, including a CFD simulation. These important topics will need to be investigated in future work.

7 FUNDING This research work was partly funded by the European Union, European Social Fund and SPIRIT Slovenia, the Slovenian Public Agency for Entrepreneurship, Innovation, Development, Investment and Tourism.

8 REFERENCES [1] Petrun, T., Flašker, J., Kegl, M. (2012). A friction model for dynamic analyses of multi-body systems with a fully functional friction clutch. Proceedings of the Institution Mechanical Engineers, Part K: Journal of Multi-Body Dynamics. DOI:10.1177/1464419312464708.

[2] Bukovnik, S., Offner, G., Čaika, V., Priebsch, H.H., Bartz1, W.J. (2007). Thermo-elasto-hydrodynamic lubrication model for journal bearing including shear rate-dependent viscosity. Lubrication Science, vol. 19, no. 4, p. 231-245, DOI:10.1002/ls.45.

[3] Fajdiga, D., Glodež, S., Flašker, J. (1998). Numerical simulation of elastohydrodynamic lubricated line contact problems. Strojniški vestnik – Journal of Mechanical Engineering, vol. 44, no. 9-10, p. 285-296.

[4] Dupont, P., Armstrong, B., Hayward, V. (2000).

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[5] Dupont, P., Armstrong, B., Hayward, V. Altpeter, F. 2002). Single State Elastoplastic Friction Models. IEEE Transactions on Automatic Control, vol. 47, no. 5, p.

787-792, DOI:10.1109/TAC.2002.1000274.

[6] Drexl, H.J., (1997). Kraftfahrzeugkupplungen: Funktion und Auslegung. Die Bibliothek der Technik, Band 138, Verlag Moderne Industrie, Landsberg/Lech. (in German) [7] Gotz, T., Parhusip, H.A. (2005). On an asymptotic expansion for Carreau fluids in porous media. Journal of Engineering Mathematics, vol. 51, no. 4, p. 351-365, DOI:10.1007/s10665-004-7468-1.

[8] Barrett, J.W., Liu, W.B. (1993). Finite element error analysis of a quasi-Newtonian flow obeying the Carreau or power law. Numerische Mathematik, vol.

64, p. 433-453, DOI:10.1007/BF01388698.

[9] Koh, J.H., Kwon, I., Jung, H.W., Hyun, J.C., (2012).

Operability window of slot coating using viscocapillary model for Carrau-type coating fluid. Korea-Australia Rheology Journal, vol. 24, no. 2, p. 137-141, DOI:10.1007/s13367-012-0016-z.

[10] Strnadel, J., Simon, M., Machač, I. (2011). Wall effects on terminal faling velociti of spherical particles movingin a Carreau model fluid. Chemical Papers, vol.

65, no. 2, p. 177-184, DOI:10.2478/s11696-011-00056.

[11] Evtushenko, O.O., Pauk, V.I. (2002). Steady-state frictional heat generation on a periodic sliding contact.

Journal of Mathematical Science, vol. 109, no. 1, p.

1266-1272, DOI:10.1023/A:1013757030298.

[12] Reibenschuh, M., Oder, G., Čuš, F., Potrč, I. (2009).

Modelling and analysis of thermal and stress loads in train disk brakes – braking form 250 km/h to standstill.

Strojniški vestnik – Journal of Mechanical Engineering, vol. 55, no. 7-8, p. 494-502.

[13] Živanović, Z., Milić, M. (2012). Thermal Load of Multidisc Wet Friction Assemblies at Braking Regime.

Strojniški vestnik – Journal of Mechanical Engineering, vol. 58, no. 1, p. 29-36, DOI:10.5545/sv-jme.2009.111.

[14] Yevtushenko, A.A., Kuciej, M. (2012). Onedimensional thermal problem of friction during braking: The history of development and actual state.

International Journal of Heat and Mass Transfer, vol. 55, no. 15-16, p. 4148-4158, DOI:10.1016/j. ijheatmasstransfer.2012.03.056.

[15] Tic, V., Lovrec, D. (2012). Design of Modern Hydraulic Tank Using Fluid Flow Simulation. International Journal of Simulation Modelling, vol. 11, no. 2, p. 7788, DOI:10.2507/IJSIMM11(2)2.202.

Petrun, T. – Flašker, J. – Kegl, M.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 367-374 © 2013 Journal of Mechanical Engineering. All rights reserved.

DOI:10.5545/sv-jme.2012.679

Received for review: 2012-06-27 Received revised form: 2012-12-15 Accepted for publication: 2013-01-08

Original Scientific Paper

Effect of Scratch Velocity on Deformation Features of C-plane Sapphire during Nanoscratching Feng, P. – Zhang, C. – Wu, Z. – Zhang, J. Pingfa Feng – Chenglong Zhang* – Zhijun Wu – Jianfu Zhang

Tsinghua University, Department of Mechanical Engineering, The State Key Laboratory of Tribology, China The effects of scratch velocity on plastic and brittle deformation features of (0001) C-plane sapphire were studied in nanoscratch tests. The test was conducted under a ramping loading condition from 40 μN to 200 mN using a nanomechanical test system. A Berkovich nanoindenter was employed in this study. The scratch velocities were set at 2, 4, 8, and 16 μm/s. Plastic and brittle deformation features were observed by scanning electron microscopy. The residual stress features of the deformation zones in the scratch groove were observed by Raman spectroscopy. Comparative studies of surface depth profiles and scratch groove features obtained with different scratch velocities reveal that the scratch velocities have distinct effects on the deformation features of C-plane sapphire.

Keywords: (0001) C-plane sapphire, nano-scratch test, scratch velocity, plastic deformation, brittle deformation, residual stress

0 INTRODUCTION As a good wave-transparent material, sapphire crystals have been used in many fields, such as highspeed integrated-circuit chips, laser substrates, etc., due to their inherent characteristics [1] to [5]. The processed surface quality of the sapphire substrate has a strong influence on the application performance of sapphire elements. Moreover, the cost of precision and ultra-precision processing of the fine surface during sapphire manufacturing for optoelectronic applications is high [6] to [9]. Therefore, an investigation into the material removal features and deformation behaviours of sapphire is necessary and can be applied to guide the sapphire manufacturing process.

The removal modes of brittle materials affect the quality of the surface processed under conventional machining conditions. Brittle fracture removal and plastic deformation removal are regarded as the two main material removal modes in the processing of brittle materials [10] to [16]. The brittle fracture removal mode has a negative effect on the processed workpiece surface and leads to a lower quality of the processed surface. Under suitable processing conditions the plastic deformation removal mode can be achieved, resulting in adequate surface quality. The scratch test, regarded as a process analogous to the machining of materials, has been employed to study the deformation behavior and tribological properties of materials [17] to [21]. The deformation patterns induced in the scratch process provide preliminary information needed to determine the material removal behaviour.

According to the removal modes of the materials mentioned in the previous section, it can be demonstrated that a smooth scratch groove can be

produced in a completely plastic mode with low depth of cut (DOC) scratching, thus helping to eliminate brittle fracture features, such as cracks and chipping dents, in brittle materials, as also reported in the literature [22] to [24]. The nanoscratch tests available for scratch experiments carried out under extremely low DOC can be employed to reveal the plastic and brittle deformation features of brittle materials.

In this paper, the nanoscratch tests were carried out to reveal the removal properties and deformation behaviors of (0001) C-plane sapphire. The effects of scratch velocities on plastic and brittle deformation features were studied. Scanning electron microscopy (SEM) was employed to observe the deformation features. Residual stress features of the deformation zones in the scratch grooves were studied by Raman spectroscopy. Comparative studies of the surface depth profiles and scratch groove features induced by different scratch velocities are presented.

1 EXPERIMENTAL PROCEDURE Scratch experiments were conducted on a nanomechanical test system (Nano Indenter G200, Agilent Corp., USA) in scratch mode. The size of the sapphire sample was 15×10×1 mm. C-plane (0001) sapphire that is suitable for infrared detector applications was selected for use in this study. A diamond Berkovich nanoindenter tip was employed.

The scratch velocities were set at 2, 4, 8, and 16 μm/s. The scratching track was set at 300 μm in all tests. The applied normal load was increased linearly from 40 μN to 200 mN along the scratch length. The process for a standard scratch is shown in Fig. 1.

First, a pre-scan was carried out in order to measure the initial surface property of the C-plane

*Corr. Author’s Address: Tsinghua University, 9003 Building, 100084, Beijing, China, zcl08thu@gmail.com

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sapphire sample under a low load of 40 μN. Then, the indenter penetrated into the sapphire workpiece with the pre-set loading characteristics. Finally, the residual deformation feature was recorded by scanning the scratch groove again (post-scan) under a 40 μN load.

Penetration depth and residual deformation data were also collected during the test, which can be used in quantitative analysis of the deformation features of the sapphire workpiece.

Fig. 1. Schematic diagram of the scratching

The sapphire sample was ultrasonically cleaned in acetone for 15 min. The morphologies of the cleaned scratch grooves were then observed by SEM (Quanta 200 FEG, FEI Corp., Netherlands). Raman spectroscopy (LabRAM HR800, HORIBA Jobin Yvon S.A.S., France) was employed to reveal the residual stress features of the deformation zones in the scratch groove. The Raman spectrometer included an Ar+ laser (514 nm) to excite the specimen. A maximum laser power of 20 mW was used. Raman spectra were collected within and outside the scratch groove at room temperature to compare the residual stress distribution features of the sapphire sample.

2 EXPERIMENTAL RESULTS AND DISCUSSION 2.1 Analysis of Scratch Profiles Fig. 2 presents the scratch profiles made with the diamond Berkovich nanoindenter tip at scratch velocities of a) 2, b) 4, c) 8, and d) 16 μm/s. The pre-scan curves show the unscratched surface of the sapphire workpiece. The penetration depth profiles (scratching) reveal the real-time deformation behaviours of the sapphire sample. The post- scan 368

curves indicate the removal features of the sapphire workpiece induced by the scratching. Negative scratch depth means the indenter is located below the initial position of the workpiece surface, and positive depth indicates the indenter is moved above the initial surface by the scattered debris produced during the scratch test.

The pre-scan profiles in Fig. 2 indicate that the surface of the prepared sapphire sample is smooth.

The profiles also show very small elastic–plastic deformations (Figs. 2a and b), which might have been caused by the pre-scan load (40 μN) as the scratch velocity is relatively low. With the increase in the applied load, there is no visible fluctuation in the scratching profile, implying that the deformation of the sapphire sample is a completely elastic–plastic deformation under this loading condition. When the load reaches a certain value, the scratch depth abruptly increases. This load is termed the “critical load” (Fc), and the depth is termed the “critical depth” (Dc); these critical values are usually used to study the ductile–brittle transition characteristics of the material removal. The critical load and critical depth were then determined from Fig. 2 and nanoscratch data were collected: these are presented in Table 1.

By comparing the collected curves, it can be observed that there are three distinct deformation stages during the scratch processing. In the first stage, the scratching and post-scan profiles are almost smooth, which is clearly visible in Figs. 2c and d but not obvious in Figs. 2a and b. This indicates that the deformation of the sapphire sample is completely elastic–plastic for the c) 8 and d) 16 μm/s scratch velocities, but the material is removed in plastic mode and leaves fish-bone-like traces for the a) 2 and b) 4 μm/s scratch velocities.

In the second stage, the scratching depth– displacement curves show small waves, while the post-scan depth–displacement curves show fluctuations. This implies that microcracks and severe fish-bone features are produced during the scratching.

In the third stage, abrupt changes and larger waves in the scratching and post-scan curves appeared, which means that larger microcracks and chipping dents emerge within the grooves and the material is thus removed by brittle deformation. The starting positions of the abrupt changes and larger fluctuations differ for the a) 2, b) 4, c), and d) 16 μm/s scratch velocities in Fig. 2, indicating that the critical loads and critical depths differ for different scratch velocities.

Therefore, as the load increases the material removal of sapphire during scratch processing is

Feng, P. – Zhang, C. – Wu, Z. – Zhang, J.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 367-374

a)

b)

c)

d) Fig. 2. Surface depth profiles for scratch velocities of a) 2, b) 4, c) 8, and d) 16 μm/s under loading conditions ramping to a maximum of 200 mN

described as elastic–plastic deformation, plastic deformation with the generation of severe fish-bone features, and brittle deformation at the highest load.

The load and depth ranges for each scratch velocity in the three stages are given in Table 1.

Table 1. Load ranges and critical depth for each scratch velocity Scratch velocity [μm/s] 2 4 8 16

Stage I [mN]

Stage II [mN]

Stage III [mN]

Fc [mN]

Dc [mN]

0 to 12 0 to 15 0 to 25 0 to -30

12 to 120 15 to 165 25 to 180 30 to 190

120 to 200 165 to 200 180 to 200 190 to 200

120 165 180 190

630 790 805 810

Table 1 shows that the critical load is 120, 165, 180, and 190 mN and the critical depth is 630, 790, 805, and 810 nm for scratch velocities of a) 2, b) 4, c) 8, and d) 16 μm/s, respectively. It can be observed that Fc and Dc increase with scratch velocity. This trend

shows that higher scratch velocity can lead to a higher proportion of plastic deformation and results in better surface quality of the workpiece.

2.2 Morphology Features of the Scratch Grooves To compare the distinct scratch regions, the scratch process was divided into four parts according to the applied load. Figs. 3a) to d are SEM micrographs of the scratch groove obtained with a scratch velocity of 2 μm/s.

In the low load stage, a smooth groove appeared with small plastic pile-ups and ironing lines within the groove (see Figs. 1a and 3a), implying a complete elastic–plastic deformation. With an increase in the normal load, the groove becomes broader and microcracks, chipping dents, and slip lines are observed. As the load increases to approximately 80 mN, microcracks are generated around the scratch

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and fish-bone traces and stick-slip lines emerge in the scratch, as shown in Fig. 3b.

mN, the damaged regions induced by the microcracks extension appear and brittle fracture removal features are clearly observed, as shown in Fig. 3d. Therefore, the material removal of the C-plane sapphire induced at a scratch velocity of 2 μm/s is a combination of plastic deformation and brittle fracture.

Fig. 4 shows SEM micrographs of deformation features induced with a scratch velocity of 16 μm/s.

The deformation features of the scratch groove obviously differ from the features induced with a scratch velocity of 2 μm/s (as shown in Fig. 3).

Plastic pile-up is seen around the groove when the applied load is low, as shown in Fig. 4a, indicating that elastic–plastic deformation occurred during this scratch process. Moreover, the slip lines clearly show scattering from the scratch groove (shown in the white elliptical region and the magnified image in the white rectangular region in Fig. 4a), but are not visible when the scratch velocity is 2 μm/s (as shown in Fig. 3a).

Fig. 3. Micrographs of deformation features induced under a scratch velocity of 2 μm/s

a)

Fig. 4. SEM morphology of deformation features induced with a scratch velocity of 16 μm/s

Chipping dents around the scratch and tearing regions within the groove are observed as the applied load reaches about 120 mN, as shown in Fig. 3c.

Owing to the increase in the applied load and the induced load, microcracks along the scratch direction within the groove are also seen. The emergence of chipping dents indicates that brittle fracture of the C-plane sapphire occurred when the applied normal load was 120 mN with a scratch velocity of 2 μm/s.

When the normal load increases to approximately 200 370

b) Fig. 5. Raman spectra and Raman shift within and outside the groove; a) Raman spectra, b) Raman peak positions

Other researchers have reported that the observed plastic deformation features of scratch processing, i.e. pile-up and microcracks, result from the tensile

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stress field [24] and [25]. To clarify the material removal mechanism of sapphires, particularly for plastic deformation, Raman spectroscopy was applied to measure the stress characteristics of the scratch groove resulting from a constant scratch velocity of 16 μm/s, as shown in Fig. 5. The Raman peaks measured in the groove were at lower wave numbers than those measured in the undeformed area of the sapphire workpiece (see Figs. 4a and 5a). The Raman peaks’ shift to lower wave numbers indicates the generation of tensile stress within the scratch groove [25]. Therefore, it can be concluded that tensile stress appears within the scratch groove during scratching.

Tensile stress is the fundamental factor causing the stacking faults, location loops, and dislocation glide that are observed during the plastic deformation mechanisms of sapphire in a nanoscratch.

With the increase in scratch load, microcracks, tears, and chipping dents are observed in and around the scratch groove at a scratch velocity of 16 μm/s.

However, no brittle fracture region can be observed when the scratch velocity is 16 μm/s. The deformation features of the maximum load (approximately 200 mN) induced at scratch velocities of 4 and 8 μm/s are also observed and compared in this study, as shown in Fig. 6.

a higher scratch velocity leads to a higher proportion of plastic deformation during the scratch process, which is very consistent with the surface depth profiles for the different scratch velocities (see Fig. 3). It is thus inferred that better processing performance can be achieved with an increase in cutting velocity, which is analogous to the scratch velocity, in the surface processing of sapphire and other brittle materials.

3 DISCUSSION To understand the mechanism of different deformation features of the sapphire sample during the scratch process with different scratch velocities, the effects of scratch velocity on strain rate and hardness and the strain rate sensitivity as scratch velocity increases are addressed in the following section.

According to its definition, the strain rate ( ε ) can be determined from the scratch velocity (v) using the scratch depth (h) in the plastic deformation stage of the sapphire sample, expressed as:

v ε = . (1) h

Fig. 7 shows the effects of scratch velocity on strain rate under different scratch loads. It can be seen that the strain rate increases with increasing scratch velocity. Increasing the scratch load will have a negative influence on strain rate.

Fig. 6. Deformation features at the maximum load (approximately 200 mN) induced with scratch velocities of 4 and 8 μm/s; a) scratch velocity is 4 μm/s, b) scratch velocity is 8 μm/s

Fig. 6 shows that brittle fracture regions are seen around the groove for a scratch velocity of 4 μm/s, but not for a scratch velocity of 8 μm/s (Fig. 6b). Plastic and brittle deformation features are observed in and around the scratch grooves for scratch velocities of 4 and 8 μm/s, and they are similar to the deformation features of the scratch grooves (as shown in Figs.

3 and 4) for scratch velocities of 2 and 16 μm/s as analysed in detail in the previous section.

Comparison analysis of the deformation features of the scratch grooves induced with different scratch velocities reveals that the scratch velocity affects the material removal features of the C-plane sapphire, i.e.

Fig. 7. Effects of scratch velocity on strain rate

Fig. 8 shows the effects of scratch velocity on hardness at lower loads where plastic deformation occurs without fracture. The calculation formula of hardness (H) is expressed as:

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H=

P , (2) A

where P is the scratch load and A is the projected area of contact surface between the indenter and the workpiece, which can be determined by scratch depth (h) in this study. It can be seen that hardness increases as scratch velocity increases. Increasing scratch load will have a negative effect on hardness.

According to the computed strain rate and hardness, the strain-rate sensitivity (m) can be obtained by the following equation:

m=

∂ (ln H ) . ∂ (ln ε )

(3)

Fig. 9 presents the strain-rate sensitivity as scratch velocity increases. It shows that the strain-rate sensitivity is positive with increasing scratch velocity.

Fig. 9. Strain-rate sensitivity as scratch velocity increases

Existing research has shown that, due to increasing strain rate, the dynamic hardness from the dynamic indentation is greater than the static hardness for the C-plane sapphire, and dynamic indentation can effectively shorten the indentation-induced crack length relative to the crack length for static indentation [26] and [27]. In this study, all the above indicates that the scratch depth decreases with increasing scratch velocity, which infers an increase in hardness and strain rate. In other words, increasing scratch velocity will influence the mechanical properties of the sapphire sample, thus increasing the hardness, and will effectively restrain the occurence and growth of cracks; this leads to fewer and smaller cracks, and the plastic deformation is thus more comparable to the lower scratch velocity.

372

4 CONCLUSIONS Nanoscratch experiments were conducted on a (0001) C-plane sapphire at scratch velocities of 2, 4, 8, and 16 μm/s. Surface depth profiles recorded with the test system and deformation features observed by SEM for the different scratch velocities are discussed in this study. The residual stress features of the deformation zones in the scratch groove were observed by Raman spectroscopy to clarify the material removal mechanism of sapphire. The following conclusions are drawn.

(1) The surface depth profiles and microstructure observations made by SEM show that scratch depth and width increase with an increase in applied load.

(2) Deformation in each scratch process can be described as plastic deformation, plastic deformation with an increase in scratch depth and the appearance of microcracks, and brittle deformation with the emergence of chipping dents and damaged regions.

(3) Raman spectra obtained with a micro-Raman spectrometer show Raman peaks at lower wave numbers within the scratch groove compared to the ones collected outside the groove. It is inferred that tensile stress is present within the scratch groove during scratching, leading to the plastic deformation of the sapphire during scratching.

(4) Comparative studies of the surface depth profiles and scratch groove features induced with different scratch velocities reveal that the scratch velocity has distinct effects on the deformation features of C-plane sapphire. With increases in the scratch velocity, the surface depth profiles show that the critical load and depth increase and the scratch groove features show more slip lines and less chipping dents, implying that a higher scratch velocity leads to a higher proportion of plastic deformation in the scratching of C-plane sapphire.

(5) Examination of the mechanism of different deformation features for different scratch velocities shows that increasing scratch velocity can improve the strain rate and hardness, and can effectively restrain the occurence and growth of cracks, which leads to fewer and smaller cracks, and thus the plastic deformation observed is more comparable to that seen at a lower scratch velocity.

5 ACKNOWLEDGMENTS This research was financially supported by the National Natural Science Foundation of China (Grant No. 50975153) and the State Key Laboratory

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of Tribology Foundation of China (Grant No. SKLT11C7).

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A new method for evaluating the scratch resistance of diamond-like carbon films by the nano-scratch technique. Diamond and Related Materials, vol. 11, no.

7, p. 1454-1459, DOI:10.1016/S0925-9635(02)000456.

[19] Subhash, G., Bandyo, R. (2005). A new scratch resistance measure for structural ceramics. Journal of the American Ceramic Society, vol. 88, no. 4, p. 918925, DOI:10.1111/j.1551-2916.2005.00181.x.

[20] Pogrebnyak, A.D., Beresnev, V.M., Kaverina, A.S., Kolesnikov, D.A., Yakushchenko, I.V., Ilyashenko, M.V., Makhmudov, N.A. (2012). Adhesive strength and physical, mechanical, and triboengineering properties of nano- and microstructural Al2O3 coatings. Journal of Friction and Wear, vol. 33, no. 3, pp. 195-202, DOI:10.3103/S1068366612030087.

[21] Pogrebnyak, A.D., Beresnev V.M., Kaverina A.S., Kolesnikov D.A., Yakushchenko I.V., Ilyashenko M.V., Makhmudov N.A. (2012). Structure, morphology, physical and mechanical properties of nano- and microstructured coatings of Al2O3 and ZrO2. Physics

Effect of Scratch Velocity on Deformation Features of C-plane Sapphire during Nanoscratching

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and Chemistry of Material Treatment, vol. 46, no. 5, p.18-25.

[22] Yao, W.L., Liu, J., Holland, T.B., Huang, L., Xiong, Y.H., Schoenung, J.M., Mukherjee, A.K. (2011). Grain size dependence of fracture toughness for fine grained alumina. Scripta Materialia, vol. 65, no. 2, p. 143-146, DOI:10.1016/j.scriptamat.2011.03.032.

[23] Klecka, M., Subhash, G. (2008). Grain size dependence of scratch-induced damage in alumina ceramics.

Wear, vol. 265, no. 5-6, p. 612-619, DOI:10.1016/j. wear.2007.12.012.

[24] Ghosh, D., Subhash, G., Radhakrishnan, R., Sudarshan, T.S. (2008). Scratch-induced microplasticity and microcracking in zirconium diboride–silicon carbide composite. Acta Materialia, vol. 56, no. 13, p. 30113022, DOI:10.1016/j.actamat.2008.02.038.

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Measurement of scratch-induced residual stress within SiC grains in ZrB2–SiC composite using micro-Raman spectroscopy. Acta Materialia, vol. 56, no. 18, p. 53455354, DOI:10.1016/j.actamat.2008.07.031.

[26] Haney, E.J., Subhash, G. (2011). Static and dynamic indentation response of basal and prism plane sapphire. Journal of the European Ceramic Society, vol. 31, no. 9, p. 1713-1721, DOI:10.1016/j. jeurceramsoc.2011.03.006.

[27] Klecka, M.A., Subhash, G. (2010). Rate-dependent indentation response of structural ceramics. Journal of the American Ceramic Society, vol. 93, no. 8, p. 23772383, DOI:10.1111/j.1551-2916.2010.03729.x.

Feng, P. – Zhang, C. – Wu, Z. – Zhang, J.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 375-386 © 2013 Journal of Mechanical Engineering. All rights reserved.

DOI:10.5545/sv-jme.2012.809

Original Scientific Paper

Received for review: 2012-09-24 Received revised form: 2013-03-13 Accepted for publication: 2013-03-15

Detecting Groan Sources in Drum Brakes of Commercial Vehicles

by TVA-FMEA: A Case Study Karabay, S. – Baynal, K. – İğdeli, C. Sedat Karabay1 – Kasım Baynal2,* – Cengiz İğdeli2

1University

of Kocaeli, Faculty of Engineering, Mechanical Engineering Department, Turkey of Kocaeli, Faculty of Engineering, Industrial Engineering Department, Turkey

2University

In this article, the strategy followed by an automobile company for detecting the root causes of groan complaints related to rear drum brakes in commercial vehicles is presented using data collected from the sold troubled vehicles and from new vehicles from the production line, as well as the drum-brake test rig at the laboratory. Drum brake groan is often very intense and can cause large numbers of customer complaints.

During a groan noise event, vehicle structure and suspension components are excited by the brake system and result in a violent event that can be heard and felt during brake application. This paper condenses experimental studies on a low frequency drum brake groan problem that has caused high warranty costs. First, the environmental conditions causing the groan were identified and the groan was reproduced. Vehicle tests were performed both at the factory and in traffic. To conclude the planned study, TVA (Total Value Analysis) and FMEA (Failure Mode and Effects Analysis) methods were used effectively. A strategy to determine the root causes was planned and implemented systematically to eliminate the secondary and tertiary effects of brake groan problems. In order to ge to the root causes, vibration and groan measurements were executed and interpreted according to TVA and FMEA charts. The sensitivity of the lining material of the brake shoes to different environmental conditions was investigated. Finally, the groan mechanism of the drum brake system is discussed and the solution to the low frequency drum brake groan problem is evaluated in detail.

Keywords: Drum brake, TVA, FMEA, NVH, Friction coefficient, Brake Groan, Brake lining

0 INTRODUCTIONS Automotive brake noise and vibration control have become increasingly important for the improvement of vehicle quietness and passenger comfort. Groan or creep groan is a high-intensity, low-frequency noise and vibration problem that affects road vehicles at very low speeds. It usually persists for short periods of time, but a skilled driver can deliberately make it last several seconds by tuning the force exerted on the brake pedal. The original cause is considered to be a self-induced vibration of the brake components, due to the friction material characteristics that make the system prone to a stick slip behaviour [1]. Groan and moan (100 to 500 Hz) are caused by vibrations due to the dynamic instability of the wheel-brake system, as is the “howl” but at higher frequencies (500 to 1000 Hz). Groan and moan are caused by stick–slip motion between the friction material and the disc surface. The superposition of several quasiharmonic vibrations or unsteady vibration impulse sequences of different intensities are its main features.

Therefore, this phenomenon is characterized by nonlinear excitation and non-linear transfer mechanisms [2]. Groan (often called creep groan) generally occurs at low vehicle speed and also occurs with cold brakes and high humidity [3]. Creep groan (low frequency frictional vibration during braking at low vehicle speed) was studied by Jang and et al. by changing the relative amounts of ingredients in a brake friction

material containing 12 ingredients. Twenty-nine friction material specimens with different relative amounts of the ingredients were manufactured according to a constrained mixture design. The difference (μ) between kinetic (μk) and static (μs) coefficients of friction for each formulation was measured to investigate the creep groan propensity of each friction material since the creep groan is caused by stick–slip phenomena. Results showed that zircon (zirconium silicate), steel wool, and phenolic resin showed a tendency to increase μ [4]. The usual method for measuring the creep groan is based on a subjective test pilot evaluation carried out on different ramp inclinations. This method is not very precise and can generate different results associated with breaking noise for different pilots. An accelerometer was installed in the caliper brake system to capture the vibration intensity by Luciano et al. [5].

The tribological contact in automotive brakes involves dry sliding contact at high speeds and high contact forces. The commonly used organic bindertype brake pad friction materials are extremely inhomogeneous and exhibit very low bulk strengths.

Despite the low strength, the specific contact surfaces that form during use give the pads very good friction and wear characteristics [6]. In an automotive disk brake system, when the disk shows thermo elastic instability (TEI) at speeds over the critical one, a hot spot forms on the disk surface and local contact friction is developed between the disk surface and

*Corr. Author’s Address: University of Kocaeli, Faculty of Engineering, Industrial Engineering Department, Kocaeli, Turkey, kbaynal@yahoo.com

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the pad due to the hot spot. This non-uniform contact friction worsens the local heat concentration on the disk surface and has a direct effect on the disk pad, through which it affects the oil pressure supplier of the disk [7]. Brake creep-groan is studied via a friction coupled torsional model consisting of driveline and brake subsystems. The model captures the main torsional modes of interest, while a suitable reduction of the higher degree-of-freedom models allows the selection of appropriate parameters.

Numerical simulations are programmed that capture groan response with stick-slip friction and transient brake pressure [8]. Detection and evaluation of “creep groan” noise has been a challenge for NVH test groups.

First, this sound typically is not purely tonal like the more common brake squeal, although ultimately it may produce a tonal subjective impression. In this work the authors study different methods that may be applied to “creep groan” detection and evaluation [9]. In Yoon et al., creep groan noise was reproduced using a chassis dynamometer. Through vibration measurements and modal impact tests, the effects of the vehicle system on creep groan noise were analysed [10]. Jung and Chung studied brake creep groan noise using a dynamometer. Their simulation results were confirmed through dynamometer testing and the groan noise contribution factor analysis between chassis components was then presented using an experimental approach [11]. Many fundamental studies have been conducted to explain the occurrence of brake groan noise in disc and drum brake systems. The elimination of brake noise, however, still remains a challenging area of research. Here, a numerical modeling approach is developed for investigating the onset of brake groan noise in a drum brake system. The brake system model is based on the modal information extracted from finite element models for individual brake components [12] and [13]. A 5-DOF non-linear model is presented to simulate the vibration of a drum brake at low frequency in the course of applying the brake. One of the main issues was the movement of the pads over the rotor fins resulting in the dynamic groan type of noise. It was important to relate this to the customer complaints of grinding [14] and [15]. In recent years, the main focus on brake groan problems has shifted from fundamental theoretical research to more practical and problem-solving oriented efforts such as TVA and FMEA [16] to [18].

1 VALUE ANALYSIS Value analysis (VA) or total value analysis (TVA) and value engineering can be defined as an organized 376

and systematic approach to providing the required function at the lowest cost consistent with specific performance, quality, and reliability. Value analysis is a functionally oriented scientific method to improve the product value from the customer’s point of view with reference to the elements of the product’s cost in order to accomplish the desired function at the cost of resources deployed to produce the product.

The following steps are to be followed for the value analysis.

• Collect data about cost function, customer needs, history, and likely future developments related to the product and its use. Determine the function of the product.

• Develop alternative design. The selected alternatives should be able to fulfill the functional requirement of the product as detailed above.

• Ascertain the cost of the alternatives.

• Evaluate the alternatives in all respects. The alternative which fulfills all the basic or primary value considerations and maximum number of secondary value considerations is the ideal alternative, subject to the cost consideration, which should be minimized.

• Recommend and implement the best solution.

Identify the control point and devise a plan for periodic measurement of the performance and correct the deviations if any.

2 DEFINITION OF PROBLEM During analysis via the global warranty tracking database, it became clear that something had gone wrong in August 2005 as indicated in Fig. 1.

According to the records, the drum casting supplier was changed in December 2004 and then the brake shoe lining process was given to a company working in the home market in Turkey in February 2005. The vehicles built after implementation of these changes had a lot of complaints related to NVH (noise, vibration, harshness) occurring during braking. The only repair method to deal with these complaints was to replace the affected parts under warranty, however this resulted in a considerable increase in warranty cost per month for the company. The collected complaints showed that the problematic unit of the light truck vehicle was the “drum brake system”. Therefore, the company organized a team to investigate the root cause of the drum brake low frequency NVH problem. In the literature, brake noises were classified according to frequency ranges [1]. Groan is the low frequency noise 100 to 500 Hz. Thus, for a better understanding of the problem, various NVH departments of the company

Karabay, S. – Baynal, K. – İğdeli, C.

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Fig. 1. Warranty cost per affected units

were asked for vibration and noise measurements and also interpretations of the records. The prepared flow chart for the planned studies on the affected systems and components is shown in Fig. 2. According to the chart, design verifications of all affected systems and/ or components (suspension system and rear brake & drum) will be reviewed in detail to understand if there were any other design changes that may have caused the problem. On the other hand, additional studies will be performed such as: receiving the complained about parts and assembling them into a master vehicle and checking if the noise is transferred to this vehicle, as well as communicating with as many customers as much as possible in order to understand the mechanism of the groan noise and trying to understand whether the source of the problem is unique to some specific markets or climates. It is clear that groan noise in brake systems is generated by vibrations. Thus, noise and vibration occur simultaneously in a problematic mechanical system. Harshness is a combination of vibration and noises. In other words, it is an adjective given by customers after subjective assessment of a defective system. The basic reasons for the low frequency brake noise can be categorized as indicated in Fig. 2.

Thus, basically four groups of studies were planned to determine the root cause of the NVH problem shown in Fig. 3. However, the problem contains a high level complexity. Thus, the main study diagram given in Fig. 3 was constructed to evaluate data simultaneously collected from different research areas. Customer explanations related to the complaints depended on subjective definitions.

Therefore, the company should construct a test rig at its laboratory to simulate the customerâ&#x20AC;&#x2122;s complaints.

If the required performance is not taken from the test

rig at the laboratory, the working team must perform some measurements on the vehicle with groan noise under real working conditions in order to detect NVH sources. On the other hand, the components of the drum brake and its effect on other vehicle components must be technically checked according to a plan. This plan has been organized into a fishbone diagram as shown in Fig. 4.

Fig. 2. Basic reason of brake groan

Fig. 3. Basic studies planned by the company to evaluate data simultaneously to determine root cause

3 EXPERIMENTAL STUDIES As indicated in the Fig. 3, test studies at the company were performed in four separate areas. Each subgroup within the company had a different interest and approach to the research. The data were collected and then interpreted by the leaders of the each subgroup.

The first group was interested in tracking groan

Detecting Groan Sources in Drum Brakes of Commercial Vehicles by TVA-FMEA: A Case Study

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Norway

Spain

Portugal

Germany

Ireland

England

Czech Republic

Italy

France

Netherland

Austria

Belgium

Table 1. Groan complaint in vehicles for countries where the vehicle was sold [%]

Denmark

reasons using the fish bone diagram, while the second group researched customer complaints and contacted customers directly regarding groan complaints in vehicles. The third group collected noise data from the test rig at the company laboratory, i.e. noise (groan) heard by the customer from the vehicle. The fourth study group performed an NVH test regarding groan complaints for vehicles in a real environment (in traffic) and in the factory test area. In the tests, vehicles with different bodies, suspension, and brake systems were used. The vehicle’s features are reported using data obtained from each test. All of the vehicles used in the tests had drum brakes and the general specifications are as follows: short wheel base vehicles; long wheel base vehicles; ABS vehicles; non-ABS vehicles; vehicles with a single leaf spring and rear stabilizer bar; vehicles with a double leaf spring and rear stabilizer bar.

2

12

20

6

8

5

4

6

8

2

11

6

10

3.2 Laboratory and Rig Tests To perform simulations of the brake groan problem in the laboratory, the problematic brake and suspension parts from the vehicle were disassembled and then a new set-up using the defective parts was employed in order to determine the root cause of the NVH source under laboratory conditions to simulate customer complaints.

3.1 Listening to VOC and Market Survey By contacting the customers through the service department, the source and mechanism of the groan was researched and test drives were performed using the affected vehicles with replaced parts, then the results and interpretations were recorded. Additionally, the vehicles sold in the largest 15 countries in Europe were studied to determine whether the drum–brake groan problem is related to different climates. As a result of statistical analysis, the groan brake problem is independent regarding the countries as indicated in Table 1.

Fig. 5. Drum brake test rig constructed to study the groan problem

Several test trials were performed. However, the vehicle mass moment of inertia cannot be created dynamically and therefore the brake groan, which is defined by customers subjectively for the activated brake, cannot be generated under laboratory

Fig. 4. Fishbone diagram to detect the reason for drum brake groan

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conditions. Therefore test rig studies were stopped and it was decided that all NVH measurements should be performed under real working conditions using the affected vehicles while driving on the road. The test rig constructed in the laboratory to measure drumbrake groan for commercial vehicles is shown in Fig. 5.

3.3 Determination of possible contributions to groan All systems and components, as well as the quality of design and production processes, are reviewed within the fish bone diagram given in Fig. 4. The details of the tests performed on the systems and components are summarized below. In the fishbone diagram the most important item is the drum component of the rear brake system. Therefore, the drum was thoroughly investigated.

3.3.1 Drum 3.3.1.1 Dimensional Control of the Drum and Its Components The main constructive parts of the rear drum brake are illustrated in Fig. 6. Here the second part of the left side is a wheel short shaft. The rear drum brake was studied in depth and some damaged surfaces on the brake table due to sharpness of corners of wheel short shaft were detected. A new technical drawing of the wheel short shaft was drafted providing a longer radius in order to decrease the sharpness of the corners.

Machining of the new parts proceeded according to the modified drawing in order to eliminate contact scratches on the brake table. The more complex part in the drum brake mechanism is the brake table, which is shown in Fig. 8. 3D surface scanners were used in quality control for the machining of this part.

Additionally an etalon was used for quality control of the mounting surfaces of the basic components.

Fig. 6. Drum brake and its components

Data collected for the numbered surfaces from 1 to 9 are summarized in Table 2. Here the most important machined surfaces are the surfaces coded as 7 and 8, which belong to the shoes and the wheel short-shaft.

First, changes in the z axis were determined to be within acceptable limits after measuring the surfaces

as indicated in Table 2. Then controlling of the other items was started.

3.3.1.2 TVA Past As indicated in Fig. 1 for December 2004, casting of the drum supplier was changed to optimize the total cost of the component. Thus, the item “brake drum” has a TVA past. However, the sub-contractor working with the vehicle company on machining of the drum parts was not changed. Therefore the brake components had to be checked in detail.

3.3.1.3 Chemical Content of Materials Due to the TVA past of the drum, its chemical content was determined using spectral analysis. After checking the related parameters, it was determined that the chemical content of the material was the same as the proposed specifications of the company. Thus, this item was also eliminated as a potential root cause of the brake groan.

3.3.1.4 NVH Measurements The natural frequencies of the former and new castmanufactured drums were measured to see whether there were any changes in their frequency levels.

The measured data are shown in Fig. 7. According to the specifications, the acceptable range of frequency characteristic function (FCF) for the drum is between 760 to 815 Hz. The new drum’s natural frequency level was 814 Hz. The natural frequency of the etalon sample is 791 Hz.

Fig. 7. Natural frequency measurement

Thus, the measured values of the former and new drums are within the specified range, which means that chemical composition and natural frequency parameters can be eliminated as a potential source of the groan.

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3.3.1.5 Machining Machining of the drum was performed according to the specifications shown in Fig. 9 up to the start of the research program on the brake groan problem.

Fig. 9 shows that machining of surfaces A, C and E was being performed at high feed rates, while surfaces coded as B and D were machined at low feed rates.

The rule defined in Fig. 9 was to eliminate axial movement of the brake shoes and its lining material from the surface of the drum when the brake was being activated. During the research program aimed at identifying the root cause of brake groan, a small change was made related to the machining procedure for the drum.

The surfaces of drum B and D were machined at high feed rates and A, C and E were machined at low feed rates as indicated in Table 3. Observations and measurements were made to determine whether there was any decrease in the number of troubled vehicles manufactured from the lines. It became obvious that the high feed rate contributes to high surface roughness.

Fig. 8. A drum brake table and its important surfaces coded with numbers for use in other component montages

A rough surface accommodates the lining material of the brake shoe to the drum very quickly. It also avoids axial displacements of the lining material with the brake shoes when the system is activated.

Thus, it was understood from the same measurements that if high feeding rates are used in the machining of the middle regions, which are the contact surfaces of the lining material, then the brake groan problem decreases as shown in Table 3. However, the results did not satisfy the project team as they did not show complete elimination of the rear drum brake groan problem. The statistical measurements did not show strong correlation between drum machining and brake groan. Therefore, the drum machining method is a reason but not the basic root cause of the groan of the

Table 2. Data collected from the 3D surface scanner for the component given in Fig. 8.

Surface 1 2 3 4 5 6 7 8 9

Measured dimensions [mm] x y z -66.805 81.201 -9.000 -61.937 -84.693 -8.952 -104.661 0.822 -9.005 60.842 -84.396 -8.856 104.136 1.419 -8.883 66.948 81.497 -8.940 25.148 15.830 -2.353 -28.833 -17.116 -2.356 8.428 -79.032 -43.665

Etalon dimensions [mm] x y z -66.805 81.201 -9.000 -61.937 -84.693 -9.000 -104.661 0.822 -9.000 60.842 -84.396 -9.000 104.136 1.419 -9.000 66.948 81.497 -9.000 25.148 15.830 -2.500 -28.833 -17.116 -2.500 8.428 -79.032 -42.750

x 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000

Variations [mm] y 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000

z 0.000 -0.047 0.005 -0.143 -0.116 -0.059 -0.146 -0.143 0.915

Difference [mm] 0.000 -0.047 0.005 -0.143 -0.116 -0.059 -0.146 -0.143 0.915

Table 3. Machining of the drums according to different feed rates and tested vehicles (Machining of B, D with high feed rates and others with small feed rates of the drum as defined in Fig. 9) Surface 1 2 3 4 5 6

380

First feed rate [mm/rpm] 0.13 0.14 0.10 0.13 0.20 0.08

Second feed rate [mm/rpm] 0.12 0.10 0.10 0.12 0.12 0.14

Number of vehicle tested 1404 812 186 68 138 219

Number of vehicle with brake groan 561 29 19 15 6 5

Karabay, S. – Baynal, K. – İğdeli, C.

% 40 4 8 22 4 2

Application of feed rates Small differences Small differences Constant feeding Small differences Big differences Big differences

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brake. Thus, other possible causes were considered according to the research plan.

Fig. 9. Drum machining details

3.4 NVH Tests on Troubled Vehicles With the help of vehicle engineering specialists, groan noise and vibration measurements in affected vehicles were performed to localize the sources of the problem and the occurrence of noise type. As described in the following sections, various kinds of NVH equipment and their accessories were used to carry out a detailed study.

The sound pressure level in-cab has been measured by mounting microphones in the ears of drivers and passengers. The location of the microphones was as shown in Fig. 10. Microphones were placed as close as possible to the ears of the passenger and driver. The type of microphones used in the measurements is the Brüel & Kjaer 4189–A–021.

Fig. 10. Microphone positioning in cabin onto driver head guard

When the vehicle was tested, all the windows were closed and the drivers and passengers were in a quiet environment. A Brüel & Kjaer microphone with type 2671 pre-amplifier was tested and calibrated up to values 93.8 dB - 1000 Hz. The NVH test of the vehicles were carried out in three stages: (1)

collecting data with accelerometers and microphones located at the outside of the backing plate, (2) taking measurements from both the leading and trailing brake shoes and (3) measurement of the leading shoe of a faulty drum brake unit. To begin with, double accelerometers were placed onto the backing plates of the left and right rear drum brakes as shown in Fig. 11.

The accelerometers used were 4 mA – 100 mV/g and type 7254A–100 and 2258–100. In addition, optical readers were used to compare the collected data with the vehicle speed. Data collected for the cabin sound level and data for the backing plate acceleration were transferred to an FFT analyser for processing. Data were analysed using Head Acoustics Artemis and SQ Lab-ΙΙΙ via computer software. The vehicle used was a van type vehicle with 90 HP and steering on the right affected by brake groan. In the test, only selected vehicles with similar properties were compared.

Therefore, trouble-free vehicles to be tested had the same properties as the troubled vehicle.

Fig. 11. Accelerometers mounted on the backing plate

Experiments have shown that the character of this type problem changes with time. For that reason, wavelet analysis was used to get reliable data about interpretation of the records of the brake system responses when it was activated. To get groan noise as heard by the customer, the hand brake of the vehicle was activated at the three tooth level and then in 1st gear, the vehicle was then accelerated up to 3000 rpm.

Under these conditions, the groan noise was generated and recorded. The same test conditions using the same NVH equipments were applied to trouble free vehicles and the required tests were performed.

The second stage of the NVH measurement in the brake mechanism was to measure acceleration of the brake shoes. For this, accelerometers were mounted in the leading and trailing shoes as shown in Fig. 12. The color of the leading shoes is white and the trailing shoe yellow. However, records collected from the separated

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shoes did not yield important results. Therefore, a third trial was considered and the required set-up of the equipment was carried out. Accelerometers were mounted on the top and bottom portions of the leading (white color lining) brake shoe as shown in the Fig. 13. Then, test drives for generating brake groan noise were performed. Data collected from the last measurements show very interesting results.

vibration can be clearly observed between 200 to 300 Hz. When we look at the charts for the right wheel of the vehicle, the x-axis of both top and bottom points shows the same variability. Moreover, on the y axis, the top portion includes more interaction, which hints that the source of the excessive vibration is near the brake lining. Finally all measurements performed in the tests point to something being wrong with the brake lining. In contrast to the other axes, records taken from the z axis show no big change. In the third test, the vehicle brake groan disappears spontaneously after applying the brake for a certain period of time (groan noise is completely eliminated after driving 5 km in traffic).

Fig. 12. Accelerometers mounted on the leading and rear brake shoes

4 TEST RESULTS

Fig. 13. Positioning the accelerometers in the top and bottom portions of the leading shoes

Analysis of the data coming from the first and second measurements did not yield any clear conclusions.

However, the third measurement study did help to explain the collected data very clearly and pointed the way to identifying root cause of the problem. In the third trial, brake groan was found to be due to an activated faulty brake, which is triggered by the front metal shoe (leading) on the both sides of the drum brake of the vehicle as shown in Fig. 13. This is why this experiment was performed with the help of two accelerometers mounted in front metal shoe (the shoe in driving direction) of the drum–brake vehicle. The recordings are presented in Fig. 14. In Fig. 14, the first row indicates the measurements of the accelerometers located in the rear drum brake shoes that are, RLLT (rear left-leading top) measures groan noise frequencies in x, y, z directions and similarly RL-LB (rear left- leading bottom) in x, y, z directions. For the right tyre, the accelerometer measured frequencies at RR-LT in the x, y, z directions and RR-LB in the x, y and z directions. After having completed an analysis of the graphics, it can be seen that the left side of the vehicle does not have any serious problems. But on the right side, variability in the acceleration of

During the third trial, the frequency of the knocking groan noise occurrence at the drum brake was also measured and a continuous error mode has been detected. Although vehicle speeds and brake forces differ from each other, records show that the braking groan noise, which is heard as “knock” at each rotation of the wheel, repeats itself regularly. The knock occurs when the wheel speed was calculated to be 2.7 cycles per second, more specifically the knocking occurrence frequency is ~2.75 rotation per second. Thus, measurements show knocking at each rotation in the braking system as shown in Fig. 15.

The groan noise recorded in the cabin is shown in Fig. 16. In Fig. 16, the blue line (R. Right - L. Bottom X) indicates groan noise magnitude collected from the lower portion of the brake metal shoe and the red line˝(R. Right - L. Top X) shows the data from the top portion of the same metal shoe. The green line (R. Right - Microphone) seen in Fig. 16 indicates the levels of braking groan noise measured in the cabin.

Although the groan created during braking is actually very high pitched, we hear only a small portion of it due to the quite effective insulation applied to the cabin. A third trial, in addition to the studies described

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Fig. 14. Accelerometer records from the front metal shoes of a troubled vehicle

above, was performed to simulate the behavior of the brake lining and its pinned metal shoe during vibration of the braking unit. To do this, data collected using “Head Acoustics Artemis HEIM and software, 8.0” was loaded and run.

brake lining movement in the range of 0.2 seconds to be observed.

Fig. 16. Brake groan heard in the cabin of the vehicle Fig. 15. Knocking period of the drum brake system

A model was established by means of “real-time operational data MeScope”. This study has attempted to understand how movement occurs between the brake lining and the metal parts. A simulation was performed with the help of data collected from the top and bottom portions of the brake metal shoe and the results have been summarized in Fig. 17. This allowed

According to the file describing this movement, when braking is triggered, the brake lining is trying to hold onto the drum surface due to the force applied by brake but gripping of the shoe lining to the drum surface is not occuring. The shoe lining tends to cause displacement of the pinned shoe surface. This can be described as a dynamically forced motion. Thus, after the braking force stops, the restoring forces on the

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Fig. 17. Vibration and groan noise records of the shoe pad contact on the drum during activation

braking shoe should displace the lining to its initial position. When this movement occurs, the friction coefficient between the drum metal and lining material decreases significantly and braking is performed only by pushing two different materials towards each other.

5 DISCUSSIONS OF RESULTS According to the simulation of the behavior of the shoe and its brake lining, a new research strategy was determined for checking whether the chemical properties of the lining material changes during activation of a troubled drum brake on the vehicle. In this section, we report on the following: 1. taking a vehicle from the production line with a troubled drum brake, 2. lifting the vehicle and disassembling the troubled drum brake, 3. using two pieces of cloth and hot water to moisten the lining material on the shoes, 4. wrapping the cloth around the shoes for 5 minutes, 5. observing the surfaces of the lining material to see whether there are any changes, 6. mounting the disassembled brake parts quickly to prevent drying of the brake lining, 7. driving the vehicle under maximum control to determine whether the troubled brake makes noise and creates vibrations. Following the procedure explained above, three troubled vehicles were reserved from the production line and then driven in the test area. It has been noted and confirmed by the drive specialists that none of the vehicles makes groan noises or vibrations due to activated drum braking after moistening the shoe lining material. The chemical process parameters of the suppliers from foreign 384

and local markets were checked carefully before and after starting of the TVA. We finally realized that the curing process of the lining material does show some differences between the local and foreign (initial) suppliers. This curing difference creates chemical bonding changes in the material. Material cured by the local supplier is open to excessive heating and therefore overcuring occurs. This effect results in a decrease in the friction coefficients of the material and an occurrence of excessive vibration and groan noise in the brake unit. The curing parameters of the suppliers are given as a) foreign supplier: 200 seconds at 195 °C; b) local supplier: 390 seconds at 200 °C. From the analysis of the parameters applied by the suppliers, it can be clearly seen that the local supplier applies excessive heating comparing to the foreign supplier (5 °C higher heating temperature and 190 seconds longer). According to this finding, it has been decided to switch back to the original (foreign) supplier in order to keep service costs low and to keep a good brand image. The local supplier was also encouraged to correct their curing process in order to keep business in the future. General vehicle sales and customer complaints were then observed for a couple of months and these confirmed that warranty costs were decreased radically by changing the supplier.

A schematic representation of all the testing and inspections are shown in Fig. 18a. The items framed with a green dotted line are within specifications. The items summarized in the boxes framed with a blue dotted line had no effect on determining the root cause of brake groan.

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material under heat, the surface of the lining material becomes excessively crystallized, which decreases braking efficiency. Due to a loss of braking effects upon contact with the drum’s internal surface, the lining works as a pressure applicator on the drum only and this results in vibration under dry friction. After having determined the root cause of the problem, the drum brakes of all vehicles in the production lines are manufactured using correctly cured lining materials.

a)

a)

b) Fig. 18. a) Schematic view of the planned tests and b) bonding process of lining material

The items presented in the red dotted line are important indicators in determining the root cause.

Fig. 18b shows the shoe bonding process and the negative effects of process variability. The effect of the curing process timing was analysed under laboratory conditions by the brake lining supplier and confirmed that the variability in the curing timing has an important effect on groan noise generation.

The friction coefficient (µ) of the brake lining has been indicated as µ = 2.0 to 2.5 in the supplier specifications. In order to understand how the friction coefficient varies between “groan noise” and “groan noise-free” brake linings, 3 sets from both vehicles were disassembled from the vehicles and shipped to the supplier for analysis. Measurements related to the friction coefficient are summarized in Figs. 19a and b for groan noise free and groan noise respectively.

According to test results, the friction coefficient µ of groan noise free brakes ranges between 2.09 and 2.52. On the other hand, the coefficient µ of the groan noise brakes shows fluctuations between 1.31 and 1.95. Because of the overcuring of the composite

b) Fig. 19. Measurements (friction coefficient, temperature and brake pressure) on a) groan noise-free and b) groan noise brake lining materials

As a result of this study, TVA and FMEA methods were employed to effectively determine the root cause of brake groan. The importance of these methods has been presented and shared with our suppliers in order to improve their plants and to help their business with other companies. The techniques used in this study also helped the suppliers to adopt a problem management approach and understand the solution flow chart.

6 CONCLUSIONS Companies must track their production, selling, and purchasing activities regularly to identify potential and actual problems. The detection and solution of the problem can then be carried out using industrial engineering tools such as TVA and FMEA methods.

These methods were used in this study to create an action plan for eliminating the groan noise generated by troubled drums in commercial light vehicles by

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determining the root cause. In our analysis, TVA showed something wrong in terms of warranty parts replacement and extra service costs. Then, using FMEA analysis, we could determine the root cause of the groan noise braking by elimination secondary and tertiary reasons. On the other hand, after performing the planned tests with the help of FMEA, the brake groan werer found to be generated by dynamic instability in the friction forces that act as nonconservative forces. In brake groan noise problems, displacements of the friction surfaces of the lining are caused by vibration due to frictional forces applied to deformed surfaces of the contact areas. Therefore, the amount of work by brake vibration structures often differs between the forward and backward vibration motions, resulting in dynamic instability of the system and some brake groan noise. In this study, the brake groan noise was found to depend on the magnitude of the friction coefficient and on the position of the contact areas between the friction material and the metal surface of the drum. Because of the overcuring of the composite material under heat, the surface of the lining material gets excessively crystallized, which decreases braking efficiency. Due to a loss of braking effects on contacting with drum’s internal surface, the lining works as pressure applicator to the drum only and this results in groan noise due to vibration under dry friction.

7 REFERENCES [1] Cantoni, C., Cesarini, R., Mastinu, G., Rocca, G., Sicigliano, R. (2009) Brake Comfort: A Review.

Vehicle System Dynamics, vol. 47, no. 8, p. 901-947, DOI:10.1080/00423110903100432.

[2] Betella, M., Harrison, F., Sharp, R.S. (2002).

Investigation of automotive creep groan noise with distributed source excitation technique. Journal of Sound and Vibration, vol. 255, no. 3, p. 531-547, DOI:10.1006/jsvi.2001.4178.

[3] Brecht, J., Schiffner, K. (2001). Influence of Friction Law on Brake Creep-Groan. SAE Technical paper, paper no. 2001-01-3138.

[4] Jang, H., Lee, J.S., Fash, J.W. (2001). Compositional effects of the brake friction material on creep groan phenomena. Wear, vol. 251, p. 1477-1483, DOI:10.1016/S0043-1648(01)00786-4.

[5] Luciano, M.A., Madruga, O., Costa, C. (2005). A Method for Measuring Creep Groan Based on Brake

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Inertial Dynamometer, SAE Technical Paper, paper no.

2005-01-4126, DOI:10.4271/2005-01-4126.

[6] Severin, D., Dörsch, S. (2001). Friction mechanism in industrial brakes. Wear, vol. 249, no. 9, p. 771-779, DOI:10.1016/S0043-1648(01)00806-7.

[7] Ahmed, I. (2006). Studying the Contact Analysis Behavior of Vehicle Drum Brake Using Finite Element Methods. SAE Technical Paper Series, paper no. 200601-3561, DOI:10.4271/2006-01-3561.

[8] Crowther, A.R., Yoon, J., Singh, R. (2007). An Explanation for Brake Groan Based on Coupled BrakeDriveline System Analysis. SAE International, paper no. 2007-01-2260, Noise and Vibration Conference and Exhibition, St. Charles, DOI:10.4271/2007-01-2260.

[9] Abdelhamid, M.K., Bray, W. (2009). Braking Systems Creep Groan Noise: Detection and Evaluation.

SAE Technical Paper, paper no. 2009-01-2103.

DOI:10.4271/2009-01-2103.

[10] Yoon, K.W., Lee, J.C., Cho, S.S. (2011). The Study of Vehicle Structural Characteristics for Creep Groan Noise. SAE International, paper no. 2011-01-2363.

[11] Jung, T., Chung, S. (2012). Research for Brake Creep Groan Noise with Dynamometer. SAE International Journal of Passenger, Cars – Mechanical Systems, vol.

5, no. 4, DOI:10.4271/2012-01-1824.

[12] Korunović, N., Trajanović, M., Stojković, M., Mišić, D., Milovanović, J. (2011). Finite Element Analysis of a Tire Steady Rolling on the Drum and Comparison with Experiment. Strojniški vestnik - Journal of Mechanical Engineering, vol. 57, no. 12, p. 888-897, DOI:10.5545/ sv-jme.2011-124.

[13] Lovrec, D., Kastrevc, M. (2011). Modeling and Simulating a Controlled Press-Brake Supply System.

International Journal of Simulation Modelling, vol. 10, no. 3, p.133-144, DOI:10.2507/IJSIMM10(3)3.184.

[14] Zhou, M., Wang, Y., Huang, Q. (2007). Study on the stability of drum brake non-linear low frequency vibration model. Archive of Applied Mechanics, vol.

77, no. 7, p.473-483, DOI:10.1007/s00419-006-0109-6.

[15] Ganguly, S., Pastor, K., Folta, G., Ianka, R., Karpenko, Y. (2011). Reduction of Groan and Grind Noise in Brake Systems. SAE International, paper no. 2011-012364, DOI:10.4271/2011-01-2364.

[16] Crow, K.A. (2002). Value analysis and Function Analysis System technique, from http://www.npdsolutions.com/va.html, accessed on 2013-01-01.

[17] AIAG (2008). Potential Failure Mode and Effects Analysis (FMEA). Automotive Industry Action Group (AIAG), 4th Ed., Southfielt [18] Stamatis, D.H. (2003). Failure Mode and Effect Analysis: FMEA from Theory to Execution. 2nd Ed., revised and expanded. ASQC Quality Press, Milwaukee.

Karabay, S. – Baynal, K. – İğdeli, C.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 387-399 © 2013 Journal of Mechanical Engineering. All rights reserved.

DOI:10.5545/sv-jme.2012.541

Original Scientific Paper

Received for review: 2012-04-21 Received revised form: 2012-10-29 Accepted for publication: 2013-02-18

Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths Peng 1 Tsinghua

Gao, P. – Yan, S. – Xie, L. – Wu, J. – Shaoze Yan1,* – Liyang Xie2 – Jianing Wu1

Gao1

University, Department of Mechanical Engineering, State Key Laboratory of Tribology, China 2 Northeastern University, School of Mechanical Engineering, China

Owing to the randomness of load applied to mechanical components, it is difficult to accurately determine the strength degradation path.

Therefore, the distribution of strength at each load application is always used to deal with the uncertainty of strength in its degradation process, which may cause errors in reliability calculations due to neglecting the correlation with the remaining strength at each load application in a strength degradation path. To deal with this problem, dynamic reliability models of mechanical components with the failure mode of fatigue are developed in this paper, based on equivalent strength degradation paths, whose uncertainty is determined by both the distribution of material parameters and the distribution of load. The proposed models can be used to quantitatively analyse the influences of the variation in statistical parameters of material parameters on the reliability and failure rate of components. Explosive bolts, which are important to the successful launch of satellites, have been chosen as representative examples to validate the effectiveness and accuracy of the proposed models. The results show that using strength distribution at each load application may lead to large errors in calculating reliability. Moreover, different material parameters have different influences on dynamic characteristics of reliability and on the failure rate of mechanical components.

Keywords: dynamic reliability, correlation, remaining strength, mechanical components, strength degradation path

0 INTRODUCTION In the design stage of mechanical components, the uncertainty in both the environmental load and the material parameters needs to be taken into consideration and a safety margin is required to guarantee the intrinsic reliability of mechanical components. The safety factor is comprehensively used in the practical design of mechanical components, which is industry specific and determined by the experience of engineers. However, the empirical safety factor cannot quantify the uncertainty and risk in mechanical design. Therefore, reliability analysis of mechanical products has gained more and more popularity [1] to [3].

Reliability is defined as the probability that a product performs its intended functions without failure during a specified time period. For mechanical components, the load–strength interference (LSI) model is the most important analytical method in reliability assessment. The conventional LSI model is essentially a static reliability model. However, gradual failure of mechanical components commonly exists in practical engineering due to the strength degradation caused by corrosion, wear, erosion creep, etc. As pointed out by Martin, constructing reliability models considering strength degradation is an important issue for reliability estimation and further research on generalized methods for the dynamic reliability analysis of mechanical components is imperative [4].

To overcome the shortcomings of conventional LSI models, reliability models based on stochastic process theory are investigated in which load and strength are modelled as two stochastic processes.

Lewis [5] analysed the time-dependent behaviour of a 1-out-of-2: G redundant system by combining the LSI model with a Markov model. Geidl and Saunders[6] introduced time-dependent elements into the reliability equation to estimate the reliability. Somasundaram and Dhas[7] put forward a generalized formula to estimate the reliability of a dynamic parallel system, in which components equally shared the load. Noortwijk and Weide [8] developed a reliability model, in which load and strength are described as two stochastic processes. Labeau et al. proposed the framework of a dynamic reliability platform and identified its main constituents [9]. Zhang et al. analysed the main methods for dynamic reliability estimation of nuclear power plants, which include discrete dynamic event trees and Monte Carlo simulation [10]. Slak analysed production planning and scheduling, cutting tools and material flow process, and manufacturing capacities [11]. Barkallah et al. proposed a method for process planning to determine the tolerance for manufacturing with statistical tools [12].

As a matter of fact, the reliability models based on stochastic process theory, such as the timedependent model, Markov model, etc, are the most important tools for dynamic reliability analysis.

The Markov models are mainly used for dynamic reliability analysis of electronic elements and multi-

*Corr. Author’s Address: Tsinghua University, Department of Mechanical Engineering, State Key Laboratory of Tribology, Beijing, China, yansz@tsinghua.edu.cn

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state systems. In the Markov models, the elements and the systems are characterised by different states at different time instants and the dynamic reliability is calculated based on the state transition matrix between two different time instants. However, it is very difficult to define and diagnose the states of mechanical components. In addition, the strength of mechanical components degrades under the application of environmental load. The state-based reliability models cannot be used for further investigation of how stress and material parameters influence the reliability of mechanical components.

In addition, much progress has been made in dynamic reliability analysis based on time-dependent models. The essence of the time-dependent models based on stochastic process theory is to calculate reliability by assuming the stress process and the strength degradation process to be two continuous processes, which are mathematically expressed as various stochastic processes. However, for mechanical components in the failure mode of fatigue, some problems exist in directly extending these reliability models to the dynamic reliability analysis of mechanical components and these are listed as follows: (1) For mechanical components with the failure mode of fatigue, the strength degradation process is discontinuous. In this case, it is meaningless to analyse the reliability at a given time instant. A more detailed explanation will be given in Section 1. The dynamic reliability analysis should be performed for a given time interval or at a specified application of load.

Therefore, it is more straightforward and reasonable to establish dynamic reliability models with respect to load application times, which can provide the basis for developing dynamic reliability models with respect to time. However, dynamic reliability models with respect to load application times considering strength degradation have not been widely used.

(2) For the convenience of developing a dynamic reliability model, various stochastic processes are assumed to model the strength degradation process in the time-dependent models without further explanation of the physical significance of the parameters involved in the strength processes. However, these proposed dynamic reliability models cannot be used to analyse how the reliability is influenced by the variation in the statistical parameters of material parameters.

(3) Due to the randomness of load magnitude at each application, it is difficult to determine the strength degradation path. Therefore, the distribution of strength at each time instant or each load application is always used for reliability calculations in the time388

dependent models, which can be derived based on the properties of the assumed strength degradation process. However, it may cause large errors to calculate reliability in this way due to neglecting the correlation between the remaining strength at each load application, something that is not pointed out in current literature.

To deal with these problems, this paper proposes dynamic reliability models that take into account the degradation mechanism of mechanical components and that can be used to quantitatively analyse the influence of the variation in the statistical characteristics of material parameters on the dynamic behaviour of reliability and failure rate of mechanical components. In the proposed models, the stress, strength, and load application times are modelled as random variables. Moreover, the proposed reliability models are established based on an analysis of the strength degradation path rather than on the distribution of strength at each load application.

1 RELIABILITY MODELS WITH RESPECT TO APPLICATION TIMES OF RANDOM LOAD Unlike the failure mode of corrosion, when only the failure mode of fatigue is taken into consideration, the load process is a discrete process. The assumption that the statistical characteristics of load are continuous functions of time indicates that there are infinite times of load application in an infinitesimal time interval ∆t, which is obviously unreasonable in reality. Therefore, the load process should be characterised by load application times and magnitude of load. In addition, under the failure mode of fatigue, strength does not degrade in the absence of load application and the change in strength is discontinuous as shown in Fig. 1.

Fig. 1. Schematic process of strength degradation

From Fig.1, it can be seen that at any time instant between two adjacent load applications, the reliability

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of the components is equal to one, which is different from the reliability in a given time interval. Hence, the reliability analysis at a given time instant is meaningless for the reliability analysis of mechanical components with the failure mode of fatigue. It is more straightforward and reasonable to model strength as a function of load application times rather than directly as the function of time. However, dynamic reliability models with respect to load application times considering strength degradation are not widely used. In this section, the dynamic reliability models of mechanical components with respect to load application times are developed, which provide the basis for dynamic reliability analysis with respect to time. Besides, the influences of the statistical characteristics of load and material parameters on reliability and failure rate are analysed.

1.1 Reliability Models with Respect to Load Application Times Due to the randomness of magnitude of load at each application, it is difficult to describe the degradation path of strength. Therefore, the method of calculating dynamic reliability based on the strength distribution at each load application, which is determined according to the assumed stochastic process of strength degradation, is adopted with all the features. However, this may lead to wrong estimates in calculating the reliability using the strength distribution at each load application, because some impossible strength degradation paths are included in the reliability calculation. For illustrative convenience, we list some possible strength degradation paths in Fig. 2. The uncertainty of the strength degradation path comes from the randomness of load magnitude at each load application. In Fig. 2, a circle represents a possible changing point of strength at a load application.

Therefore, the strength degradation path can be characterised by the changing points. The possible paths in Fig. 2 are summarised in Table 1.

Table 1. Strength degradation path Strength degradation path r0-1-3-7 r0-1-3-8 r0-1-4-9 r0-1-4-10 r0-2-5-11 r0-2-5-12 r0-2-6-13 r0-2-6-14

t1 1 1 1 1 2 2 2 2

Changing point t2 3 3 4 4 5 5 6 6

t3 7 8 9 10 11 12 13 14

From Table 1, it can be seen that there are two possible changing points of strength (1-2), four possible changing points of strength (3-6) and eight possible changing points (7-14) at t1, t2 and t3, respectively. When calculating reliability by using the strength distribution at each load application, paths with all possible combinations of changing points of strength are taken into account, which also include impossible paths, such as path r0-1-6-10, path r0-2-4-12, etc. Therefore, the method based on the distribution of strength at each load application may result in errors in the assessment of reliability, which will be illustrated later. An alternative method for dynamic reliability analysis is to perform a Monte Carlo simulation, in which the strength degradation is simulated based on the degradation mechanism of mechanical components and the random loads are generated according to their probability distributions.

However, the Monte Carlo simulation is considerably time consuming and the time for a Monte Carlo simulation increases rapidly with the increase in load application times, which limits the practical applicability of this simulation. In addition, the Monte Carlo simulation method cannot be used to quantitatively analyse the influences of the variation in statistical parameters of material parameters on the reliability and failure rate of mechanical components.

To deal with these problems, dynamic reliability models with respect to load application times are developed in this section, which can then be used to calculate the reliability of mechanical components using the failure mode of fatigue under the application of random load for arbitrary times.

The strength degradation path for a given sample mechanical component is deterministic. In general, the remaining strength of the mechanical components can be expressed in the following form [13].

r (n) = r0 [1 â&#x2C6;&#x2019; D(n)]a , (1)

Fig. 2. Strength degradation path Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths

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where n and are the load application times and initial strength, respectively, and a is the material parameter.

D(n) is the cumulative damage caused by load, which is determined by both load application times n and the magnitude of load. According to the Miner linear damage accumulation rule [14], the damage caused by a load with a magnitude of once is:

1 Di (ni ) = , (2) Ni

where Ni is the lifetime of a component under a load with a magnitude of si. Analogously, the damage caused by a load with a magnitude of s0 once is:

D0 (1) = 1 / N 0 , (3)

where N0 is the lifetime of a component under the load with the magnitude of s0. According to the S-N Curve theory of components, the relationship between the load si applied to a components and the corresponding lifetime Ni of the component under the application of si can be mathematically expressed as follows:

m

si N i = C , (4)

where m and C are material parameters and the dispersion of lifetime is represented by the dispersion of the parameter C. Similarly, the relationship between and N0 can be written as:

s0 m N 0 = C. (5) From Eq. (4) and Eq. (5), it can be derived that:

Di (= 1)

1 sm = i , (6) Ni C

and

D0 (= 1)

m

s 1 = 0 . (7) N0 C

Therefore, from Eqs. (6) and (7), it can be derived that the damage caused by a load with a magnitude of si once is equal to the damage caused by a load with a magnitude of s0 for ni0 times under the condition that:

s ni 0 = ( 1 ) m . (8) s0

According to the total probability theorem, the damage caused by the application of a random load with a probability density function (pdf) of fs(s) once 390

can be approximated by the damage caused by a magnitude of s0 load applied n0 times, which is given by: 1 ∞

n0 = m ∫ s m f s ( s )ds. (9) s0 −∞ Therefore, for a deterministic initial strength, the remaining strength in an equivalent strength degradation path can be expressed according to Eq. (1) as follows: nn r (n) = r0 [1 − D(n)]a = r0 (1 − 0 ) a = N0 ∞

= r0 (1 −

n ∫ s m f s ( s )ds −∞

C

)a .

(10)

For a given component with deterministic initial strength r0 and material parameter C, the reliability under the application of random load n times can be derived as: ∞ i ∫ s m f s ( s )ds n −1 r0 (1− −∞ )a C R ( n) = ∏ [ ∫ f s ( s )ds ]. (11)

i =0

−∞

To consider the randomness of the initial strength and the material parameter C, we denote the pdf of r0 and C by fr0(r0) and fC(C), respectively. According to the Bayes law for continuous variables, the reliability of components with respect to load application times considering strength degradation can be expressed as follows: i ∫ s m f s ( s )ds  n−1  r0 (1− −∞ )a   C f C (C ) ∏ [ ∫ f s ( s )ds ] dCdr0 . −∞ i = 0     ∞

R ( n) = ∫

∞

−∞

f r0 (r0 ) ∫

∞

−∞

(12)

According to the definition of failure rate, the failure rate of components with respect to load application times can be written as: h( n) =

F (n + 1) − F (n) = R ( n)

∞  i ∫ s m f s ( s )ds  n−1  ∞ r0 (1− −∞ )a  ∞   C =  ∫ f r0 (r0 ) ∫ fC (C ) ∏ [ ∫ f s ( s )ds ] dCdr0 − −∞ −∞ −∞   i =0    

∞

∞

−∞

−∞

− ∫ f r0 (r0 ) ∫

∞  i ∫ s m f s ( s )ds  n  r0 (1− −∞ )a    C f s ( s )ds ] dCd r0  / f C (C ) ∏ [ ∫ −∞  i =0     

∞   i ∫ s m f s ( s )ds  n−1  ∞ r0 (1− −∞ )a  ∞    C / f ( r ) f ( C ) [ f ( s ) s d ] dCdr  ∫−∞ r0 0 ∫−∞ C ∏ ∫−∞  0  . (13) s i = 0        

In the case where strength degradation does not take place, Eq. (12) degenerates into the following form:

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∞

r0

−∞

−∞

R(n) = ∫ f r0 (r0 )[ ∫ f s ( s )ds ]n dr0 . (14)

Specifically, when the value of n is equal to 1, Eq. (14) reduces to the conventional LSI model.

1.2 Numerical Examples In this section, explosive bolts are chosen as representative examples to illustrate the proposed reliability models. Explosive bolts are of considerable importance to the successful launch of satellites as a connection and pyrotechnical separation device. The structure of the explosive bolt is shown in Fig. 3 [15].

In the launch process of satellites, the explosive bolt is used for the connection between the payload adapter and the interface ring, which are connected to the launch vehicle and the satellite, respectively. In the departure process of satellites and launch vehicles, the explosive bolt fractures at the site where a groove is designated on the outer cylinder of the explosive bolt with the help of an internal power source generated by an explosive charge. The launch process of satellites lasts longer than the duration of the departure process so that failure due to the strength degradation of the explosive bolt may occur. In this section, we will concentrate on a dynamic reliability analysis of explosive bolts in the launch process of satellites.

bolts as shown in Fig. 4. The distribution of initial strength can be obtained by experimental tests [11].

The method of constructing the finite element model of bolted joints and the statistical method for the distributions of random variables is found in reference [16]. In addition, Crocombe developed a method for estimating the energy dissipated in the bolted joints of a satellite structure[17]. Nethercot investigated the behaviour of stainless steel bolted connections using the finite element models [18]. Oskouei analysed the stress of aircraft structural double-lapbolted joints by using the finite-element method [19]. Nassar proposed a method of calculating the bearing friction forces on the rotating contact surface of threaded fasteners [20].

In this section, we focus on validating the proposed dynamic reliability models and analysing the influences of the variation in statistical characteristics of material parameters on the reliability and failure rate of explosive bolts.

Fig. 4. Finite element model of an explosive bolt

Fig. 3. Structure of the explosive bolt

Due to the randomness of the environmental load in the launch process of satellites and the variation in the manufacturing process of explosive bolts, the stress and the initial strength should be modelled as random variables. In the launch process of satellites, the explosive bolts are mainly used to withstand the random environmental axial forces while the shear forces are sustained by other mechanical components [11]. The distribution of stress can be obtained through a finite element analysis (FEA) of explosive

The material parameters of the explosive bolts are given by m = 2, α = 1 and C = 109 MPa2. The initial strength of the explosive bolt is normally distributed with a mean value of μ(r0) and a standard deviation of σ(r0). The stress at each load application follows the normal distribution with a mean value of μ(s) and a standard deviation of σ(s). The value of the mean value and standard deviation of the initial strength and stress are listed in Table 2.

Table 2. Statistical parameters of stress and initial strength μ(r0) [MPa] 600

σ(r0) [MPa] 20

μ(s) [MPa] 500

σ(s) [MPa] 20

In order to validate the reliability model proposed in Section 1.1, a Monte Carlo simulation for evaluating

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the dynamic reliability of explosive bolts is carried out in this section and the flowchart of the Monte Carlo simulation is shown in Fig. 5. In the Monte Carlo simulation, the stress at each load application is generated according to its distribution and the strength degradation of an explosive bolt sample is simulated based on the degradation mechanism and all the stress generated in its strength degradation path.

Therefore, the Monte Carlo simulation represents the actual strength degradation process of explosive bolts. Additionally, the distribution of strength at each load application can be obtained based on Eq. (10).

To illustrate the errors that can be caused by using the distribution of strength at each load application in the calculation of reliability, the reliability calculated by using the proposed model, the reliability calculated according to the strength distribution at each load application, and the reliability from the Monte Carlo simulation are shown in Fig. 6.

Start Determine k and n; set i=0, m=1, j=1 Generate random initial strength r0; set rj=r0 Generate random load sj m=m+1; set j=1

j=j+1

Calculate remaining strength r; set rj=r

strength at each load application in a strength degradation path and takes into account non-existent strength degradation paths.

Fig. 6. Comparison between proposed method and Monte Carlo simulation

In addition, to analyse the influences of the variation in statistical parameters of material parameters on the reliability and failure rate of explosive bolts, consider the following four cases: Case 1: The material parameters of the explosive bolts are given by m = 2, α = 1 and r0 = 600 MPa. The statistical parameters of stress and C are listed in Table 3. The reliability and failure rate of the explosive bolts with different mean values of C are shown in Figs. 7 and 8, respectively.

No rj>sj ?

Yes

Table 3. Statistical parameters of stress and material parameters C of explosive bolts

j=n ?

No i=i+1

Yes

1 2 3

No m=k ?

Yes

σ(s) [MPa] 20 20 20

μ(C) [MPa2] 109 1.5×109 2×109

σ(C) [MPa2] 106 106 106

Case 2: The material parameters of the explosive bolts are given by m = 2, α = 1 and r0 = 600 MPa. The statistical parameters of stress and C are listed in Table 4. The reliability and failure rate of the explosive bolts with different standard deviations of C are shown in Figs. 9 and 10, respectively.

R=1-i/k

End

Fig. 5. Flowchart of Monte Carlo simulation

From Fig. 6, we can see that the reliability calculated by using the proposed method in this paper shows good agreement with the results obtained from the Monte Carlo simulation. However, it may result in erroneous calculation of the reliability by using the distribution of strength at each load application, which neglects the correlation between the remaining 392

μ(s) [MPa] 500 500 500

Table 4. Statistical parameters of stress and material parameters C of explosive bolts 1 2 3

Gao, P. – Yan, S. – Xie, L. – Wu, J.

μ(s) [MPa] 500 500 500

σ(s) [MPa] 20 20 20

μ(C) [MPa2] 109 109 109

σ(C) [MPa2] 106 5×106 107

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 387-399

Fig. 7. Reliability of explosive bolts with different mean values of C

Fig. 9. Reliability of explosive bolts with different dispersions of C

Fig. 8. Failure rate of explosive bolts with different mean values of C

Fig. 10. Failure rate of explosive bolts with different dispersions of C

Case 3: The material parameters of the explosive bolts are given by m=2, α=1 and C=109 MPa2. The statistical characteristics of both stress and initial strength are listed in Table 5. The reliability and failure rate of the explosive bolts with different mean values of initial strength are shown in Figs. 11 and 12, respectively.

rate of the explosive bolts with different standard deviation of initial strength are shown in Figs. 13 and 14, respectively.

Table 5. Statistical parameters of stress and initial strength of explosive bolts 1 2 3

μ(r0) [MPa] 550 600 650

σ(r0) [MPa] 30 30 30

μ(s) [MPa] 500 500 500

σ(s) [MPa] 20 20 20

Case 4: The material parameters of the explosive bolts are given by m = 2, α = 1 and C = 109 MPa2.

The statistical characteristics of both stress and initial strength are listed in Table 6. The reliability and failure

Table 6. Statistical parameters of stress and initial strength of explosive bolts 1 2 3

μ(r0) [MPa] 600 600 600

σ(r0) [MPa] 20 30 40

μ(s) [MPa] 500 500 500

σ(s) [MPa] 20 20 20

Case 5: The material parameters of the explosive bolts are given by m=2, α=1 and C=109 MPa2. The statistical parameters of stress and r0 are listed in Table 7. The reliability of the explosive bolts with different dispersions of stress are shown in Fig. 15.

Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths

393

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 387-399

Fig. 11. Reliability of explosive bolts with different mean values of initial strength

Fig. 13. Reliability of explosive bolts with different dispersions of initial strength

Fig. 12. Failure rate of explosive bolts with different mean values of initial strength

Fig. 14. Failure rate of explosive bolts with different dispersions of initial strength

Table 7. Statistical parameters of stress and material parameters C of explosive bolts 1 2 3

μ(s) [MPa] 500 500 500

σ(s) [MPa] 10 20 30

μ(r0) [MPa] 600 600 600

σ(r0) [MPa] 30 30 30

From Figs. 7 to 12, it can be inferred that the mean value of the initial strength and C has a strong influence on the reliability and failure rate of explosive bolts. The reliability increases and the failure rate decreases with the increase in the mean value of the initial strength and C. In addition, the dispersion of C has a negligible influence on the reliability and failure rate of explosive bolts and can therefore be neglected in the reliability analysis of explosive bolts. Thus, Eqs. (12) and (13) can be rewritten as follows: 394

i ∫ s m f s ( s )ds  n−1  r0 (1− −∞ )a   C f r0 (r0 ) ∏ [ ∫ f s ( s )ds ] dr0 , −∞  i =0    ∞

R ( n) = ∫

∞

−∞

∞  i ∫ s m f s ( s )ds  n−1  r0 (1− −∞ )a  ∞   C h(n) =  ∫ f r0 (r0 ) ∏ [ ∫ f s ( s )ds ] dr0 − −∞ −∞  i =0      ∞ m  i ∫ s f s ( s )ds  n  ∞ r0 (1− −∞ )a    C f s ( s )ds ] dr  / − ∫ f r0 (r0 ) ∏ [ ∫ −∞ −∞ 0 i =       ∞  i ∫ s m f s ( s )ds  n−1   r0 (1− −∞ )a  ∞    C /  ∫ f r0 (r0 ) ∏ [ ∫ f s ( s )ds ] dr0  . −∞ −∞  i =0       

In addition, it is traditionally considered that a large dispersion leads to lower reliability. However,

Gao, P. – Yan, S. – Xie, L. – Wu, J.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 387-399

from Figs. 13 and 14, it can be seen that the dispersion of initial strength has different influences on the reliability and failure rate of explosive bolts at different stages of their lifetime. This is because during the initial period of lifetime, the reliability is high and a large dispersion of initial strength increases the possibility that the remaining strength will have a small value, which leads to low reliability. At the stage of lifetime when the reliability is low, a large dispersion of initial strength increases the possibility that the remaining strength takes a large value, which leads to a high reliability.

2.1 Dynamic Reliability Models with Respect to Time As mentioned above, for mechanical components with the failure mode of fatigue, the assumption of continuous statistical characteristics of load with respect to time indicates infinite times of load application in an infinitesimal time interval ∆t, which is unreasonable in reality. The load process should be characterised by the load application times and the magnitude of load. Therefore, the reliability models proposed in Section 1.1 provide a basis for developing reliability models with respect to time. When the relationship between load application times and time is obtained, dynamic reliability of components with respect to time can be further developed based on the models proposed in Section 1.1. Provided that the load application times in a time interval is deterministic, which is expressed by using a deterministic function n = ft (t) with respect to time, the dynamic reliability of mechanical components can be calculated as follows: i ∫ s m f s ( s )ds  f (t )−1  r0 (1− −∞ )a t  C f r0 (r0 )  ∏ [ ∫ f s ( s )ds ] dr0 . (15) −∞  i =0    ∞

R(t ) = ∫

∞

−∞

Fig. 15. Reliability of explosive bolts under stress with different standard deviations

From Fig. 15, it can be inferred that the dispersion of stress has a strong influence on the dynamic reliability of explosive bolts. The reliability decreases with the increase in the dispersion of stress. This is because a large dispersion of stress increases the possibility that the stress will exceed the remaining strength in the load application process.

2 DYNAMIC RELIABILITY ANALYSIS OF MECHANICAL COMPONENTS WITH RESPECT TO TIME In this section, dynamic reliability models with respect to time are established based on the models proposed in Section 1.1, which take account of the failure mechanism and the stochastic strength degradation path. Moreover, numerical examples are given to analyse the influences of the statistical characteristics of initial strength on the dynamic behaviour of reliability and on the failure rate of mechanical components.

However, in some situations, the occurrence times of load are random, which can only be analysed by using the stochastic process theory. In practice, it has been proved that the Poisson process is an effective stochastic process to describe the random occurrence times of random load in a time interval. According to the Poisson process theory, the probability that the random load appears n times in a time interval t can be given by reference [6]: t

Pr[n(t ) − n(0) = n] =

( ∫ λ (t )dt ) n 0

n!

t

exp(− ∫ λ (t )dt ), (16) 0

where λ(t) is the intensity of the Poisson process. For a mechanical component with deterministic initial strength r, according to the total probability theorem, the reliability in the time interval of t can be expressed as follows: ∞

R(t ) = ∑ P(n(t ) = k )R(k ) = k =0

t

∞

= exp(− ∫ λ (t )dt ) + ∑ 0

t

( ∫ λ (t )dt ) n 0

n =1

n!

×

i ∫ s m f s ( s )ds  n−1  r (1− −∞ )a   C f s ( s ) ds ]  . × exp(− ∫ λ (t )dt ) ∏ [ ∫ 0 −∞  i =0    ∞

t

Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths

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When considering the distribution of initial strength characterised by its pdf of fr(r), the reliability can be obtained by using the Bayes law for continuous variables as follows: R(t ) = ∫

∞

−∞

∞

t

f r (r ) exp(− ∫ λ (t )dt ) + ∑ 0

t

( ∫ λ (t )dt ) n

n =1

0

n!

×

i ∫ s m f s ( s )ds  n−1  r (1− −∞ )a   C × exp(− ∫ λ (t )dt ) ∏ [ ∫ f s ( s )ds ] dr. (17) 0 −∞  i =0    ∞

t

statistical characteristics of stress and initial strength are listed in Table 8. The reliability calculated by using the models proposed in this paper and the reliability calculated based on the distribution of strength at each load application are shown in Fig. 16.

Table 8. Statistical parameters of stress and initial strength of explosive bolts μ(r0) [MPa] 400

σ(r0) [MPa] 30

μ(s) [MPa] 300

σ(s) [MPa] 20

Correspondingly, the failure rate of the component can be written as: t n −1   ∞ ( λ (t )dt ) ∞   ∫ 0  n − t λ (t )dt  × h(t ) = λ (t ) ∫ f r (r ) 1 − ∑ −∞  ∫0  n !   n=1  ∞ i ∫ s m f s ( s )ds  n−1   r (1− −∞ )a    C f s ( s )ds ] dr  / × ∏ [ ∫ −∞  i =0      t n   ∞ ( λ (t )dt )  ∞  ∫ × /  ∫ f r (r ) 1 + ∑ 0 −∞ n! = 1 n    ∞ i ∫ s m f s ( s )ds  n−1   r (1− −∞ )a    C f s ( s )ds ] dr  . × ∏ [ ∫ −∞  i =0     

Time t [h]

(18)

Although the dynamic reliability model with respect to time is developed based on the theory of the Poisson process, it should be noted that the reliability model proposed in section 2.1 can be easily extended to other dynamic reliability models with respect to time, as long as the statistical characteristics of the load application times are obtained. In addition, from the derivation of Eqs. (17) and (18), it can be seen that the proposed dynamic reliability model with respect to time takes the stochastic strength degradation path of mechanical components into consideration. To illustrate the error caused by estimating reliability based on the strength distribution at each load application, numerical examples will be given in the following section.

2.2 Numerical Examples Consider the explosive bolts operating under the application of random load, whose occurrence times follow the Poisson process with an intensity of 0.6 h–1.

The initial strength and the stress follow the normal distribution. The material parameters of the explosive bolts are given by m = 2, α = 1 and C = 108 MPa2. The 396

Fig. 16. Comparison between the reliability calculated by using the proposed method and the reliability calculated based on the strength distribution at each load application

From Fig. 16, we can see that the proposed reliability models can be used to represent the variation in reliability with time. In addition, it leads to lower reliability to use the distribution of strength at each load application. The error is caused by taking into account strength degradation paths which do not exist in practice. Therefore, dynamic reliability models with respect to time should be developed based on the strength degradation path rather than on the distribution of strength at each load application.

In order to analyse the influences of the mean value and dispersion of initial strength on the reliability and failure rate of explosive bolts, consider the following two cases: Case 1: The material parameters of the explosive bolts are given by m = 2, α = 1 and C = 108 MPa2.

The statistical characteristics of both stress and initial strength are listed in Table 9. The reliability and failure rate of the explosive bolts with different mean values of initial strength are shown in Figs. 17 and 18, respectively.

Case 2: The material parameters of the explosive bolts are given by m = 2, α = 1 and C = 108 MPa2.

Gao, P. – Yan, S. – Xie, L. – Wu, J.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 387-399

Time t [h]

Fig. 17. Reliability of explosive bolts with different mean values of initial strength

Time t [h]

Fig. 19. Reliability of explosive bolts with different dispersions of initial strength

Time t [h]

Time t [h]

Fig. 18. Failure rate of explosive bolts with different mean values of initial strength

Fig. 20. Failure rate of explosive bolts with different dispersions of intial strength

The statistical characteristics of both stress and initial strength are listed in Table 10. The reliability and failure rate of the explosive bolts with different standard deviations of initial strength are shown in Figs. 19 and 20, respectively.

of the variation in statistical parameters of material parameters on the reliability and failure rate. The reliability increases and failure rate decreases with the increase in the mean value of initial strength.

Moreover, the dispersion of initial strength has different influences on the reliability and failure rate of the explosive bolts at different periods of lifetime.

In addition, the bathtub curve is always used to represent the variation in the failure rate of mechanical components with time as shown in Fig. 21.

Table 9. Statistical parameters of stress and inistial strength of explosive bolts 1 2 3

μ(r0) [MPa] 350 400 450

σ(r0) [MPa] 30 30 30

μ(s) [MPa] 300 300 300

σ(s) [MPa] 20 20 20

From Figs. 17 to 20, we can see that the proposed dynamic reliability models can be used to analyse the dynamic characteristics of reliability and failure rate and quantitatively analyse the influence

Table 10. Statistical parameters of stress and initial strength of explosive bolts 1 2 3

μ(r0) [MPa] 400 400 400

σ(r0) [MPa] 20 30 50

Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths

μ(s) [MPa] 300 300 300

σ(s) [MPa] 20 20 20

397

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 387-399

Fig. 21. Schematic bathtub curve of mechanical components

From Figs. 18 and 20, it can be seen that the failure rate curve obtained from the reliability model developed in this paper is consistent with the bathtub curve theory. Furthermore, the slope of the failure rate curve for mechanical components in the random failure period tends to decrease with the increase in the mean value and the dispersion of initial strength.

3 CONCLUSIONS AND FUTURE WORK Dynamic reliability models based on strength degradation paths are developed in this paper.

Owing to the difficulty in mathematically describing the strength degradation path, which is caused by the randomness of loads applied to mechanical components, the distribution of strength at each load application is always used to analyse the dynamic reliability of the mechanical components. However, it may lead to large errors in the reliability calculation to treat the uncertainty of strength in this way due to neglecting the correlation with the remaining strength at each load application in a strength degradation path.

The proposed reliability models can be used to quantitatively analyse the influence of the variation in statistical parameters of material parameters on the dynamic characteristics of reliability and on the failure rate of mechanical components. Traditionally, it is considered that a large dispersion of initial strength results in a lower reliability. However, when considering strength degradation, the dispersion of initial strength has different influences on the reliability of mechanical components at different stages of lifetime. Moreover, it should be noted that the statistical characteristics of initial strength have strong influences on the dynamic behaviour of the failure rate of mechanical components. The slope of the failure rate curve for mechanical components in the random failure period tends to decrease with the 398

increase in the mean value and the dispersion of initial strength.

Further work is in progress to include other variables in the reliability models to achieve more accurate predictive results. Moreover, extensions of the proposed method to problems of reliability-based design optimisation are currently being investigated by the authors.

4 ACKNOWLEDGEMENTS This work was supported by the National Science Foundation of China under Contract No. 11072123, the National High Technology Research and Development Program of China (863 Program) under Contract No. 2009AA04Z401, the Major State Basic Research Development Program of China (973 Program), and the Project sponsored by SRF for ROCS, SEM. 5 REFERENCES [1] Dasic, P., Natsis, A., Petropoulos, G. (2008). Models of reliability for cutting tools: Examples in manufacturing and agricultural engineering. Strojniški vestnik – Journal of Mechanical Engineering, vol. 54, no. 2, p.

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[5] Levis, E.E. (2001). A Load-capacity interference model for common-mode failures in 1-out-of-2: G systems.

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IEEE Transactions on Reliability, vol. 50, no. 1, p. 4751, DOI:10.1109/24.935017.

[6] Geidl, V., Saunders, S. (1987). Calculation of reliability for time-varying loads and resistances. Structural Safety, vol. 4, no. 4, p. 285-292, DOI:10.1016/01674730(87)90003-8.

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Reliability of a dynamic n-unit shared load parallel system under different failure times. Microelectronics and Reliability, vol. 37, no. 5, p. 869-871, DOI:10.1016/S0026-2714(96)00100-X. [8] Van Noortwijk, J.M., Van der Weide, J.A.M., Kallen, M.J., Pandey, M.D. (2007). Gamma processes and peaks-over-threshold distributions for time-dependent reliability. Reliability Engineering and System Safety, vol. 92, no. 12, p. 1651-1658, DOI:10.1016/j. ress.2006.11.003.

[9] Labeaua, P. E., Smidts, C., Swaminathan, S. (2000).

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[12] Barkallah, M., Louati, J., Haddar, M. (2012).

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895-911, DOI:10.1016/j.apm.2006.03.020.

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Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 400-408 © 2013 Journal of Mechanical Engineering. All rights reserved.

DOI:10.5545/sv-jme.2013.959

Original Scientific Paper

Received for review: 2013-01-08 Received revised form: 2013-02-07 Accepted for publication: 2013-04-08

Selecting the Most Adaptable Work Equipment Berlec, T. – Kušar, J. – Rihar, L. – Starbek, M. Tomaž Berlec – Janez Kušar* – Lidija Rihar – Marko Starbek

University of Ljubljana, Faculty of Mechanical Engineering, Slovenia Selecting the most suitable work equipment is a key factor in creating the value of products; therefore, the design and selection of work equipment are increasingly focused on its adaptability. The adaptability of work equipment with regards to flexibility and responsiveness is defined by its universality, mobility, modularity, compatibility and economy. This article presents a methodology of evaluating the adaptability of work equipment and a procedure for selecting work equipment with the maximum adaptability index (i.e. the optimal one) from a group of available work equipment. The results of designing and weighing the criteria of the target tree are presented, as well as the results of coarse and fine selection of the optimal CNC lathe, considering its expected adaptability.

Keywords: adaptability, flexibility, responsiveness, universality, mobility, modularity, compatibility, target tree

0 INTRODUCTION The production programmes of small and medium enterprises are subject to continuous changes and ever increasing customer requests for the shortest possible delivery times, the best possible quality and the lowest possible product prices. Companies are faced with unstable demands regarding the types and quantities of products. Due to these demands, frequent changes of functions and the layout of work equipment are required. Problems with the adaptability of work equipment frequently occur.

Hernandez [1] asserted that adaptability is the potential of the company to carry out both the purchase of low-cost goal-oriented work equipment and the reconfigurations of it. Reinhar [2] and Zäh [3] understood adaptability to be an extension of flexibility. Nyhuis [4] said that adaptability is associated with additional costs of investment and consumption of time; however, the costs only apply when changes are carried out. Wiendahl [5] wrote that it is necessary to distinguish between five levels of adaptability: universality, mobility, modularity, compatibility and economy.

As a carrier of value creation, work equipment is the key factor of production. Today, the design and selection of work equipment are increasingly oriented towards its adaptability, mostly because of unreliable market forecasts [6] to [8].

The adaptability of work equipment is its ability to be adapted (at low cost) according to internal or external technological, structural, or organisational changes.

In general, adaptability consists of the flexibility and responsiveness of work equipment; this can be classified under universality, mobility, modularity, compatibility and economy.

400

Flexibility means that the work equipment is greater than needed with respect to the current functions, performance and accuracy. It allows the management of future, planned-in-advance scenarios.

Additional functions are available and can be activated when needed. Work equipment must be able to adapt to new circumstances and new needs at low costs [9].

The responsiveness of work equipment is the capability of reacting to new circumstances that were not foreseen in the planning phase.

Such a responsiveness is carried out by using its capability of being reconfigured. The desired result of responsiveness is modular, reconfigured work equipment.

When designing or selecting work equipment, it is necessary to take into account the requirements that must be met by it with regard to its technological functions and adaptability. An exact specification of the work equipment adaptability is required.

1 COMPONENTS OF WORK EQUIPMENT ADAPTABILITY Adaptability of work equipment in the form of flexibility and the responsiveness of work equipment is defined by the universality, mobility, modularity, compatibility and economy of work equipment [5].

The universality of work equipment refers to its design and dimensions so that it can carry out various tasks.

Today, work equipment is highly universal; e.g. a work piece can be entirely processed on a turningmilling centre.

The mobility of work equipment refers to its ability to be moved if the structure or layout of the factory is changed. This mobility is influenced by its mass, size and transport mode. For work equipment that cannot be transported as a whole, it is essential

*Corr. Author’s Address: University of Ljubljana, Faculty of Mechanical Engineering, Aškerčeva 6, Ljubljana SI-1000, Slovenia, janez.kusar@fs.uni-lj.si

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 400-408

that its components can be easily dismantled and reassembled afterwards.

The modularity of work equipment refers to modular design and standardisation of interfaces.

Interfaces have to fulfil the requirements of simple assembly and disassembly, precise positioning and high rigidity. Modularity must be carried out so that the buyers can change the work equipment by themselves.

The compatibility of work equipment refers to the possibility of its integration into the existing production structure.

The following interfaces are crucial for factory planning: control, IT and communication, mechanical, energy, user and material flow. In addition to connections, the term ‘interface’ also refers to the systems of the tools. The technology of connections is the decisive factor of compatibility. The producer of work equipment must sell only standard connections.

The extension of the available functions of work equipment always depends on the economy of doing so.

If extensions of functions were foreseen by the producer of the work equipment, the extension is usually economical; otherwise, it is not.

2 SYSTEM OF ASSESSMENT OF WORK EQUIPMENT ADAPTABILITY The results of the assessment of work equipment adaptability will be presented as indices of adaptability of work equipment K with regard to universality, mobility, modularity, compatibility and economy.

The target tree procedure can be used to assess the adaptability of a piece of work equipment [10]. This is a comparative procedure that allows an assessment of at least two work equipment units. An essential part of the target tree procedure is the hierarchy of criteria, i.e. the logical classification and grouping of criteria.

The target tree in Fig. 1 shows a general tree of criteria for the assessing adaptability of work equipment.

A tree of criteria is the result of teamwork [11] and [12]. On the first level of the target tree, there is a target criterion corresponding to the searched target index value of work equipment adaptability.

The following sub-criteria are on the second level of the target tree: universality, mobility, modularity, compatibility and economy [13]. On the third level of the target tree, there are the required basic criteria for which the schemes of fulfilment are made and from which the values of meeting the criteria di,j can be found. The schematics regarding the fulfilment

of criteria on the third level are usually set up in cooperation with the suppliers of work equipment. For each basic criterion on the third level of the target tree, a scheme regarding the fulfilment of the criterion must be formed and, on its basis, the value of fulfilment is determined.

Fig. 1. Hierarchy of criteria for assessing adaptability of work equipment

Fig. 2 presents an example of a scheme regarding the fulfilment of the ‘mass’ criterion.

Fulfilment of di,j criterion [points] 120 100 80 60 40 20 0

0 1

5

10

15

20

25 Mass [t]

di,j – the value of fulfilment of the ith criterion of the jth work equipment

Fig. 2. Scheme on fulfilment of the ‘mass’ criterion

It can be seen from Fig. 2 that if the mass of the work equipment is 1 t, the work equipment obtains 100 points, while it obtains only 10 points if its mass is 20 t.

The assessment of all criteria on the third level of the target tree is carried out in the same way as with the evaluation scale from 0 to 100 points.

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Adaptability criteria have different levels of importance; therefore, they must be assigned different weights, ui, on all hierarchical levels.

When determining the weight factors for criteria of work equipment adaptability, it is necessary to take care that the relations between criteria make sense. The transitivity rule is valid here, i.e. if the first criterion has a larger weight than the second criterion: u1 > u2 , and if the second criterion has larger weight than the third criterion: u2 > u3 , then the first criterion has larger weight than the third criterion: u1 > u3 .

The method of pair comparison is usually used for determining the weights of criteria [14].

It is necessary to consider the following when determining the weights of criteria, using the pair comparison method: • the criterion that is more important than the other one is circled, • if there are two equally important criteria, they are set within an oval, • the weight of a criterion is a quotient between the number of its advantages and the total number of advantages, • the sum of the weights of the criteria is 1, • the criterion with the maximum weight is the most important and is placed as first.

Mobility

100%

Economy

Universality

Compatibility

Modularity

total index of work equipment adaptability partial indices of work equipment adaptability

Fig. 3. Profile of work equipment adaptability

Known values of meeting the criteria on the third level of work equipment, di,j, and weights, ui, 402

allow the calculation of indices of work equipment adaptability regarding universality, mobility, modularity, compatibility, economy, and total index.

The results of determining the indices of work equipment adaptability are entered in a polar chart; such an example is shown in Fig. 3.

3 SELECTION OF THE OPTIMAL WORK EQUIPMENT WITH RESPECT TO ADAPTABILITY The known system of evaluating work equipment adaptability allows the selection of the work equipment with the highest adaptability index (i.e. the optimal work equipment) from a group of available units.

An analysis of optimisation methods showed that the benefit-analysis method would be the most useful for a selection of the optimal work equipment [15].

The cost-benefit analysis is the most commonly used method for decision making: it is extremely simple to use and gives the target values of the benefits.

This method requires data from suppliers of work equipment regarding the fulfilment of criteria, di,j, and the data from the buyer of the work equipment regarding weights of criteria, ui.

After the creativity workshop had been carried out [16] regarding the weights of criteria, it was concluded that the selection would be carried out in two phases.

In the first phase, a coarse selection of a smaller group of work equipment suppliers (up to five) that best meet the requirements according to the defined criteria would be made. The basic data for a coarse selection would be provided by normalised data on the fulfilment of the required criteria in the smaller group of work equipment suppliers (di,j equals 0 or 1).

In the second phase, a fine selection of work equipment would be carried out, i.e. the work equipment with the maximum benefit (maximum adaptability index) would be selected from the smaller group. Basic data for the fine selection would be provided by the data on the tests performed regarding the fulfilment of the di,j criteria in the smaller group of suppliers (0 points ≤ di,j ≤ 100 points).

In theory, there are many possible algorithms for multi-objective optimization [17]. We decided to use the weighted sum method, because of its simplicity.

Berlec, T. – Kušar, J. – Rihar, L. – Starbek, M.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 400-408

3.1 Coarse Selection of the Smaller Group of Work Equipment During coarse selection of work equipment, whether each unit meets the target tree criteria can be determined by using the data from the suppliers.

For the jth offered work equipment, an ntuple of numbers regarding the fulfilment of criteria of the work equipment is created:

 j work equipment fulfils 1, the i th criterion  , = j th work equipment doesn't  0, fulfil the i th criterion  th

di , j

where di,j is a fulfilment of the ith criterion (1 ≤ i ≤ m) on the jth work equipment (1 ≤ j ≤ m).

On the basis of the data on the weights of criteria of a piece of work equipment ui (weights of criteria are normalised so that their sum is equal to 1) u1, …, ui, …, un , where ui is weight of the ith criterion of a piece of work equipment (1 ≤ i ≤ n) and the values of criteria of the jth work equipment (di,j), the sum of the products of weights and values of criteria of the jth work equipment is calculated as: u1∙d1, j + … + ui∙di,j + … + un∙dn,j .

This is adaptability or index of adaptability of the jth work equipment.

n

k j = ∑ ui ⋅ di , j , i =1

where kj is adaptability of the jth work equipment, ui the weight of the ith criterion and di,j the value of the ith criterion of the jth work equipment.

It is convenient to write the calculation of adaptability of all offered work equipment units regarding their adaptability in a matrix form.

Three matrices must be created for this purpose: • matrix of weights U, which has one line and n columns: U = [u1, …, ui, …, un], •

•

... d1,m    ... di ,m  ,    ... d n ,m 

matrix K of adaptability of the offered work equipment, which has one line and m columns:

K = [k1, …, kj, …, km].

d1,j, …, di,j, …, dn,j, where:

 d1,1 ... d1, j     D =  di ,1 ... di , j     d  n ,1 ... d n , j

matrix D of values of functions, which has n lines and m columns:

Using the rule for multiplying matrices, we obtain the matrix of adaptability of the work equipment offered: K = U ∙ D. By ranking elements of matrix K, it is possible to obtain a smaller group of work equipment units for the fine selection.

3.2 Fine Selection of Work Equipment Fine selection of work equipment starts with tests at suppliers of the smaller group of work equipment, and by taking notes on the values of fulfilment of the required criteria, di,j*.

An n-tuple of numbers, di,j*, regarding the fulfilment of the required functions is created for the jth work equipment: d1,j*, …, di,j*, …, dn,j*.

The fulfilment of the criteria is between: 0 points ≤ di,j* ≤ 100 points, where di,j* is the value of fulfilment of the ith criterion of the jth work equipment (points).

On the basis of the data on weights of the required criteria of work equipment ui, and fulfilment of the criteria of the jth work equipment of the smaller group di,j*, it is possible to create a sum of products of weight factors and criteria fulfilment values of the jth work equipment: u1∙d1,j* + … + ui∙di,j* + … + un∙dn,j*.

This value is the adaptability of the jth work equipment:

n

k *j = ∑ ui ⋅ di*, j , i =1

where kj* is adaptability of the jth work equipment in a smaller group.

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In the fine selection, it is convenient to write the calculation of the adaptability of the smaller group of work equipment in a matrix form. For this purpose, it is necessary to create two matrices: • matrix D* of the values of fulfilment of criteria of the smaller group of work equipment, which has n lines and m columns:  d1*,1    D* =  di,*1     *  d n,1

•

d1*, j  di,* j  d n,* j

d1*,m     *  di,m ,    *  d n,m 

matrix K* of adaptability kj* of work equipment in the smaller group, which has one line and three columns:

K* = [k1*, k2*, k3*].

By multiplying the matrices, we obtain the matrix of adaptability of the smaller group of work equipment:

K* = U ∙ D*.

By ranking elements of adaptability matrix K*, it is finally possible to obtain the work equipment with the maximum adaptability, i.e. with the maximum adaptability index:

{ }

* kopt = max k *j . j =1, 2 ,3

4 CASE STUDY OF SELECTING THE OPTIMAL CNC-LATHE WITH RESPECT TO ADAPTABILITY By conducting an analysis of the production programme over the previous five years, a company’s management found that the production programme (considering types of products and their quantities) changed considerably from year to year.

The management decided that before buying work equipment in the future, it would be necessary to pay attention to the adaptability of it, in order to make the company more adaptable to the constant changes of the production programme. Using the defined criteria, it would be necessary to select work equipment that ensures the maximum adaptability.

After confirmation of the production programme for that particular year, it was found that the company urgently needed a new CNC lathe. A project team was established and tasked with selecting the CNC lathe that would ensure the maximum adaptability.

A creativity workshop was organised in the company [10], [11] and [16] with the aim of the team developing a target tree of criteria for assessing the adaptability of the CNC lathe. The results of development of the target tree of criteria are presented in Fig. 4.

Table 1. Questionnaire on fulfilment of criteria SUPPLIER of CNC lathe

Basic criterion no.

Criterion fulfilled

No Self-diagnosis – control 1 Self-acquisition and processing of data 2 Intelligence Decentralised control 3 Self-control – control of collision 4 Universality Functional flexibility – size 5 Flexibility Functional flexibility – function 6 Flexibility of layout 7 Size 8 Mass 9 Mobility Design of work equipment Robustness – stiffness for transport 10 Simple assembly/disassembly 11 /\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\ Operating conditions – vibrations 46 Operating conditions – temperature 47 Effect of environment on Compatibility CNC lathe Operating conditions – humidity 48 Self-sufficient supply – energy 49 Economy Investment costs 50

404

Berlec, T. – Kušar, J. – Rihar, L. – Starbek, M.

Yes

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 400-408

LEVEL:

1. target criterion

2. subcriterion

3. criterion

Intelligence UNIVERSALITY Flexibility

CNC lathe design

Installations

MOBILITY

Transport concept

ADAPTABILITY OF CNC LATHE

Reconfiguration (re-layout) MODULARITY

Machine modules Device modules

Standardisation

Flexibility

COMPATIBILITY

Neutrality of operation Environmental effects on the CNC lathe

ECONOMY

4. basic criteria 1 Self diagnosis - control 2 Self-acquisition and processing of data 3 Decentralised control 4 Self-control – control of collision 5 Functional flexibility – size 6 Functional flexibility – function 7 Flexibility of layout 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27

Size Mass Robustness – stiffness for transport Simple assembly / disassembly Used components Air connectors Sockets – direct Fuses Foundations Connection time Environmental effect – vibrations Environmental effect – temperature Environmental effect – moisture Environmental effect – sun Carts Crane Forklift Container Extraordinary transport Transport vehicle – height

28 29 30 31 32 33

Modular system of peripherals Fast construction / decomposition – delivery time Quick start of operation

Use in the manufacture of machine Reconfiguration with device modules

34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49

Interfaces Components of CNC lathe Capabilities of CNC lathes Maintenance of CNC lathe IT and communication technology Control – interchageability Control – postprocessors System of tools System of fixing CNC lathe environmental emissions (target value = 0) Law, regulations, guidelines CE Operating conditions – vibrations Operating conditions – temperature Operating conditions – humidity Self-sufficient supply – energy

Reconfiguration with functional and technology modules

50 Investment costs

Fig. 4. The criteria for assessing the adaptability of a CNC lathe

The project team, in collaboration with suppliers of CNC lathes, created a fulfilment scheme for each criterion defined on the fourth level of the target tree (Fig. 4). Because the CNC lathe adaptability criteria had various levels of importance, the project team assigned different weights to them by using the paircomparison method.

Known criteria for the CNC lathe adaptability and known weights of criteria allowed the project team to carry out the coarse and fine selection of the CNC lathe.

In order to carry out the coarse selection of CNC lathes, the project team sent a questionnaire regarding the fulfilment of criteria for evaluating the adaptability

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Table 2. Adaptability matrix of nine CNC lathes CNC-1 CNC-2 CNC-3 CNC-4 CNC-5 CNC-6 CNC-7 CNC-8 CNC-9 Weights of Basic required criterion criteria no. ui di,1 ui∙di,1 di,2 ui∙di,2 di,3 ui∙di,3 di,4 ui∙di,4 di,5 ui∙di,5 di,6 ui∙di,6 di,7 ui∙di,7 di,8 ui∙di,8 di,9 ui∙di,9 1 0.01 0 0.00 1 0.01 0 0.00 1 0.01 1 0.01 0 0.00 1 0.01 0 0.00 0 0.00 2 0.02 0 0.00 1 0.02 0 0.00 1 0.02 1 0.02 0 0.00 1 0.02 0 0.00 0 0.00 3 0 0 0.00 0 0.00 0 0.00 0 0.00 1 0.00 0 0.00 1 0.00 1 0.00 0 0.00 4 0.05 0 0.00 0 0.00 0 0.00 0 0.00 1 0.05 0 0.00 1 0.05 0 0.00 0 0.00 5 0.03 0 0.00 0 0.00 0 0.00 0 0.00 1 0.03 0 0.00 1 0.03 0 0.00 0 0.00 6 0.02 0 0.00 0 0.00 0 0.00 0 0.00 1 0.02 0 0.00 1 0.02 0 0.00 0 0.00 7 0.02 1 0.02 0 0.00 0 0.00 1 0.02 0 0.00 0 0.00 0 0.00 1 0.02 0 0.00 8 0.015 1 0.02 1 0.02 1 0.02 0 0.00 1 0.02 1 0.02 1 0.02 1 0.02 1 0.02 9 0.02 1 0.02 1 0.02 0 0.00 0 0.00 1 0.02 1 0.02 1 0.01 1 0.02 1 0.02 10 0.025 0 0.00 1 0.03 0 0.00 1 0.03 1 0.03 1 0.03 1 0.03 1 0.03 1 0.03 11 0 0 0.00 1 0.00 0 0.00 1 0.00 1 0.00 0 0.00 1 0.00 0 0.00 1 0.00 /\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\ 46 0.005 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 47 0.005 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 1 0.01 0 0.00 0 0.00 0 0.00 48 0 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 1 0.00 0 0.00 0 0.00 0 0.00 49 0 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 0 0.00 50 0.2 1 0.20 0 0.00 1 0.20 1 0.20 1 0.20 0 0.00 1 0.20 0 0.00 1 0.20 50

k j = ∑ ui ⋅ di , j ,

0.40

0.47

0.27

0.41

0.80

0.20

0.86

0.29

0.35

i =1

Rank of the offered CNC lathes:

5.

3.

8.

4.

2.

9.

1.

7.

6.

Table 3. Adaptability matrix of three CNC lathes Smaller group of CNC lathes j = 2, 5, 7 CNC-5 CNC-7 di,2* ui∙di,2* di,5* ui∙di,5* di,7* ui∙di,7* 1 Self-diagnosis – control 0.01 0.50 0.01 0.90 0.01 1.00 0.01 2 Self-acquisition and processing of data 0.02 0.50 0.01 0.90 0.02 1.00 0.02 3 Decentralised control 0 0.20 0.00 0.80 0.00 0.80 0.00 4 Self-control – control of collision 0.05 0.00 0.00 0.50 0.03 1.00 0.05 5 Functional flexibility – size 0.03 0.20 0.01 0.40 0.01 1.00 0.03 6 Functional flexibility – function 0.02 0.10 0.00 0.40 0.01 1.00 0.02 7 Flexibility of layout 0.02 0.00 0.00 0.00 0.00 0.00 0.00 8 Size 0.015 0.80 0.01 0.90 0.01 1.00 0.02 9 Mass 0.02 0.60 0.01 0.70 0.01 0.50 0.01 10 Robustness – stiffness for transport 0.025 1.00 0.03 0.90 0.02 0.90 0.02 11 Simple assembly/disassembly 0 1.00 0.00 0.80 0.00 0.60 0.00 /\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\/\ 46 Operating conditions – vibrations 0.005 0.10 0.00 0.20 0.00 0.30 0.00 47 Operating conditions – temperature 0.005 0.20 0.00 0.20 0.00 0.30 0.00 48 Operating conditions – humidity 0 0.20 0.00 0.20 0.00 0.20 0.00 49 Self-sufficient supply – energy 0 0.00 0.00 0.00 0.00 0.00 0.00 50 Investment costs 0.2 0.15 0.03 0.40 0.08 0.90 0.18 Basic criterion no.

Description of required criteria of a CNC lathe:

Weights of required criteria ui

CNC-2

50

k *j = ∑ ui ⋅ di*, j ,

0.45

0.58

0.75

i =1

RANK OF THE NARROWER GROUP of CNC lathes:

406

3.

Berlec, T. – Kušar, J. – Rihar, L. – Starbek, M.

2.

1.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 400-408

of CNC lathes to suppliers of CNC lathes. Table 1 shows part of the questionnaire.

After receiving answers from nine suppliers of CNC lathes, the project team made a coarse selection of the three CNC lathes that best fulfilled the required criteria. Using the known matrix D of fulfilment of the required criteria of the offered CNC lathes, and matrix U of criteria weights, the elements of adaptability matrix were calculated:

K = U ∙ D.

Some results of the adaptability matrix are shown in Table 2.

After selecting a smaller group of CNC lathe suppliers, the project team, in collaboration with the suppliers, made a plan of tests regarding the fulfilment of the criteria to be performed at a particular supplier.

During the tests, the project team members made notes on fulfilment of the required criteria on a particular CNC lathe. The lathe that best fulfilled a particular criterion received 100 points, while the other two lathes received penalty points.

Using the known matrix D* of achieved points, and matrix U of weights of criteria, the elements of the adaptability matrix were calculated: K* = U ∙ D*.

Some of the results of the adaptability matrix of the three CNC lathes from the smaller group are shown in Table 3.

By ranking the smaller group of three CNC lathes according to their adaptability (i.e. their adaptability index), the project team concluded that the company should buy the CNC-4 lathe that had the highest adaptability index.

5 CONCLUSIONS This article defines the components of adaptability of work equipment. The system of adaptability assessment is also presented.

The limitation of the suggested method of selecting the most adaptable work equipment is the input data of working equipment available on the market.

The target tree is the basis of the system for assessing the adaptability of work equipment, showing the hierarchy of criteria for assessment of work equipment adaptability. A scheme regarding the fulfilment of each basic criterion of the target tree is designed. Criteria have different importance and thus different weights.

Known values of meeting the criteria and corresponding weights of criteria allow the calculation of an adaptability index of the observed work equipment. By first making a coarse and then a fine selection of the available work equipment using the adaptability method, it is possible to determine the work equipment with the maximum adaptability index.

Future work will be focused on better optimisation with the use of more accurate multiobjective optimization [17].

6 REFERENCES [1] Hernandez, R. (2003).

Systematik der Wandlungsfähigkeit in der Fabrikplanung. VDI Verlag, Düsseldorf. (in German) [2] Reinhart, G., Berlak, J., Effert, C., Selke, C. (2002).

Wandlungsfähige Fabrikgestaltung. Zeitschrift für wirtschaftlichen Fabrikbetrieb (ZWF), vol. 97, no. 1/2, p. 18-23. (in German) [3] Zäh, M.F., Müller, N., Prasch, M., Sudhoff, W. (2004).

Methodik zur Erhöhung der Wandlungsfähigkeit von Produktionssystemen. Zeitschrift für wirtschaftlichen Fabrikbetrieb (ZWF), vol. 99, no. 4, p. 173-177. (in German) [4] Nyhuis, P., Heinen, T., Reinhart, G., Rimpau, C. (2008).

Wandlungsfähige Produktionssysteme.

Werkstattstechnik, vol. 98, no. 1/2, p. 85-91. (in German) [5] Wiendahl, H.P., Reichardt, J., Nyhuis, P. (2009).

Handbuch Fabrikplanung: Konzept, Gestaltung und Umsetzung wandlungsfähiger Produktionsstätten, Hanser Verlag, München. (in German), DOI:10.3139/9783446423237.

[6] Palčič, I., Buchmeister, B., Polajnar, A. (2010). Analysis of innovation concepts in Slovenian manufacturing companies. Strojniški vestnik - Journal of Mechanical Engineering, vol. 56, no. 12, p. 803-810.

[7] Anišić, Z., Krsmanović, C. (2008). Assembly initiated production as a prerequisite for mass customization and effective manufacturing, Strojniški vestnik - Journal of Mechanical Engineering, vol. 54, no. 9, p. 607-618.

[8] Božičković, R., Radošević, M., Ćosić, I., Soković, M., Rikalović. A. (2012). Integration of Simulation and Lean Tools in Effective Production Systems – Case Study, Strojniški vestnik - Journal of Mechanical Engineering, vol. 58, no. 11, p. 642-652, DOI: DOI:10.5545/sv-jme.2012.387.

[9] Kušar, J., Berlec, T., Žefran, F., Starbek, M. (2010).

Reduction of machine setup time. Strojniški vestnik Journal of Mechanical Engineering, vol. 56, no. 12, p.

833-845.

[10] Michalski, W.J. (2003). Six Sigma Tool Navigator: The Master Guide for Teams. Productivity Press, New York [11] Rihar, L., Kušar, J., Gorenc, S., Starbek, M. (2012).

Teamwork in the simultaneous product realisation.

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Strojniški vestnik – Journal of Mechanical Engineering, vol. 58, no. 9, p. 534-544, DOI:10.5545/ sv-jme.2012.420.

[12] Rihar, L., Kušar, J., Duhovnik, J., Starbek, M. (2010).

Teamwork as a precondition for simultaneous product realization. Concurrent Engineering: Research and Applications, vol. 18, no. 4, p. 261-273.

[13] Reddy, B.S.P., Rao, C.S.P. (2011). Flexible Manufacturing System Modelling and Performance Evaluation Using AutoMod. International Journal of Simulation Modelling, vol. 10, no. 2, p.78-90, DOI:10.2507/IJSIMM10(2)3.176.

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[14] Berkum, E.E.M. van (1987). Optimal Paired Comparison Designs for Factorial Experiments.

Centrum voor wiskunde en informatica, Amsterdam.

[15] Boardman, A.E. (2006). Cost-Benefit Analysis and Practice. Prentice Hall, Upper Saddle River.

[16] Scherer, J. (2007). Kreativitätstechniken. Gabal Verlag, Offenbach. (in German) [17] Zitzler, E. (1999). Evolutionary Algorithms for Multi Objective Optimization: Methods And Applications.

PhD Thesis, Swiss Federal Institute of Technology, Zürich.

Berlec, T. – Kušar, J. – Rihar, L. – Starbek, M.

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 409-417 © 2013 Journal of Mechanical Engineering. All rights reserved.

DOI:10.5545/sv-jme.2012.830

Received for review: 2012-10-11 Received revised form: 2013-02-11 Accepted for publication: 2013-03-14

Original Scientific Paper

Analysing Energy and Material Saving Technologies’ Adoption and Adopters Palčič, I. – Pons, M. – Bikfalvi, A. – Llach, J. – Buchmeister, B. Iztok Palčič1,* – Marc Pons2 – Andrea Bikfalvi2 – Josep Llach2 – Borut Buchmeister1 1 University

of Maribor, Faculty of Mechanical Engineering, Slovenia 2 Universitat de Girona, Spain

The main objective of this paper is to map the adoption of technologies for energy reduction and resources consumption in production.

The aim is also to contribute to the identification and understanding of the characteristics of the manufacturing firms that use these kinds of energy and material saving technologies. Our research is based on data from the largest European manufacturing survey to date and it includes data from Spain and Slovenia. The results show that the use of specific energy saving technologies and material saving technologies in manufacturing firms is still modest. Dividing manufacturing firms based on technology intensity sectors and based on their relative energy efficiency we have concluded that firms in high technology industries focus less on energy efficiency than low technology firms. Some other specific relationships between the use of energy efficient technologies and adopters’ characteristics (e. g. use of environmental management systems) are presented in this paper.

Keywords: energy efficiency, manufacturing firm, energy saving technology, material saving technology, European manufacturing survey

0 INTRODUCTION Manufacturing is defined as the transformation of materials and information into goods for the satisfaction of human needs. Turning raw materials into consumer products is also a major source of environmental pollution. Waste coming from manufacturing activities is an environmental threat originating in several regions around the world [1].

Therefore, in recent years, mostly in response to increasing pressure from environmental regulations, many manufacturing firms have made significant efforts to use cleaner production methods [2] to [4].

Industrial energy efficiency plays a central role as the manufacturing industry accounts for about 75% of the world’s yearly coal consumption, 44% of the world’s natural gas consumption, and 20% of global oil consumption. In addition, these manufacturing firms also use 42% of all the electricity generated [5]. Although renewable energy technologies, such as photovoltaic technology, might be a long-term solution, more efficient energy use can make the greatest and most economic contribution towards solving these problems in the short run. Using the available energy more efficiently is an effective countermeasure to rising energy needs and insecure energy supplies [6] and[7]. Bunse et al. [8] argue that examples in the literature and in real world practice show that although the manufacturing sector has made continuous improvements in energy efficiency, the economically beneficial energy efficiency potential has not yet been fully exploited [8] to [10].

This paper is based on an empirical study in the field of energy and material efficiency technologies.

The objective of this paper is firstly to map the adoption of technologies for the reduction of energy and resource consumption in production and, second, to contribute to the identification and understanding of the characteristics of the manufacturing firms that use these kinds of innovative technologies.

The paper is organized as follows. The introduction comprises a background and literature review of energy efficiency in production. Next, the research methodology and methods used to analyse the characteristics of energy and material saving technologies’ adoption and their adopters are presented. The results and findings are presented for the manufacturing firms with the use of descriptive statistics and simple correlation tests. Finally, we discuss our results and present some implications.

1 LITERATURE REVIEW The energy efficiency of manufacturing processes is becoming increasingly important due to rising energy costs and climate altering greenhouse gas emissions [11]. Improving energy efficiency is regarded as one of the most important options for reducing greenhouse gas emissions and for reducing the dependency of countries on energy imports [12]. Measuring energy efficiency is the basis for controlling energy consumption in the production processes and for deciding on improvement measures, as well as for tracking changes and improvements in energy efficiency [8]. Studies on the energy consumption of manufacturing processes have provided fundamental information for improving energy efficiency and building a comprehensive foundation aimed at

*Corr. Author’s Address: University of Maribor, Faculty of Mechanical Engineering, Smetanova 17, SI-2000 Maribor, Slovenia, iztok.palcic@uni-mb.si

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reducing the energy consumption of manufacturing processes [11]. There is also an on-going debate regarding the reasons why profitable investments to reduce energy consumption have not been carried out in companies [13] and [14]. There are several barriers to implementing energy efficiency improvement measures in firms, e. g. payback periods, limited capital, a low priority given to energy efficiency by the management, lack of information, or “difficult-tomeasure components” of energy investments [8] and [15] to [17].

Bunse et al. [8] argue that many industrial firms still lack appropriate methods to effectively address energy efficiency in production management.

Current approaches to integrating energy efficiency performance as a relevant criterion in production management seem to have shortcomings in their comprehensiveness and practicality. The authors of this paper argue that there are two reasons for this: the first is that there is no consensus on the definition of energy efficiency. The second reason is the variety of ways of measuring and monitoring energy efficiency.

When discussing energy efficiency in the industrial sector, different definitions are used [8] and [18] to [21]. Bunse et al. [8] define energy efficiency as “the ratio of energy services out to energy input (meaning) getting the most out of every energy unit you buy”. Increased energy efficiency may be accomplished by more efficient technology, energy recovery in the same process or further use of energy waste in different processes, increased energy conversion efficiency or optimized operational practices.

Energy efficiency developments can be monitored by quantifying the ratio of energy input and the useful output of a certain activity over time. The useful output of an activity can be defined in either physical (e.g. litres of beer produced or person kilometres driven) or monetary units (e.g. GDP of a country or value added of a sector) [12].

Energy efficiency indicators are usually ratios describing the relationship between an activity and the required energy. In the industrial sector, activities such as the production process of a product can be described in either economic or physical terms resulting in either economic or physical indicators.

Economic indicators are useful at an aggregated level, such as for comparing different sectors; however, to gain insight into particular manufacturing processes, physical indicators are more suitable [22]. Examples of physical indicators are specific energy consumption [22] to [25], final energy efficiency improvement [26], thermodynamic energy efficiency [19], etc. There is no 410

single energy efficiency indicator that can be applied in every situation, but the appropriate indicators have to be defined depending on the decision to be made or the decision tool to be applied [25].

Only several studies used for environmental variables the use of production activities or energy efficient technologies. One of the most recent is from Zeng et al. [27] who found an overall positive impact of cleaner production on firms’ business performance, but not under all circumstances. They argue that the cleaner production activities from low-cost schemes (e. g. improve employee environmental consciousness through training, improve working conditions to reduce waste, strictly enforce rules on cleaner production, increase the recyclability of the products and components) make a bigger contribution to financial performance than high-cost scheme activities (e. g. using energy efficient and clean technologies or using renewable resources as raw materials), which require significant financial investment but may not result in immediate economic benefit. Thus, lowcost schemes for cleaner production activities do not require significant financial input but may bring immediate financial benefits.

2 METHODOLOGY We used data from the European Manufacturing Survey (EMS) for our research. The EMS is the largest European survey on manufacturing activities and is coordinated by the Fraunhofer Institute for Systems and Innovation Research (ISI), Germany.

The survey collects data on manufacturing strategies, the application of innovative organisational and technological concepts in production, personnel deployment and qualification, the production offshoring and back-sourcing activities, cooperation patterns, etc. Data on firm characteristics and performance indicators (R&D expenses, productivity, returns on sales,) is also collected.

The 2009 EMS edition was carried out in 12 countries. This paper uses data from the Spanish and Slovenian sub-samples. The Spanish sub-sample had 116 responses and the Slovenian accounted for 64, altogether 180 responses. The survey was performed in manufacturing firms (NACE codes from 15 to 37) with at least 20 employees.

In recent years, only a few surveys in the world have been launched that analyse energy efficiency in manufacturing firms and their energy saving technologies (EST) and material saving technologies (MST) use. These existing surveys cover only some industrial sectors - monitoring very specific

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technologies or cover only American and Asian countries. None of them include the European countries covered by EMS, which also encompasses all manufacturing industries. The latest survey added several questions related to environmental and energy issues; the EMS defines 10 general groups of technologies: 8 for energy efficiency and 2 for material consumption saving. Any specific technology can be classified into one of these broad groups, thereby creating a global map of their use and level of implementation.

The EST included were: T1. control system for shut down of machines in offpeak periods, T2. electric motors with rotation speed regulation, T3. compressed air contracting, T4. highly efficient pumps, T5. low-temperature joining processes, T6. retrieval of kinetic and process energy, T7. combined cold, heat and power – Bi-/Trigeneration and T8. waste material for in-house energy generation.

We included two MST: T9. utilisation of recycled material in product manufacturing and T10. product recovery after product life cycle.

EST and MST are characterized in terms of use and also in terms of usage levels (extent of use) through a descriptive and a frequency analysis. The extent of actual use is referred to by comparing the actual use of the technology in the firm to the most reasonable potential use. There are three levels: the extent of utilised potential is “low” for an initial attempt to utilise, “medium” for partly utilized, and “high” for extensive use.

We have analysed the characteristics of EST and MST adopters according to the OECD’s taxonomy of manufacturing industries classified by their technological intensity [28]. We have formed three groups: “Low technology” with firms from NACE 1516, 17-19, 20-22, 36-37; “Medium-Low technology” with firms from NACE 23, 25, 26, 351, 27, 28; and “Medium-High and High technology” with firms from NACE 24, 31, 34 excl. 2423, 352+359, 29 and 353, 2423, 30, 32, 33. As shown in Table 1, the majority of firms fall within the Medium-Low technology group.

If we join Medium-Low technology and MediumHigh technology, this group consists of 131 firms, thus making Medium technology industry the largest group. Since our High technology industry group (NACE 353, 2423, 30, 32 and 33) involves only 13 firms, this group was merged with the Medium-High technology industry group in order to reduce the

number of groups. We created a discrete variable to group this classification into three categories: “Low technology” – value 1, “Medium-Low technology” – value 2, and “Medium-High and High technology” – value 3. “Medium-Low technology” was taken as a reference variable.

Next, we classified technology adopters into three groups that represent the relative energy and materials consumption efficiency in production. These groups were created from the responses to the question regarding the perception of their production efficiency in terms of actual material and energy consumption in comparison with other factories in their industry.

Energy efficiency is therefore measured on a relative scale with values from 1 to 5. The scale ranges from 1 meaning much less efficient (0.5%) to a value of 5 meaning much more efficient (2.2%). The value 2 indicates somewhat less efficient (6.1%), 3 indicates equally efficient (62.6%), and 4 indicates somewhat more efficient (28.5%). In the analyses, three groups have been created from this variable: “Less efficient” (firms rated with values 1 or 2), “Equally efficient” (firms rated with a value of 3) and “More efficient” (firms rated with values 4 or 5).

3 RESULTS AND FINDINGS Table 1 presents the results according to the OECD’s taxonomy of manufacturing industries classified by their technological intensity. The results show that adopters in higher technologically intensive industries have, on average, a higher number of employees, higher percentage of firms with R&D expenditure, superior exportation intensity (more than 50% of sales abroad), and a strong use of environmental management systems, such as ISO 14000 [29]. Firms in Medium-high and High Technology industrial sectors also had an average turnover in 2008 of more than the double of each one of the other two technological groups (62 vs. 22 and 30 M€). From the average estimation of material and energy efficiency in production (max. range=5, min. range=1), there is hardly any difference between these groups of technological sectors relative to the average of material and energy efficiency in production.

Table 2 presents the results according to three groups representing the relative energy and materials consumption efficiency in production. There are only 12 firms in the “Less efficient” group. The majority of firms are in the “Equally efficient” group. There were only 4 firms that claimed that they were considerably more efficient than other firms in their industry.

The “More efficient” group had altogether 55 firms.

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Table 1. Summary of descriptive features of the sample by technological intensity Low technology N Number of employees 2008. N=187 Turnover 2008. N=169 Firms with R&D expenditures. N=183 High exportation intensity firms. N=172 Firms with ISO 14000 implemented. N=187 Average energy efficiency in production. N=179 T1: Control system for machine shut down. N=179 T2: Speed regulation. N=184 T3: Compressed air contracting. N=182 T4: Highly efficient pumps. N=181 T5: Low-temperature joining processes. N=182 T6: Energy retrieval. N=178 T7: Bi-/Tri-generation. N=178 T8: Waste material for energy. N=178 T9: Recycled material in production. N=178 T10: Product recovery. N=178

43 (23%) 76 30 M€ 50% 24% 14% 3.25 29% 57% 26% 26% 0% 10% 5% 10% 45% 35%

Medium-Low technology 82 (44%) 137 22 M€ 53% 39% 30% 3.32 24% 51% 16% 16% 9% 7% 6% 9% 30% 26%

Medium-High and High technology 62 (49+13) (33%) 402 62 M€ 75% 53% 34% 3.18 30% 62% 31% 24% 15% 9% 12% 10% 34% 28%

187 211 37 M€ 60% 41% 28% 3.26 27% 56% 24% 21% 9% 8% 8% 10% 35% 29%

More efficient 55 (51+4) (30%) 377 64 M€ 60% 37% 45% 2.07 42% 75% 25% 28% 11% 15% 9% 15% 39% 37%

Total 179 216 38 M€ 61% 42% 28% 2.12 28% 57% 24% 22% 9% 9% 8% 10% 35% 30%

Total

Table 2. Summary of descriptive features of the sample by relative efficiency in production N Number of employees 2008. N=179 Turnover 2008. N=164 Firms with R&D expenditures. N=176 High exportation intensity firms. N=165 Firms with ISO 14000 implemented. N=179 Average of Technological intensity. N=179 T1: Control system for machine shut down. N=171 T2: Speed regulation. N=176 T3: Compressed air contracting. N=174 T4: Highly efficient pumps. N=173 T5: Low-temperature joining processes. N=174 T6: Energy retrieval. N=170 T7: Bi-/Tri-generation. N=170 T8: Waste material for energy. N=170 T9: Recycled material in production. N=170 T10: Product recovery. N=170

Less efficient 12 (1+11) (7%) 140 13 M€ 58% 64% 33% 2.23 33% 42% 17% 8% 8% 0% 8% 0% 42% 0%

Regarding energy and materials efficiency, descriptive analysis shows that firms belonging to the more relatively efficient groups have, on average, higher numbers of employees (much more than equally or less efficient firms: 377 vs. 146 and 140). Average firm turnover also increases as the relative efficiency of these firms increases (the “More efficient” group has an average turnover of 64 vs. 23 and 13 M€).

However, high exportation intensity (more than 50% of sales abroad) is reduced on average as the studied groups gain relative efficiency. Looking at the 412

Equally efficient 112 (63%) 146 23 M€ 61% 42% 17% 2.13 21% 50% 25% 20% 8% 7% 7% 9% 32% 29%

technological intensity of the firms we can see a slight decrease in the technological intensity values from the “Less efficient” group to the “More efficient” group.

And finally, looking at R&D expenditure and the use of environmental management systems, such as ISO 14000, hardly any trend is visible. Interestingly, the average use of environmental management systems in the “Equally efficient” group is much smaller than in the “Less efficient” group.

Fig. 1 depicts the use of EST and MST for all manufacturing sectors presented. Here we see that

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“Speed control” is the most highly used technology with 56% of affirmative responses. The second and third place technologies in the use ranking MST were “Recycled material in production” with a 35% and ”Product recovery” with a 29%. The second EST “Control system for shut down of machines in off-peak periods” (27%) is in the fourth position and has much lower use than the top ranking “speed control.” However, the highest ranking “Speed control” technology and its large lead over the other technologies could be misleading since this technology may be misunderstood or very broadly interpreted. The term “Electric motors with rotation speed regulation” can be understood to mean that almost any machine that produces any kind of motion or rotation with a common speed regulation system over the engine has implemented this technology.

For most machines, however, this is not an option, but rather an intrinsic characteristic. Therefore, it is questionable whether “Speed control” should be considered an EST.

This fact is more evident for EST and less so for MST as both MST technologies are more widely used and both have relatively high extensive share of use. Only “Energy retrieval” technology and “Control system for shut down of machines in off-peak periods” technology have the smallest usage share in the “high use” group.

Fig. 2. Degree of implementation of EST and MST for all manufacturing sectors

Fig. 1. Use of EST and MST for all manufacturing sectors

Fig. 3. Implementation percentage of EST and MST by technological sector

The graph in Fig. 2 shows a distribution of technologies used according their implementation degree and ranked from the highest implementation level to the lowest. This ranking compared to the simple use has changed. “Bi-/Tri-generation” is the EST with the greatest high implementation rate (43%), together with “Product recovery,” which is an MST. The second EST in the ranking of highly implemented technologies is “Highly efficient pumps” with 42%. The “Speed regulation” technology was the most widely used technology, but only 31% of firms acknowledged strong use of this technology, giving it a ranking of 7. This fact could be again related to the possible misunderstanding of the term “Speed regulation”.

Nevertheless, a percentages variance of the highly implemented technologies is smaller compared with the percentages of the simple use of these technologies.

Fig. 3 presents EST and MST in accordance with the three technological intensity groups. The technologies are ranked based on the share of use in the “Medium-High and High technology” group (from the highest to the lowest share). “Lowtemperature joining processes” is the technology with the highest percentage of use in “Medium-High and High technology” with 56%. No firm utilises this technology in the “Low technology” group. “Bi-/ Tri-generation” is the only other technology that is predominantly used in the “Medium-High and High technology” group with 50%. We can see that none of these technologies is widely used in the “Low technology” group (firms in this group represent a percentage of the total used technologies, which is always below 30%). On the other hand at least 30% of the used technologies are within the “Medium-High and High technology” group (from 31 to 56%). MST

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have the lowest use among firms in the “MediumHigh and High technology” group.

Fig. 4. High implementation percentage of EST and MST by technological sector

We have also analysed the use of highly implemented technologies according to three technological intensity groups (Fig. 4). The highly implemented technologies are ranked based on the share of use in the “Medium-High and High technology” group (from the highest to the lowest share). This ranking, compared to the general implementation degree, has changed.

Interestingly, the average percentage of highly implemented technologies in the “Medium-High and High technology” group is lower than for the implementation of EST and MST in general. This leads to the conclusion that the analysed EST and MST are predominately highly implemented in low and medium-low technology groups.

with 53%. No firm (0%) in the “Less efficient” group uses either “Energy retrieval,” “Waste material for energy” or “Product recovery.”. “Energy retrieval” is the only technology most widely used in the “More efficient” group. All the other technologies are most widely used in the “Equally efficient” group (47 to 64%), which represents 65% of the total number of firms. It is very obvious that EST and MST are hardly used in the “Less efficient” group, with a share always lower than 9%. On the other hand at least 30% of the technologies are used within the “More efficient” group (from 31 to 53%).

We have also analysed the use of highly implemented technologies according to three groups that represent the relative energy and materials consumption efficiency in production (Fig. 6). The highly implemented technologies are ranked based on the share of use in the “More effective” group (from the highest to the lowest share). This ranking has changed compared to the general implementation degree, but not very drastically. More importantly, we can see that the analysed EST and MST are usually highly implemented in firms that claim to be more energy efficient than other firms in their industry.

Fig. 6. High implementation percentage of EST and MST by level of efficiency relative to the sector

Fig. 5. Implementation percentage of EST and MST by level of efficiency relative to the sector

Fig. 5 presents EST and MST implementation in three groups that represent the relative energy and materials consumption efficiency in production. The technologies are ranked based on the share of use in the “More effective” group (from the highest to the lowest share). “Energy retrieval” is the technology with the highest share of the “More efficient” group 414

In order to test the possible relationship between technology level (intensity) and environmental management systems implementation and the number of EST and MST implemented or the number of these technologies highly implemented, several tests of correlation have been carried out (Tables 3 and 4). We conducted correlation tests using the Pearson correlation value.

Table 3 shows a correlation matrix between the firm’s technology level and use and the high use of EST and MST. The results show that no significant correlation appears between technology level and EST and MST use and high use.

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Table 3. Correlation matrix between firm’s technology level and use and high use of EST and MST Technology level Technology level

Number of EST implemented Number of EST highly implemented Number of EST and MST implemented Number of EST and MST highly implemented

Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N

Number of EST implemented

1

0.076 0.300 187 1

187 0.076 0.300 187 0.076 0.300 187 0.042 0.565 187 -0.063 0.390 187

187 0.531* 0.000 187 0.927* 0.000 187 0.480* 0.000 187

Number of EST highly implemented -0.021 0.781 187 0.531* 0.000 187 1

Number of EST and MST implemented 0.042 0.565 187 0.927* 0.000 187 0.509* 0.000 187 1

187 0.509* 0.000 187 0.928* 0.000 187

187 0.519* 0.000 187

Number of EST and MST highly implemented -0.063 0.390 187 0.480* 0.000 187 0.928* 0.000 187 0.519* 0.000 187 1 187

* Correlation is significant at the 0.01 level (2-tailed)

Table 4. Correlation matrix between environmental management systems use and the use and high use of EST and MST Technology level Technology level

Number of EST implemented Number of EST highly implemented Number of EST and MST implemented Number of EST and MST highly implemented

Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N Pearson Correlation Sig. (2-tailed) N

Number of EST implemented

1 187 0.238* 0.001 187 0.056 0.445 187 0.292* 0.000 187 0.087 0.237 187

0.238* 0.001 187 1 187 0.531* 0.000 187 0.927* 0.000 187 0.480* 0.000 187

Number of EST highly implemented 0.056 0.445 187 0.531* 0.000 187 1

Number of EST and MST implemented 0.292* 0.000 187 0.927* 0.000 187 0.509* 0.000 187 1

187 0.509* 0.000 187 0.928* 0.000 187

187 0.519* 0.000 187

Number of EST and MST highly implemented 0.087 0.237 187 0.480* 0.000 187 0.928* 0.000 187 0.519* 0.000 187 1 187

* Correlation is significant at the 0.01 level (2-tailed)

We also wanted to explore the relationship between environmental management systems use and the use and high use of EST and MST. As shown in Table 4, only the simple use of EST and MST is significantly correlated with environmental management systems such as ISO 14000, but not with the high use of these technologies. In these cases, both Pearson correlation coefficients are significant at 0.01 level (2-tailed), and the one that also considers MST is higher than the one considering only EST. Consequently, the relationship strength is also slightly higher.

4 CONCLUSIONS Based on our analysis five conclusions can be drawn.

A general observation on the use of EST and MST is that the use of these technologies in manufacturing firms is still relatively low (from 8 to 35%). The only exception is “Speed control” technology with 56%. The first conclusion is that in analysing energy efficiency groups we have observed a slight decrease in the technological intensity values from the “Less efficient” group to the “More efficient” group. On the other hand, the Low Technology group has a slightly higher average of material and energy efficiency

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in production than the Medium-High and High technology group (3.25 vs. 3.18). Both these facts could reveal a possible negative relationship between energy efficiency in production and technological intensity of firms, at least on average. This suggests that firms in high technology industries focus less on energy efficiency than low technology firms.

Secondly, both MST are ranked second and third in general use. But it is interesting to note that they are mostly used in the low and medium technology sector, not in the high technology one. “Product recovery after product life cycle” is the most widely highly utilised technology, being mostly used in the low and medium technology sector.

Thirdly, only 7% of all manufacturing firms claims to be less energy efficient than firms from their sector, 30% believe they are more energy efficient than others. We calculate that MST and EST are on average used 41% in the more efficient group, 55% in the medium efficient group, and 4% in the less efficient group of firms. Based on this fact, manufacturing firms are more efficient if they use at least one EST or MST. Fourthly, in analysing EST and MST we focused on manufacturing firms that showed high implementation of these technologies. We have analysed these technologies according to their use in different technology intensity sectors and based on the energy efficiency of the firms. We found that the analysed EST and MST are predominately highly implemented in low and medium-low technology groups and less so in the “Medium-High and High technology” group. This observation could again prove that firms in high technology industries focus less on energy efficiency than low technology firms.

However, our results show that there is no significant correlation between technology level and the percentage of EST and MST use and high use.

On the other hand, the analysed EST and MST are usually highly implemented in firms that claim to be more energy efficient than other firms in their industry.

This leads to a potentially positive relationship between being energy and material efficient and using energy efficient technologies, especially if they are highly implemented.

Our final conclusion concerns the implementation of environmental management systems. Our results showed a positive significant relationship between energy and material efficiency, but only with use (not high use) of these technologies.

Our research has several limitations. The first is that only descriptive statistics and correlation tests were used to map the characteristics of energy 416

efficient technologies and their adopters. To draw further conclusions in the future several advanced statistical methods will have to be used (e. g. linear regression for quantitative independent variables and ordinal logistic regression). We will further explore the relationship between the implementation of energy efficient technologies and the environmental performance of manufacturing firms. In addition, we will also examine the use of these technologies and the economic performance of manufacturing firms.

Another limitation is also the narrow geographical coverage and the fact that no similar previous data exists with which to compare our findings. This shortcoming is already being addressed by the inclusion of energy efficiency questions in the new European Manufacturing Survey 2012.

Despite these shortcomings, our contribution categorizes the use of energy efficient technologies, describes the characteristics of their adopters, and indicates a possible influence of these technologies on the environmental performance of manufacturing firms.

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30, p. 663-679, DOI:10.1016/S0301-4215(02)00023-X. [23] Farla, J., Blok, K., Schipper, L. (1997). Energy efficiency developments in the pulp and paper industry: a cross-country comparison using physical production data. Energy Policy, vol. 25, p. 745-758, DOI:10.1016/ S0301-4215(97)00065-7.

[24] Duhovnik, J., Žargi, U., Kušar, J., Starbek, M. (2009). Project-driven concurrent product development.

Concurrent Engineering-Research and Applications, vol. 17, no. 3, p. 225-236, DOI:10.1177/1063293X09343823.

[25] International Energy Agency. (2007). Tracking Industrial, Energy Efficiency and CO2 Emissions.

Retrieved on 20. 8. 2012, from http://www.iea.org/ textbase/nppdf/free/2007/tracking_emissions.pdf.

[26] Irrek, W., Thomas, S. (2006). Der Energie Spar Fonds für Deutschland. Hans-Böckler-Stiftung, Düsseldorf. (in German) [27] Zeng, S.X., Meng, X.H., Yin, H.T., Tam, C.M., Sun, L. (2010). Impact of cleaner production on business performance. Journal of Cleaner Production, vol. 18, p. 975-983, DOI:10.1016/j.jclepro.2010.02.019.

[28] OECD (2005) Directorate for Science, Technology and Industry, stan indicators (2005 edition): 1980-2003. Organisation for Economic Cooperation and Development, Paris. Retrieved on 29. 9. 2012, from http://www.oecd.org/industry/ industryandglobalisation/40230754.pdf.

[29] ISO 14000. (2004). Environmental management The ISO 14000 family of International Standards.

International Organization for Standardization. Geneva.

Analysing Energy and Material Saving Technologies’ Adoption and Adopters

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DOI:10.5545/sv-jme.2012.648

Short Scientific Paper

Received for review: 2012-06-12 Received revised form: 2013-02-11 Accepted for publication: 2013-03-13

Modeling Nonlinear Viscoelastic Nanoindentation of PVAc at Different Unloading Rates Kucuk Y. Yilmaz Kucuk

Bartin University, Department of Mechanical Engineering, Turkey In this study, nanoindentation technique was used to determine the nonlinear viscoelastic behavior of polyvinyl acetate (PVAc) with a glass transition temperature of 29 °C. Nanoindentation was conducted with a Berkovich nanoindenter tip under the same constant loading rate but different unloading rates. The Burgers nonlinear viscoelastic model was implemented in ABAQUS/Standard finite element code by means of a user-defined subroutine. After fitting the complete load-displacement curves, parameters in the nonlinear Maxwell and Voigt elements were obtained to describe the nonlinear viscoelastic behavior of PVAc. The results show that PVAc exhibits negative unloading stiffness under low unloading rates. However, if the unloading rate is higher than 0.1 mN/s, PVAc appears more elastic behavior. The nonlinear Burgers model used can successfully model the nonlinear viscoelastic behavior of PVAc at different unloading rates, including the unloading “nose” associated with negative unloading slope at slow unloading rate for a highly viscoelastic material.

Keywords: nanolinear viscoelastic nanoindentation, PVAc, viscoelasticity, finite element simulation, FEM, nonlinear Burgers model

0 INTRODUCTION Polyvinyl acetate (PVAc) has a wide application area as general sensitive adhesives in addition to be used as a constituent in paint, textile, and paper coating products. Depending on the concentration of PVAc, the ethylene polymers find application in a range of areas including high clarity and flexible packaging materials, impact resistant footwear, and adhesives.

Nanoindentation is an effective technique to measure mechanical properties of materials and it can also characterize material behavior at micron/ submicron scale. The fundamentals of nanoindentation have been well-established [1] to measure some properties of elastic-plastic materials. However, PVAc often shows negative unloading slope, giving negative contact stiffness under nanoindentation and as a result, the elastic analysis is not appropriate to extract its viscoelastic properties. In some previous studies, the viscoelastic properties of PVAc have been examined under both linear [2] and [3] and nonlinear [4] states by the use of its time-dependent relaxation modulus and creep compliance. Huang and Lu [5], carried out some indentation tests on two polymeric materials (PMMA and PVAc) by using both spherical and Berkovich (three-faced) indenters. Using two different indenter shapes, they were able to evaluate two independent material relaxation functions.

They employed the method of Lee and Radok [6] to incorporate viscoelastic behavior by implementing hereditary integral operators into the elastic solution expressions that were presented by Sneddon [7]. In nanoindentation, the load and deformation relations (P-h curves) have been studied originally by Hertz [8] and developed by Sneddon et al., with additional 418

studies [7], [9] and [10]. More recently Kucuk, et al. investigated the nonlinear viscoelastic nanoindentation of polymethyl methacrylate (PMMA) by modeling the complete loading and unloading curves [11].

In the past several decades, a number of authors have studied constitutive law, which emphasizes the effect of strain-induced increase in the stress relaxation process, to determine the nonlinear viscoelastic response of the polymers.

Shay and Caruthers [12] demonstrated that yield could be calculated using their constitutive equation in numerical simulations of constant extension rate experiments. Knauss and Emri [13] and [14] developed a free volume based constitutive equation. Wineman and Waldron [15] considered the effect of strain on the acceleration of stress relaxation by applying a form of their constitutive equation. By means of the linear viscoelastic operators that are employed in time-dependent integrals, material properties can be determined based on linear mechanics. However, even in the case of elastic materials, it is known that a plastic deformation zone occurs in the specimen underneath the contact area between the nanoindenter tip and test specimen. In most analytical studies on nanoindentation, only the loading portion of the response has been examined to predict the viscoelastic behavior of test materials. Recently, Wang et al. [16], performed some nanoindentation tests on PVAc to examine the viscoelastic behavior especially in nonlinear viscoelastic region. However, there is a lack of consideration to evaluate the effect of unloading portion of the response on nonlinear viscoelastic behavior of PVAc. In this study, the nanoindentation tests including a constant loading rate and different unloading rates

*Corr. Author’s Address: Bartin University, Department of Mechanical Engineering, Turkey, yilmazkucuk75@gmail.com

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, 418-422

were performed to predict the nonlinear viscoelastic behavior of PVAc. The Burgers nonlinear viscoelastic model was implemented to ABAQUS/Standard code with a user-defined subroutine. After fitting the loaddisplacement curves to the simulation curves obtained from the Burgers model curves, the nonlinear behavior of PVAc could be predicted under different strain rates.

1 THEORY Recently, Chen et al. [17] prepared a review paper summarizing different three-dimensional viscoelastic models, including Hooke, Newton, Maxwell, Voigt, Boltzmann, Zener, Tsay, Burgers, Weichert, and Kelvin models. The creep and stress relaxation terms cannot be described by the Maxwell and the Voigt elements, respectively [18]. The nonlinear viscoelastic behavior under all situations cannot be described by using available nonlinear viscoelastic models in the literature [19]. In this study, a phenomenological model that is generally called four-parameter Burgers model [20] was selected. The theory applied to the Burgers model has been introduced in detail in a paper recently published by Kucuk et al. [11].

2 EXPERIMENTAL PVAc specimen having flat prismatic shape with thickness of 1 mm was cut and fixed properly on the sample tray of indentation device. Then, the nanoindentation test parameters that are necessary to input in first stage of process were recorded on the software provided by the nanoindentation device (Agilent Nano Indenter G200 system).

The nanoindentation tests were carried out with Berkovich indenter tip at the room temperature (23 °C). For the load of 8 mN with a constant loading rate of 0.05 mN/s and five unloading rates (0.05, 0.1, 0.2, 0.4 and 0.8 mN/s) were implemented, respectively.

To ensure the consistency of results obtained from loading-unloading steps, twelve nanoindentation tests in a 4×3 grid have been performed. The distance between the neighboring indents (in both X and Y directions) was selected as 100 μm for each indent.

All nanoindentation experiments could be started after reaching the drift rate to 0.05 nm/s or below this rate.

For simulations, the indenter was modeled as a rigid shell and full integration 3D stress elements were selected for the sample. Fig. 1 shows the mesh of 1/6 symmetric PVAc sample and the nanoindenter tip.

Fig. 1. Berkovich indenter using rigid shell element and 1/6 symmetric PVAc sample

3 RESULTS AND DISCUSSION In this study, the nonlinear viscoelastic behavior of PVAc at different unloading rates was investigated by nanoindentation technique. The viscoelastic materials can exhibit negative stiffness during unloading part of indentation process for slow unloading rates. The Oliver-Pharr method cannot give the modulus under such a condition due to negative unloading stiffness or high unloading stiffness induced by viscoelastic effects. Burgers model parameters that are summarized in Table 1 were determined by curve fitting method between the loading-unloading and Burgers model response curves. Therefore, these material parameters are then used in the nonlinear Burgers model to simulate the nanoindentation under other unloading rates. The constant loading rate of 0.05 mN/s was used for all unloading rates. It can clearly be seen from the simulation results that the model used can capture the negative stiffness during initial unloading (Fig. 2). According to the load-displacement curves obtained from experiments and simulations, the PVAc specimen used exhibits negative stiffness in both unloading rates of 0.05 and 0.1 mN/s (Figs 2a and b).

This behavior can be explained as creep under a low unloading rate.

At the beginning of the unloading portion, the load is almost at maximum and thus, the effect of creep is currently predominant because of the decrease slightly at maximum load. However, this viscoelastic behavior cannot be maintained for higher unloading rate due to the diminishing effect of viscosity (Figs. 2c to e). Another point that must be emphasized here is of the recovery capability of PVAc. After reaching to unloading rate of 0.1 mN/s, PVAc specimen exhibited the positive stiffness and an increase in its capability of recovery was also observed. Consequently, this results support this comment that the unloading

Modeling Nonlinear Viscoelastic Nanoindentation of PVAc at Different Unloading Rates

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Fig. 2. Comparison of load vs. displacement curves obtained from simulation and nanoindentation results for constant load of 8 mN at different unloading rates of a2) 0.05 mN/s b2) 0.1 mN/s, c2) 0.2 mN/s, d2) 0.4 mN/s, e2) 0.8 mN/s: left column contour plot of the displacements into the surface, right column comparison between the nanoindation experimental data and simulation results

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Table 1. Nonlinear Burgers model parameters used in simulations Elastic Parameters ν E [GPa] 3.5

0.35

ct mt tξ [s]

1 0.85 0.1 0.5

2 0.70 0.25 3

Transient Parameters (Voigt Units) 3 4 5 6 0.70 0.70 0.70 0.4 0.15 0.25 0.3 0.3 50 100 125 150

rate has strictly effect on the linear and/or nonlinear viscoelastic behaviors of PVAc specimen used.

4 CONCLUSIONS In this study, nonlinear viscoelastic behavior of PVAc has been modeled using a nonlinear Burgers model. A set of parameters in the nonlinear Burgers model were fitted to the nanoindentation load-displacement curve with a peak load of 8 mN under 0.05 mN/s loading/ unloading rate, and all subsequent simulations were carried out using these parameters. PVAc exhibits a negative stiffness during unloading with slow rates (0.05 and 0.1 mN/s). However, if the unloading rate is higher than 0.1 mN/s, PVAc exhibits more elastic behavior. A good agreement can be achieved between simulations and nanoindentation experiments under different unloading rates. The results indicate that the nonlinear Burgers model is capable of modeling the nonlinear behavior of PVAc under different unloading rates from 0.05 to 0.8 mN/s. The unloading rate has explicit effect on the viscos properties of PVAc specimen used. Also, an increase in unloading rate leads to an increase in capability of recovery for PVAc sample. By the use of the model parameters extracted, the further predictions regarding PVAc could be achieved by applying different loading & unloading conditions.

5 ACKNOWLEDGEMENT The author would like to thank Dr. Hongbing Lu at the University of Texas at Dallas for providing equipment used in this paper, and the support of DOE Nuclear Energy University Program (NEUP) grant 09-416, the ONR MURI BAA 10-026, and NSF CMMI-1031829 & CMMI-1132174.

6 REFERENCES [1] Oliver, W.C., Pharr, G.M. (1992). An Improved technique for determining hardness and elastic modulus using load and displacement sensing indentation

7 0.05 0.3 250

8 0.05 0.3 300

Steady Parameters cS mS 0.040 0.060

experiments. Journal of Materials Research, vol. 7, no.

6, p. 1564-1583, DOI:10.1557/JMR.1992.1564.

[2] Knauss, W.G., Kenner, V.H. (1980). On the hygrothermomechanical characterization of polyvinyl acetate. Journal of Applied Physics, vol. 51, no. 10, p.

5131-5136, DOI:10.1063/1.327458.

[3] Deng, T.H., Knauss, W.G. (1997). The temperature and frequency dependence of the bulk compliance of poly(vinyl acetate). Mechanics of TimeDependent Materials, vol. 1, no. 1, p. 33-49, DOI:10.1023/A:1009734225304.

[4] Arenz, R.J. (1999). Nonlinear shear behavior of poly(vinyl acetate) material. Mechanics of TimeDependent Materials, vol. 2, no. 4, p. 287-305, DOI:10.1023/A:1009827310712.

[5] Huang, G., Lu, H. (2007). Measurements of two independent viscoelastic functions by nanoindentation.

Experimental Mechanics, vol. 47, no. 1, p. 87-98, DOI:10.1007/s11340-006-8277-4.

[6] Lee, E.H., Radok, J.R.M. (1960). The contact problem for viscoelastic bodies. Journal of Applied Mechanics, vol. 27, no. 3, p. 438-444, DOI:10.1115/1.3644020.

[7] Sneddon, I.N. (1965). The relation between load and penetration in the axisymmetric boussinesq problem for a punch of arbitrary profile. International Journal of Engineering Science, vol. 3, no. 1, p. 47-57, DOI:10.1016/0020-7225(65)90019-4.

[8] Hertz, H. (1882). Uber die beruhrung fester elastischer korper. Journal fur die Reine und Angewandte Mathematik, vol. 1882, no. 92, p. 156-171, DOI:10.1515/crll.1882.92.156. (in German) [9] Harding, J.W., Sneddon, I.N. (1945). The elastic stresses produced by the indentation of the plane surface of a semi-infinite elastic solid by a rigid punch. Proceedings of the Cambridge Philosophical Society, vol. 41, no. 1, p. 16-26, DOI:10.1017/S0305004100022325.

[10] Sneddon, I.N. (1948). Boussinesq’s problem for a rigid cone. Proceedings of the Cambridge Philosophical Society, vol. 44, no. 4, p. 492-507, DOI:10.1017/ S0305004100024518.

[11] Kucuk, Y., Mollamahmutoglu, C., Wang, Y., Lu, H. (2013). Nonlinearly viscoelastic nanoindentation of PMMA under a spherical tip. Experimental Mechanics, vol. 53, no. 5, p. 731-742, DOI:10.1007/s11340-0129695-0.

[12] Shay, R.M., Caruthers, J.M. (1986). A new nonlinear viscoelastic constitutive equation for predicting yield

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Proceedings of the SEM Series, vol. 15, p. 93-100, DOI:10.1007/978-1-4419-9794-4_16.

[17] Chen, D-L., Yang, P-F., Lai Y-S. (2012). A review of three-dimensional viscoelastic models with an application to viscoelasticity characterization using nanoindentation Microelectronics Reliability, vol. 52, no. 3, p. 541-558, DOI:10.1016/j.microrel.2011.10.001.

[18] Ferry, J.D. (1980). Viscoelastic properties of polymers, 3rd ed. John Wiley&Sons, New York.

[19] Findley, W.N., Lai, J.S., Onaran, K. (1989). Creep and Relaxation of Nonlinear Viscoelastic Materials with an Introduction to Linear Viscoelasticity, 1st ed., Dover Pub., New York.

[20] Richter, H., Misawa, E.A., Lucca, D.A., Lu, H. (2001). Modeling nonlinear behavior in a piezoelectric actuator. Precision Engineering, vol. 25, no. 2, p. 128137, DOI:10.1016/S0141-6359(00)00067-2.

in amorphous solid polymers. Journal of Rheology, vol.

30, no. 4, p. 781-827, DOI:10.1122/1.549869.

[13] Knauss, W.G., Emri, I. (1981). Non-linear viscoelasticity based on free volume consideration.

Computers and Structures, vol. 13, no. 1-3, p. 123-128, DOI:10.1016/0045-7949(81)90116-4.

[14] Knauss, W.G., Emri, I. (1987). Volume change and the nonlinearly thermo-viscoelastic constitution of polymers. Polymer Engineering and Science, vol. 27, no. 1, p. 86-100, DOI:10.1002/pen.760270113.

[15] Wineman, A.S., Waldron, W.K. (1993). Interaction of nonhomogeneous shear, nonlinear viscoelasticity, and yield of a solid polymer. Polymer Engineering and Science, vol. 33, no. 18, p. 1217-1228, DOI:10.1002/ pen.760331810.

[16] Wang, F. Wang, Y., Fu, B., Lu, H. (2011). Nonlinear viscoelastic nanoindentation of PVAc. Conference

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Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6 Vsebina

Vsebina Strojniški vestnik - Journal of Mechanical Engineering letnik 59, (2013), številka 6 Ljubljana, junij 2013 ISSN 0039-2480 Izhaja mesečno

Razširjeni povzetki člankov Vytautas Ostasevicius, Rimvydas Gaidys, Rolanas Dauksevicius, Sandra Mikuckyte: Študija vibracijskega rezkanja za izboljšanje kakovosti površine materialov, težavnih za obdelavo Tomaž Petrun, Jože Flašker, Marko Kegl: Teoretična in numerična študija dodatnega viskoznega člena v modificiranem elastoplastičnem modelu trenja za simulacijo mokre torne sklopke Pingfa Feng, Chenglong Zhang, Zhijun Wu, Jianfu Zhang: Vpliv hitrosti razenja na deformacijske lastnosti pri nanorazenju safirja, orientiranega v ravnini C Sedat Karabay, Kasım Baynal, Cengiz İğdeli: Iskanje virov ječanja bobnastih zavor gospodarskih vozil po metodi TVA-FMEA: študija primera Peng Gao, Shaoze Yan, Liyang Xie, Jianing Wu: Analiza dinamične zanesljivosti mehanskih komponent na osnovi enakovrednih poti degradacije trdnosti Tomaž Berlec, Janez Kušar, Lidija Rihar, Marko Starbek: Izbira najbolj prilagodljive delovne opreme Iztok Palčič, Marc Pons, Andrea Bikfalvi, Josep Llach, Borut Buchmeister: Analiza uporabe in uporabnikov okolju prijaznih tehnologij v proizvodnih podjetjih Yilmaz Kucuk: Modeliranje nelinearne viskoelastične nanoindentacije PVAc pri različnih hitrostih razbremenjevanja Osebne vesti Doktorske disertacije, znanstveno magistrsko delo, diplomske naloge

SI 67 SI 68 SI 69 SI 70 SI 71 SI 72 SI 73 SI 74 SI 75

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Prejeto v recenzijo: 2012-11-08 Prejeto popravljeno: 2013-03-05 Odobreno za objavo: 2013-03-15

Študija vibracijskega rezkanja za izboljšanje kakovosti površine materialov, težavnih za obdelavo Vytautas Ostasevicius – Rimvydas Gaidys – Rolanas Dauksevicius* – Sandra Mikuckyte 1 Institut

za razvoj visokih tehnologij, Fakulteta za strojništvo in mehatroniko, Tehniška univerza v Kaunasu, Litva

Predmet raziskave v tem članku je vibracijsko (podprto) rezkanje visokotrdnih kovinskih zlitin, s ciljem opredelitve vpliva pogojev vzbujanja rezalnega orodja na kakovost površine obdelovancev z uporabo kvalitativnih in kvantitativnih metod karakterizacije površin.

Na osnovi prototipa orodja je bil pripravljen in eksperimentalno potrjen model orodja za vibracijsko rezkanje na osnovi končnih elementov. Kompleksna zgradba orodja je bila v numeričnem modelu reducirana na eno samo vnaprej zavito konzolo, ki je bila izpostavljena robnim pogojem (vpetja in vzbujanja), ki natančno reproducirajo dejansko vibracijsko orodje. Model je bil uporabljen za napovedovanje resonančne frekvence aksialne vibracijske zvrsti rezkarjev dveh različnih dolžin. Relativno odstopanje med simulirano in izmerjeno resonančno frekvenco je bilo manjše od 2 %. Opravljena je bila vrsta eksperimentov z običajnim in vibracijsko podprtim rezkarjem, pri čemer je bila ugotovljena učinkovitost razvitega orodja za vibracijsko rezkanje pri izboljševanju kakovosti površine nerjavnega jekla (1.4301) in titanove zlitine (GOST 22178-1976).

Površina obdelovancev je bila posneta z vrstičnim elektronskim mikroskopom. Ugotovljeno je bilo, da deli, obdelani z visokofrekvenčnimi vibracijami, nimajo mikrorazpok, karakterizira pa jih fina pravilna struktura.

Rezultati meritev kažejo, da imajo zlitine z visoko trdnostjo po vibracijsko podprtem rezkanju kakovostnejšo površino kot po konvencionalnem procesu (površinska hrapavost je približno za en razred nižja). Najboljša kakovost površine se je izkazala pri rezkanju nerjavnega jekla, kar je mogoče pripisati boljši obdelovalnosti nerjavnega jekla v primerjavi s titanovo zlitino.

Podatki meritev hrapavosti so bili statistično obdelani za podrobnejšo karakterizacijo učinkovitosti orodja za vibracijsko rezkanje. Statistična analiza je pokazala, da imajo največji vpliv na kakovost površine dinamične lastnosti (frekvenca vzbujanja) orodja in način obdelave (s podporo visokofrekvenčnih vibracij ali brez nje). To pomeni, da je treba orodje za povečanje ugodnega vpliva vibracij na proces odrezavanja vzbujati z vibracijami takšne frekvence, ki ustreza aksialni resonančni frekvenci rezkarja. Posledično se ojači amplituda aksialnih vibracij in se okrepi sukanje rezalne konice zaradi sklopitve aksialnih in torzijskih deformacij vijačnega rezkarja.

Rezultati predstavljene raziskave kažejo, da je za doseganje želene zvrsti vibracij rezkarja in s tem optimalne izboljšave kakovosti površine nujno dinamično prilagajanje frekvence vzbujanja vibracijsko podprtega rezkalnega orodja.

Glavni prispevek članka je v tem, da dokazuje učinkovitost predlaganega pristopa k obdelavi zlitin visoke trdnosti, ki lahko pomembno olajša obdelavo trdih in krhkih materialov kot so keramika, steklo in kompoziti.

Predstavljeni eksperimenti vibracijskega rezkanja so bili uspešno izvedeni v pogojih suhe obdelave, to pa pomeni, da bi lahko odrezavanje s podporo vibracij pomagalo tudi pri uvajanju metod minimalnega mazanja v industrijske izdelovalne postopke.

Ključne besede: vibracijsko odrezavanje, model s končnimi elementi, vnaprej zavita konzola, aksialna zvrst, hrapavost

*Naslov avtorja za dopisovanje: Fakulteta za strojništvo in mehatroniko, Tehniška univerza v Kaunasu, Studentu 65, Kaunas, Litva, rolanasd@centras.lt

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Teoretična in numerična študija dodatnega viskoznega člena v modificiranem elastoplastičnem modelu trenja za simulacijo mokre torne sklopke Petrun, T. – Flašker, J. – Kegl, M. Tomaž Petrun1,* – Jože Flašker2 – Marko Kegl2 1 AVL-AST

2

d.o.o., Slovenija Univerze v Mariboru, Fakulteta za strojništvo, Slovenija

Prispevek obravnava teoretično in numerično analizo možnosti vgradnje dodatnega viskoznega člena v modificiran elastoplastični model trenja. Model je bil razvit in validiran za uporabo v dinamičnem modelu suhe torne sklopke, z dodatnim viskoznim členom pa bi model trenja omogočal numerične simulacije tudi za mokre torne sklopke. Uporabljeni so bili različni enostavni viskozni modeli, pri čemer je bil analiziran njihov vpliv na količino prenesenega vrtilnega momenta s tornim kontaktom torne sklopke. Namen je namreč razviti model trenja za model torne sklopke za numerične analize dinamike sistemov teles s popolnoma funkcionalno suho ali mokro torno sklopko, ki bo namenjen vsakodnevni inženirski uporabi. Zaradi omejitev modificiranega elastoplastičnega modela trenja, ki je enodimenzionalen, ter zahtev in omejitev, ki izhajajo iz programskega okolja za implementacijo AVL EXCITE, so bili uporabljeni zgolj enostavni viskozni modeli, pri katerih je viskoznost funkcija strižne napetosti. Viskozni modeli, kjer je viskoznost funkcija temperature, porazdelitve tlaka in podobno, zaradi omejitev modificiranega elastoplastičnega modela trenja niso bili uporabljeni.

Kot osnova za dano raziskavo je služil Carreaujev tekočinski model. Gre za enostaven 1D-tekočinski model, kjer je viskoznost funkcija zgolj strižne napetosti, hkrati pa je model sposoben opisati Newtonske, dilatantne in psevdoplastične tekočine, ki se po lastnostih bistveno razlikujejo. Poleg teoretičnih raziskav so bile izvedene tudi numerične analize vpliva dodatnega viskoznega člena na rezultate dinamike celotnega pogonskega sklopa vozila s popolnoma funkcionalno torno sklopko v smislu vpliva na količino prenesenega vrtilnega momenta s tornim kontaktom in vpliva na inducirane dinamične pojave zaradi trenja. Obravnavan je bil tudi vpliv dodatnega viskoznega člena na modificiran elastoplastični model trenja ter vpliv na reševanje sistema enačb in numerično stabilnost. Predstavljena je tudi merilna proga in potek eksperimentalnih meritev, ki so bile uporabljene za validacijo modificiranega elastoplastičnega modela trenja za uporabo v modelu suhe torne sklopke.

Rezultati numeričnih analiz uporabe modela trenja z dodatnim viskoznim členom so bili primerjani z rezultati validacijskega primera uporabe modificiranega elastoplastičnega modela trenja za suho torno sklopko. Primerjava rezultatov je pokazala, da izbira definicije viskoznosti tekočine v tornem kontaktu odločilno vpliva na količino prenesenega vrtilnega momenta s tornim kontaktom torne sklopke v fazi drsenja – sinhronizacije in razhoda, kar posledično znatno vpliva na potek in trajanje procesa sinhronizacije, kakor tudi na količino prenesenega vrtilnega momenta v razklopljenem stanju torne sklopke. Raziskava je hkrati pokazala tudi slabosti takšnega pristopa, kjer so uporabljeni enostavni viskozni modeli. Ugotovljeno je bilo, da so rezultati drastično odvisni od uporabljene definicije viskoznosti že pri uporabi enostavnih viskoznih modelov, zato uporaba le-teh za podrobne analize dinamike celotnega pogonskega sklopa z mokro torno sklopko ni primerna. V ta namen bo treba uporabiti kompleksne modele viskoznosti, ki poleg strižne napetosti upoštevajo tudi temperaturo, porazdelitev tlaka v kontaktu in podobno.

Raziskava je pokazala tudi, da modificiran elastoplastični model omogoča enostavno dodajanje različnih členov, ki kakor koli prispevajo k rezultirajoči sili trenja v tornem kontaktu, ne da bi s tem vplivali na osnovno strukturo samega elastoplastičnega modela. Z dodajanem členov se spremenijo zgolj prispevki posameznega člena v danih pogojih. Razen viskoznosti je predstavljena tudi teoretična osnova za izračun količine generirane toplote zaradi trenja v tornem kontaktu in njenega vpliva na temperaturo tornega kontakta in okoliških teles, ter posledično vpliva na tribološke lastnosti kontakta. Podane so teoretične osnove in vodilne enačbe za termične preračune, kakor tudi omejitve ciljnega programskega paketa za implementacijo, ki trenutno ne omogočajo izračuna temperature kontakta in njenega vpliva na tribološke lastnosti. Podane so smernice za nadaljnje raziskovalno delo na tem področju.

Ključne besede: dodaten viskozni člen, modificiran elastoplastični model trenja, torna sklopka, dinamika sistema teles, Carreaujev fluidni model, enostavni viskozni modeli SI 68

*Naslov avtorja za dopisovanje: AVL-AST d.o.o., Trg Leona Štuklja 5, 2000 Maribor, Slovenija, tomaz.petrun@yahoo.com

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 69 © 2013 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2012-06-27 Prejeto popravljeno: 2012-12-15 Odobreno za objavo: 2013-01-08

Vpliv hitrosti razenja na deformacijske lastnosti pri nanorazenju safirja, orientiranega v ravnini C Feng, P. – Zhang, C. – Wu, Z. – Zhang, J. Pingfa Feng – Chenglong Zhang* – Zhijun Wu – Jianfu Zhang

Državni laboratorij za tribologijo, Oddelek za natančne instrumente in strojništvo, Univerza Tsinghua, Kitajska

Kristal safirja kot material z dobro prepustnostjo za valovanja se uporablja na številnih področjih, npr. kot substrat za čipe in v laserski tehniki. Stroški natančne in ultranatančne obdelave površine pri proizvodnji safirja so zelo visoki. Zato je nujna raziskava možnosti obdelave safirja z odvzemanjem materiala in njegovih deformacijskih lastnosti, katere rezultati bodo uporabni pri proizvodnji izdelkov iz safirja. Preskus razenja je proces, ki je analogen obdelavi z odvzemanjem in je bil zato uporabljen za preučitev deformacijskih lastnosti in lastnosti obdelave z odvzemanjem materiala.

V članku so predstavljeni preskusi nanorazenja, ki razkrivajo deformacijske lastnosti in lastnosti obdelave z odvzemanjem materiala za safir, orientiran v ravnini C (0001). Preučen je vpliv hitrosti razenja na lastnosti plastične in krhke deformacije. Izbrane so bile hitrosti razenja 2, 4, 8 in 16 μm/s. Test z nanomehanskim preskusnim sistemom je bil opravljen v pogojih naraščajoče sile od 40 μN do 200 mN. V študiji je bilo uporabljeno Berkovičevo nanovtiskalo, lastnosti plastične in krhke deformacije pa so bile analizirane s pomočjo vrstičnega elektronskega mikroskopa (SEM). Za analizo zaostalih napetosti v deformacijskem območju brazde in za ugotavljanje mehanizma odvzema materiala pri safirju je bil uporabljen Ramanski spektroskop. Predstavljena je tudi primerjava globinskega profila površine in lastnosti brazd, ki nastanejo pri različnih hitrostih razenja.

Analiza globinskega profila površine in mikrostrukture s SEM je pokazala, da se globina in širina brazde povečujeta sorazmerno z uporabljeno silo. Deformacije ob vsakem postopku razenja je mogoče opisati kot plastične deformacije, plastične deformacije s povečano globino brazde in pojavom mikrorazpok, ter krhke deformacije s pojavom krušenja in poškodb. Primerjava globinskih profilov površine in lastnosti brazd, ki nastanejo pri različnih hitrostih razenja, je pokazala, da ima hitrost razenja značilen vpliv na deformacijske lastnosti safirja, orientiranega v ravnini C. Globinski profil površine kaže, da se s povečevanjem hitrosti razenja povečata tudi kritična sila in globina, analiza brazde pa pokaže več linij zdrsa in manj vdrtin zaradi krušenja. To pomeni, da večja hitrost razenja prinaša večji delež plastičnih deformacij pri razenju safirja, orientiranega v ravnini C. Ramanski spektri, izmerjeni z mikroramanskim spektrometrom, imajo v brazdi Ramanske vrhove pri nižjem valovnem številu kot zunaj brazde. Sklepamo lahko, da se v brazdi med razenjem pojavljajo natezne napetosti.

Natezne napetosti so glavni povzročitelj napak zloga, dislokacijskih zank in dislokacijskih zdrsov, ki štejejo med elemente mehanizma plastične deformacije. Zato je mogoče sklepati, da so med razenjem natezne napetosti v brazdi zaslužne za plastične deformacije safirja.

Za razjasnitev mehanizmov različnih deformacijskih vedenj vzorca safirja med postopkom razenja pri različnih hitrostih je obravnavan vpliv hitrosti razenja na preoblikovalno hitrost in trdoto, kakor tudi občutljivost na preoblikovalno hitrost z ozirom na naraščajočo hitrost razenja. Iz diskusije mehanizmov različnih deformacijskih vedenj pri različnih hitrostih razenja sledi, da lahko povečanje hitrosti razenja izboljša preoblikovalno hitrost in trdoto ter učinkovito omeji pojav in rast razpok, ki jih je zato manj in so manjše, plastične deformacije pa so tako bolj primerljive s tistimi pri nižjih hitrostih razenja. Sledi, da je za izboljšanje površinske obdelave safirja smiselno povečati rezalno hitrost, ki je analogna hitrosti razenja.

Ključne besede: safir, orientiran v ravnini C (0001), preskus nanorazenja, hitrost razenja, plastična deformacija, krhka deformacija, preostale napetosti, preoblikovalna hitrost

*Naslov avtorja za dopisovanje: Državni laboratorij za tribologijo, Oddelek za natančne instrumente in strojništvo, Univerza Tsinghua, Kitajska, zcl08thu@gmail.com

SI 69

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 70 © 2013 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2012-09-24 Prejeto popravljeno: 2013-03-13 Odobreno za objavo: 2013-03-15

Iskanje virov ječanja bobnastih zavor gospodarskih vozil po metodi TVA-FMEA: študija primera Karabay, S. – Baynal, K. – İğdeli, C. Sedat Karabay1 – Kasım Baynal2,* – Cengiz İğdeli2 1 Univerza v Kocaeliju, Tehniška fakulteta, Oddelek za strojništvo, Turčija 2 Univerza v Kocaeliju, Tehniška fakulteta, Oddelek za industrijski inženiring, Turčija

V članku je predstavljena strategija za iskanje glavnih vzrokov reklamacij zaradi ječanja zadnjih bobnastih zavor gospodarskih vozil, ki jo je uporabil proizvajalec avtomobilov. Pri tem je uporabil podatke, pridobljene na reklamiranih vozilih, vozilih na proizvodni liniji in laboratorijskem preizkuševališču zavor.

Iz bobnastih zavor pogosto prihaja močan zvok ječanja, zaradi katerega kupci uveljavljajo reklamacije. Ječanje je posledica vzbujanja konstrukcije vozila in komponent vzmetenja z zavornim sistemom, ki med zaviranjem povzroči močno slišno in občutno reakcijo.

V članku so opisane eksperimentalne študije nizkofrekvenčnega ječanja bobnastih zavor, ki je proizvajalcu povzročilo visoke stroške reklamacij. Najprej so bili identificirani pogoji okolice in reproducirano ječanje. Za simulacijo ječanja so bili z vozila demontirani problematični deli zavor in vzmetenja ter nato uporabljeni v testni postavitvi za ugotavljanje glavnih virov hrupa, vibracij in grobega delovanja v laboratorijskih pogojih. Preskusi na vozilih so bili opravljeni v tovarni in v prometu. S pomočjo strokovnjakov za tehniko vozil so bile opravljene meritve ječanja in vibracij za lokalizacijo virov hrupa na reklamiranih vozilih. Kot je opisano v nadaljevanju, je bila za izvedbo načrtovane strategije podrobne študije uporabljena različna oprema za analizo hrupa, vibracij in grobega delovanja s pripadajočim priborom. Pri študiji sta bili uspešno uporabljeni metodi TVA (analiza celotne vrednosti) in FMEA (analiza možnih napak in njihovih posledic). Načrtovana strategija iskanja glavnih vzrokov je sistematično izločila sekundarne in terciarne vplive ječanja zavor. Meritve vibracij in ječanja v okviru ugotavljanja glavnih vzrokov so bile izvedene in interpretirane v skladu s kartami TVA in FMEA. Preučena je bila občutljivost materiala oblog zavornih čeljusti na različne okoljske pogoje. Končno je obravnavan tudi mehanizem ječanja sistema zavornega bobna in podrobna rešitev za odpravo nizkofrekvenčnega ječanja zavornih bobnov.

Ugotovljeno je bilo, da je ječanje zavor posledica dinamične nestabilnosti nekonzervativnih tornih sil.

Vibracije zaradi delovanja tornih sil na deformirane kontaktne površine povzročajo premike tornih površin oblog in težave z ječanjem zavor. Količina dela vibrirajočih komponent zavor se zato razlikuje pri vibracijskem gibanju naprej in nazaj, zaradi česar se pojavita dinamična nestabilnost sistema in zvok ječanja zavor. Raziskava je pokazala, da je zvok ječanja zavor odvisen od koeficienta trenja ter od lege kontaktnih površin tornega materiala in kovinske površine bobna. Površina obloge zaradi predolgega časa sintranja kompozitnega materiala čezmerno kristalizira in zmanjša se učinkovitost zaviranja. Obloga zaradi manjšega zavornega učinka ob stiku z notranjo površino bobna pritiska na boben, vibracije suhega trenja pa povzročajo zvok ječanja.

Odpirajo se možnosti za razvoj novega pristopa za podjetja, ki se ukvarjajo z množično proizvodnjo tehničnih izdelkov, vključno s tehnikami za hitro in natančno preizkušanje termomehansko oblikovanih kompozitnih materialov.

Ugotovljeno je bilo, da je čas sintranja zavornih oblog ključen za tiho delovanje bobnastih zavor lahkih gospodarskih vozil. Pri izdelavi kompozitnih oblog je zato treba upoštevati postopek obdelave, ki je bil potrjen v laboratorijskih preizkusih. Z drugimi besedami, za 5 °C previsoka temperatura in za 190 sekund predolg čas sintranja sta morda povzročila za več milijonov dolarjev škode.

Ključne besede: bobnasta zavora, TVA, FMEA, NVH, koeficient trenja, ječanje zavor, zavorne obloge

SI 70

*Naslov avtorja za dopisovanje: Univerza v Kocaeliju, Tehniška fakulteta, Oddelek za industrijski inženiring, Kocaeli, Turčija, kbaynal@yahoo.com

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 71 © 2013 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2012-04-21 Prejeto popravljeno: 2012-10-29 Odobreno za objavo: 2013-02-24

Analiza dinamične zanesljivosti mehanskih komponent na osnovi enakovrednih poti degradacije trdnosti Peng 1 Univerza

Gao, P. – Yan, S. – Xie, L. – Wu, J. – Shaoze Yan1,* – Liyang Xie2 – Jianing Wu1

Gao1

Tsinghua, Oddelek za strojništvo, Državni laboratorij za tribologijo, Kitajska 2 Severovzhodna univerza, Šola za strojništvo, Kitajska

Članek podaja predloge modelov dinamične zanesljivosti, ki vključujejo naključno naravo poti degradacije trdnosti pri ocenjevanju dinamične zanesljivosti, kakor tudi analizo vpliva variabilnosti statističnih parametrov lastnosti materiala na zanesljivost in stopnjo odpovedi. Pot degradacije trdnosti je težko natančno določiti zaradi naključnosti obremenitev mehanskih komponent. Negotovost trdnosti v procesu degradacije je zato upoštevana s porazdelitvijo trdnosti pri vsaki aplikaciji obremenitve, zaradi česar lahko pride do napak pri izračunavanju zanesljivosti, ker ni upoštevana korelacija preostale trdnosti pri vsaki aplikaciji obremenitve v poti degradacije trdnosti. V članku so kot odgovor na ta problem predstavljeni dinamični modeli zanesljivosti mehanskih komponent z utrujenostno odpovedjo na osnovi enakovrednih poti degradacije trdnosti. V predlaganih modelih sta uporabljena mehanizem degradacije mehanskih komponent z načinom odpovedi ter teorija stohastičnih procesov za obvladovanje negotovosti procesa degradacije trdnosti, ki ga določata tako porazdelitev lastnosti materiala kot statistične lastnosti obremenitev v procesu obremenjevanja.

Za vzorčni primer pri validaciji učinkovitosti in natančnosti predlaganih modelov so bili izbrani eksplozivni elementi, ki se uporabljajo pri lansiranju satelitov. V numeričnih primerih je bila uporabljena simulacija Monte Carlo za validacijo učinkovitosti in natančnosti predlaganih modelov dinamične zanesljivosti. V simulaciji Monte Carlo se degradacija trdnosti simulira na osnovi mehanizma degradacije mehanskih komponent, naključne obremenitve pa so generirane na podlagi porazdelitve verjetnosti.

Rezultati kažejo, da se zanesljivost, izračunana po predlagani metodi, dobro ujema z rezultati simulacije Monte Carlo. Pri izračunavanju zanesljivosti na podlagi porazdelitve trdnosti pri vsaki obremenitveni aplikaciji lahko nastopi večja napaka tudi zato, ker je zanemarjena korelacija preostale trdnosti pri vsaki obremenitveni aplikaciji v poti degradacije trdnosti. Različni parametri materiala, uporabljeni v predlaganih modelih zanesljivosti, imajo tudi različen vpliv na dinamične lastnosti zanesljivosti in stopnjo odpovedi mehanskih komponent. Tradicionalno velja, da velik raztros začetne trdnosti povzroči manjšo zanesljivost. Pri obravnavi degradacije trdnosti pa ima raztros začetne trdnosti različen vpliv na zanesljivost mehanskih komponent v različnih fazah življenjskega cikla. Omeniti je treba tudi to, da statistične značilnosti začetne trdnosti pomembno vplivajo na dinamiko stopnje odpovedi mehanskih komponent. Naklon krivulje stopnje odpovedi za mehanske komponente v naključnem obdobju odpovedi se zmanjšuje s povečevanjem srednje vrednosti in raztrosa začetne trdnosti.

Predlagani modeli omogočajo analizo dinamične zanesljivosti mehanskih komponent in so uporabni za kvantitativno analizo vpliva variabilnosti statističnih parametrov značilnosti materiala na zanesljivost in stopnjo odpovedi komponent. V prihodnjih raziskavah bodo v modele zanesljivosti vključene dodatne spremenljivke za natančnejše napovedi. Avtorji se ukvarjajo tudi s širitvijo predlagane metode na področje optimizacije konstrukcij na osnovi zanesljivosti.

Ključne besede: dinamična zanesljivost, korelacija, preostala trdnost, mehanske komponente, pot degradacije trdnosti, stopnja odpovedi

*Naslov avtorja za dopisovanje: Univerza Tsinghua, Oddelek za strojništvo, Državni laboratorij za tribologijo, Peking, Kitaljska, yansz@tsinghua.edu.cn

SI 71

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 72 © 2013 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2013-01-08 Prejeto popravljeno: 2013-02-07 Odobreno za objavo: 2013-04-08

Izbira najbolj prilagodljive delovne opreme Berlec, T. – Kušar, J. – Rihar, L. – Starbek, M. Tomaž Berlec – Janez Kušar* – Lidija Rihar – Marko Starbek Univerza v Ljubljani, Fakulteta za strojništvo, Slovenija

V članku je prikazan postopek, kako med na trgu razpoložljivo delovno opremo (delovna sredstva) izbrati tisto, ki bo najbolj prilagodljiva glede na spremembe proizvodnega programa podjetja. Delovna oprema je ključni faktor ustvarjanja vrednosti izdelkov, zato se oblikovanje in izbor le te vse bolj orientira na njene zmožnosti spreminjanja.

Zmožnosti spreminjanja delovne opreme v obliki fleksibilnosti in reaktiviranja pa so definirane z univerzalnostjo, mobilnostjo, modularnostjo, kompatibilnostjo in gospodarnostjo delovne opreme.

Proizvodni program malih in srednje velikih podjetij je podvržen kontinuiranemu spreminjanju in vse večjim zahtevam kupcev po čim krajšem dobavnem času, čim boljši kakovosti in čim nižji ceni izdelkov. V podjetjih se srečujejo z nestabilnim povpraševanjem glede na vrsto in količino izdelkov. Nestabilno povpraševanje po izdelkih pa zahteva pogosto spreminjanje funkcij in razmestitve delovne opreme. Pogosto se pri tem pojavi problem zmožnosti spreminjanja oziroma prilagajanja delovne opreme spremembam v proizvodnem programu podjetja.

Delovna oprema je nosilec ustvarjanja dodane vrednosti izdelkom in zato ključni faktor vsake proizvodnje.

Oblikovanje in izbor delovne opreme se danes vse bolj orientira na zmožnost spreminjanja le te in to predvsem zaradi nezanesljivih prognoz trga. Zmožnost spreminjanja delovne opreme presoja primernost delovnih sredstev, da se lahko izvede prilagoditev le teh glede na interno in eksterno vzbujene tehnološke, strukturne in organizacijske spremembe in to ob čim manjših stroških. Zmožnost spreminjanja delovne opreme se sestoji iz deleža fleksibilnosti in deleža reaktiviranja delovne opreme.

Fleksibilnost pomeni predimenzioniranje delovne opreme glede na trenutno zahtevane funkcije, učinek in točnost. Omogoča obvladovanje prihodnjih, vnaprej planiranih scenarijev, torej dodatne funkcije so že razpoložljive in se lahko po potrebi aktivirajo. Od delovne opreme se zahteva, da se lahko prilagodi na novo situacijo in nove potrebe in to z majhnimi stroški.

Reaktiviranje delovne opreme pa pomeni zahtevano sposobnost, da se lahko reagira tudi na nove, v fazi planiranja nepoznane zahteve. Reaktiviranje delovne opreme se realizira z zmožnostjo rekonfiguracije delovne opreme. Cilj reaktiviranja delovne opreme je modularna, rekonfigurirana delovna oprema.

Pri oblikovanju oziroma izboru delovne opreme je potrebno upoštevati katere zahteve mora delovna oprema izpolniti glede na tehnološke funkcije in zmožnost spreminjanja. Potrebna je točna specifikacija profila zmožnosti spreminjanja delovne opreme.

Za izbor najbolj prilagodljive opreme se lahko uporabi razširjeno metodo analize koristnosti, ki se izvede v sedmih korakih: 1. Določitev ciljev.

2. Oblikovanje drevesa kriterijev.

3. Določitev uteži kriterijev 4. Pridobitev podatkov o izpolnjevanju kriterijev ponujane opreme 5. Pred-izbor delovne opreme (izmed vseh ponudb se izbere 3, ki v največji meri izpolnjujejo kriterije – dosežejo največjo korist).

6. Pridobitev dodatnih (podrobnih) podatkov o izpolnjevanju kriterijev za izbrane tri ponudbe.

7. Končni izbor delovne opreme (izmed izbranih 3 ponudb se izbere tisto, ki doseže največjo korist).

Prikazan je tudi praktični primer izbora najprimernejše CNC stružnice. Določenih je bilo 50 kriterijev, za katere so bili v pred-izboru pridobljeni podatki za 9 ponudb. Izmed teh ponudb so bile izbrane 3, za katere so ponudniki dopolnili podatke. Končni izbor je pokazal, katera izmed izbranih ponudb je najprimernejša.

Rezultati, prikazani v prispevku, so pokazali, da se z razširjeno metodo analize koristnosti zelo učinkovito izbere najprimernejšo opremo in to na osnovi več kriterijalnega odločanja.

Ključne besede: delovna oprema, zmožnost spreminjanja, fleksibilnost, reaktiviranje, univerzalnost, mobilnost, modularnost, kompatibilnost

SI 72

*Naslov avtorja za dopisovanje: Univerza v Ljubljani, Fakulteta za strojništvo, Aškerčeva 6, 1000 Ljubljana, Slovenija, janez.kusar@fs.uni-lj.si

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 73 © 2013 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2012-10-11 Prejeto popravljeno: 2013-02-11 Odobreno za objavo: 2013-03-14

Analiza uporabe in uporabnikov okolju prijaznih tehnologij v proizvodnih podjetjih Palčič, I. – Pons, M. – Bikfalvi, A. – Llach, J. – Buchmeister, B. Iztok Palčič1,* – Marc Pons2 – Andrea Bikfalvi2 – Josep Llach2 – Borut Buchmeister1 1 Univerza

v Mariboru, Fakulteta za strojništvo, Slovenija 2 Univerza v Gironi, Španija

Namen prispevka je prikazati uporabo tehnologij za zniževanje rabe energije in potrebne porabe virov v proizvodnji. Cilj je tudi prispevati k identifikaciji in razumevanju značilnosti proizvodnih podjetij, ki uporabljajo omenjene tehnologije.

Energijska učinkovitost v proizvodni industriji je izjemno pomembna, saj proizvodna podjetja letno porabijo 75 % celotne svetovne potrošnje premoga, 44 % naravnih plinov in 20 % naftnih derivatov. Kljub temu večje število raziskav dokazuje, da imajo proizvodna podjetja pri izboljševanju energijske učinkovitosti še zmeraj velike rezerve, še posebej glede vpliva na ekonomski vidik poslovanja podjetij.

V okviru naše raziskave smo izbrali osem tehnologij za zniževanje rabe energije v proizvodnji in dve tehnologiji za zniževanje porabe virov (materialov) v proizvodnji. Izbrane tehnologije smo opazovali z dveh vidikov: pogostosti uporabe in nivoja uporabe glede na potencial izkoriščenosti. Proizvodna podjetja smo razdelili glede na njihovo tehnološko intenzivnost (v skladu s klasifikacijo NACE) in glede na relativno energijsko učinkovitost, kjer so podjetja v anketi ocenila, ali so bolj ali manj učinkovita od svojih konkurentov.

Naša raziskava temelji na podatkih iz največje evropske ankete o proizvodni dejavnosti in vključuje podatke iz Slovenije in Španije. V anketi sprašujemo podjetja o proizvodnih strategijah, rabi tehniških in organizacijskih inovacij, selitvi proizvodnje, tipih proizvodnje in izdelkov, konkurenčnih kriterijih, kvalifikacijah in izobrazbi zaposlenih. Zbiramo tudi podatke o produktivnosti, fleksibilnosti, kakovosti, donosih ipd. Anketo smo v zadnji verziji temeljito posodobili, dodali nekaj novih perečih tematik, predvsem s področja učinkovite rabe energije.

Rezultati kažejo, da je v povprečju uporaba tehnologij za zniževanje rabe energije in potrebne porabe virov v proizvodnji še zmeraj precej nizka. Ugotovili smo, da podjetja v visoko tehnoloških industrijah posvečajo manj pozornosti energijski učinkovitosti kot podjetja iz nizko tehnoloških industrij. Izmed vseh obravnavanih tehnologij sta bili na drugem in tretjem mestu po pogostosti uporabe tehnologiji za zniževanje porabe virov (materialov) v proizvodnji (recikliranje materialov in obnovitev izdelka po preteku življenjske dobe). Še posebej izrazito ju uporabljajo v podjetjih iz nizko tehnoloških industrij. Statistične analize so tudi pokazale, da so energijsko bolj učinkovita tista proizvodna podjetja, ki v povprečju uporabljajo več okolju prijaznih tehnologij. Pri opazovanju zgolj tistih tehnologij z visokim nivojem izkoriščenosti, smo ugotovili, da so tehnologije bolje izkoriščene pretežno v podjetjih iz nizko tehnoloških industrij. To ponovno nakazuje na dejstvo, da podjetja v nizko tehnoloških industrijah posvečajo več pozornosti energijski učinkovitosti kot podjetja iz višje tehnoloških industrij. Sočasno pa smo ugotovili, da so tista podjetja, ki močno izkoriščajo potencial obravnavanih tehnologij, tudi bolj energijsko učinkovita od svojih konkurentov. Zanimala nas je tudi povezava med uporabo okolju prijaznih tehnologij in sistemi ravnanja z okoljem (ISO 14000). Ugotovili smo, da obstaja pozitivna korelacija med uporabo sistema ISO 14000 in pogostostjo uporabe analiziranih tehnologij.

Naša raziskava je ena redkih, kjer ugotavljamo vpliv uporabe izbranih sodobnih okolju prijaznih tehnologij na energijsko učinkovitost podjetij. Podane relacije so lahko v veliko pomoč pri strateškem odločanju v proizvodnih podjetjih pri razmišljanju o investicijah v nove tehnologije.

Ključne besede: energijska učinkovitost, proizvodno podjetje, tehnologije za varčevanje z energijo, tehnologije za varčevanje z materialom, evropska anketa o proizvodni dejavnosti

*Naslov avtorja za dopisovanje: Univerza v Mariboru, Fakulteta za strojništvo, Smetanova ulica 17, 2000 Maribor, Slovenija, iztok.palcic@uni-mb.si

SI 73

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 74 © 2013 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2012-06-12 Prejeto popravljeno: 2013-02-11 Odobreno za objavo: 2013-03-13

Modeliranje nelinearne viskoelastične nanoindentacije PVAc pri različnih hitrostih razbremenjevanja Kucuk Y. Yilmaz Kucuk

Univerza Bartin, Fakulteta za strojništvo, Turčija

Nelinearno viskoelastično vedenje polimernih materialov se v zadnjih letih preizkuša predvsem z metodo nanoindentacije. Nanoindentacija, znana tudi kot vtiskavanje z zaznavanjem globine, je učinkovita tehnika za merjenje lokalnih mehanskih lastnosti na mikro- in nanoravni. Oliver in Pharr sta razvila pristop za merjenje mehanskih lastnosti, kot sta modul elastičnosti in trdote brez neposrednega opazovanja projicirane površine vtiska.

Doemer in Nix sta nakazala, da je trdoto in modul elastičnosti mogoče izračunati tudi iz podatkov obremenilne krivulje in naklona razbremenilne krivulje.

Tehnika nanoindentacije se je dobro uveljavila pri izotropnih materialih, ki niso časovno odvisni. V zadnjih letih pa je vse več zanimanja tudi za analizo odziva viskoelastičnih materialov pri nanoindentacijskem preizkusu. V tej študiji je bila uporabljena tehnika nanoindentacije za določitev nelinearnega viskoelastičnega vedenja polivinil acetata (PVAc) s temperaturo steklastega prehoda 29 °C. Podatki nanoindentacijskega preskusa so sestavljeni iz dveh glavnih delov: obremenitve in razbremenitve. Večina analitičnih študij polimerov se je ukvarjala samo z obremenitvenim delom odziva viskoelastičnih materialov, čeprav se veliko informacij skriva tudi v razbremenitvenem delu odziva. Oliver-Pharrova metoda npr. ni primerna za razne polimere zlasti če krivulja razbremenitve pri analizi rezultatov nanoindentacije izkazuje negativno togost oz obliko nosu. Ta problem se pojavlja zlasti pri majhnih hitrostih obremenjevanja in razbremenjevanja, ki se uporabljajo pri nanoindentaciji viskoelastičnih materialov.

V literaturi je opisanih veliko nelinearnih modelov viskoelastičnosti, noben pa ne pokriva vseh primerov.

Modeli materialov so lahko izdelani na osnovi empiričnih opažanj ali fizikalnih zakonitosti. Fizikalne modele za nelinearne viskoelastične materiale je zelo težko razvijati in uporabljati. V tej študiji je bil za izhodišče zaradi enostavnosti in izvedljivosti vzet fenomenološki model po študiji Marina in Paa. Ta model je v splošnem linearnem primeru štiriparametrski Burgersov model, oblikovan z zaporedno vezavo Maxwellove trdine in Voigtove enote.

Burgersov nelinearni viskoelastični model, ki sestoji iz Maxwellove in Voigtove enote, je bil prirejen za paket ABAQUS-FEA s subrutino aUMAT. Preskusi nanoindentacije so bili opravljeni z Berkovičevim vtiskalom pri konstantni hitrosti obremenitve in različnih hitrostih razbremenitve. Nabor parametrov nelinearnega Burgersovega modela je bil prilagojen krivulji obremenitev–deformacij pri nanoindentaciji z vršno obremenitvijo 8 mN in hitrosti obremenjevanja/ razbremenjevanja 0,05 mN/s, vse nadaljnje simulacije pa so bile izvedene s temi parametri. Po zlaganju krivulj obremenitev–deformacij s simuliranimi odzivi so bili pridobljeni parametri nelinearnih Maxwellovih in Voigtovih elementov za analizo nelinearnega viskoelastičnega vedenja PVAc. Rezultati kažejo, da ima PVAc negativno togost pri majhnih hitrostih razbremenitev (0,05 in 0,1 mN/s). Če je hitrost razbremenitve večja od 0,1 mN/s, je vedenje PVAc bolj elastično. Doseženo je bilo dobro ujemanje med simulacijami in eksperimenti nanoindentacije pri različnih hitrostih razbremenitve. Rezultati kažejo, da je z nelinearnim Burgersovim modelom mogoče popisati nelinearno vedenje PVAc pri različnih hitrostih razbremenitve od 0,05 do 0,8 mN/s. Povečanje hitrosti razbremenitve omogoča tudi povečanje zmogljivosti obnovitve vzorca PVAc. Določiti pa je mogoče tudi nelinearne viskoelastične lastnosti PVAc kot so prehodni in stacionarni parametri.

Ključne besede: nanoindentacija, nelinearno viskoelastično vedenje, nelinearen Burgersov model, polimer, PVAc, metoda končnih elementov, viskoelastičnost, simulacija

SI 74

*Naslov avtorja za dopisovanje: Univerza Bartin, Fakulteta za strojništvo, Turčija, yilmazkucuk75@gmail.com

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 75-77 Osebne objave

Doktorske disertacije, znanstveno magistrsko delo, diplomske naloge

DOKTORSKE DISERTACIJE Na Fakulteti za strojništvo Univerze v Mariboru so obranili svojo doktorsko disertacijo: ● dne 20. maja 2013 Tomaž PETRUN z naslovom: »Razvoj matematičnega modela trenja za simulacijo dinamike torne sklopke v sistemu teles« (mentor: prof. dr. Jože Flašker); Doktorska disertacija obravnava razvoj in validacijo modela trenja za uporabo v modelu popolnoma funkcionalne torne sklopke v programskem paketu AVL EXCITE, ki bo omogočal numerične simulacije dinamike sistema teles celotnega pogonskega sklopa. Na podlagi zahtev ciljnega programskega paketa za implementacijo in zahtev, ki izhajajo iz aplikacije in pregleda obstoječe literature, je bil izbran elasto-plastični model trenja kot najprimernejši model za nadaljnjo raziskovalno delo. S pomočjo numeričnih testov je bilo ugotovljeno, da izbran model ne ustreza zahtevam dane aplikacije zaradi nezveznosti modela za določene naključne kombinacije parametrov modela in vhodnih podatkov. Zaradi tega je bila narejena modifikacija elasto-plastičnega modela trenja, ki je odpravila pomanjkljivosti. Razvit modificiran elasto-plastični model je bil validiran na podlagi eksperimentalnih meritev. Za izvedbo eksperimentalnih meritev je bila izdelana posebna merilna proga s poenostavljeno torno sklopko. Merilna proga omogoča meritve prenesenega vrtilnega momenta, normalne sile in vrtilnih hitrosti pod obratovalnimi pogoji, ki se zelo približajo realnim obratovalnim pogojem v pogonskem sklopu z motorjem z notranjim zgorevanjem. Primerjava rezultatov eksperimentalnih meritev in pripadajočih numeričnih simulacij sistema teles z uporabo razvitega modificiranega elasto-plastičnega modela trenja v modelu popolnoma funkcionalne torne sklopke je pokazala zelo dobro ujemanje. Ugotovljeno je bilo, da je razvit modificiran elasto-plastični model trenja sposoben pravilno opisati vse obratovalne faze torne sklopke. Model je sposoben zvezno izračunati količino dejansko prenesenega vrtilnega momenta v vseh obratovalnih fazah torne sklopke. Model je tudi sposoben pravilno upoštevati inducirane dinamične pojave zaradi trenja, kot je na primer „stick/slip“ pojav. Razvit modificiran elasto-plastični model trenja ustreza vsem podanim zahtevam za dano aplikacijo in vsem zahtevam za implementacijo v ciljni programski paket; ● dne 21. maja 2013 Marko REIBENSCHUH z naslovom: »Inteligentni sistem za korekcijo rezalnih

parametrov pri frezanju gravur orodij« (mentor: prof. dr. Franci Čuš); Kvaliteta obdelane površine in stroški obdelave sta glavna faktorja, s katerima podjetje dosega konkurenčnost na svetovnem tržišču izdelave preoblikovalnih orodij. Z uporabo najnovejših materialov in modernih postopkov obdelave podjetja poskušajo doseči optimalno izkoriščenost strojev in orodij ter si tako zagotoviti tržno prednost. Izbira rezalnih parametrov ima pri obdelavi velik vpliv na stroške, zato veliko raziskovalcev izvaja postopke off-line optimizacije. Na področju materialov se razvijajo vedno bolj kompleksni materiali, med katere prištevamo tudi gradientne materiale. Znane offline rešitve optimizacij ne omogočajo zadovoljivih rezultatov pri optimizaciji rezalnih parametrov. V raziskavi smo razvili, izdelali in vpeljali v proizvodnjo orodij nov inteligentni sistem za korekcijo rezalnih parametrov pri frezanju gravur (ISK) za obdelavo standardnih in gradientnih materialov. Predstavljen sistem omogoča optimizacijo rezalnih parametrov v on-line načinu in izključuje negativne vplive človeškega faktorja v proizvodnjo. Z vizualnim nadzorom odrezka, obrabe orodja in merjenjem rezalnih sil sistem samodejno izvaja korekcijo rezalnih parametrov (vrtilne frekvence in podajalne hitrosti) ter tako omogoča optimalno izkoriščenost stroja, orodja in hkrati omogoča natančnejšo izdelavo. Raziskava zajema predstavitev problema, iskanje rešitev, sestavljanje in umerjanje sistema ter serijo preizkusov, kjer je sistem testiran v laboratorijskih in delavniških razmerah. Izdelan je tudi uporabniški vmesnik, ki na podlagi govornega vodenja operaterju stroja omogoča hitrejši odziv na nepredvidljive spremembe med obdelavo. V zaključku raziskave so podani rezultati, ki utemeljujejo in dokazujejo uporabnost sistema. Podani so tudi predlogi za nadaljnji razvoj in raziskave; ● dne 27. maja 2013 Jasmina FILIPIČ z naslovom: »Vpliv magnetnega polja na bakterije v bioloških čistilnih napravah in na pretvorbo dušika« (mentor: izr. prof. dr. Vanja Kokol); V doktorski disertaciji predstavljamo študijo čiščenja odpadne vode z vidika odstranjevanja dušikovih spojin s pomočjo srednje močnega statičnega magnetnega polja (SMP). Veljavna zakonodaja in visoki stroški okoljskih dajatev za onesnaževanje voda, namreč težijo k iskanju novih naprednih tehnologij čiščenja odpadne vode. V dosedanjih raziskavah je bilo ugotovljeno, da se pod vplivom SMP, poveča odstranjevanje organskih substratov iz odpadne vode. Naše raziskave kažejo, SI 75

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 75-77

da SMP gostote, B = (30, 50) mT (toda ne B = 10 mT) pozitivno vpliva na odstranjevanje dušikovih spojin v SBR reaktorjih. Amonij-oksidacijska hitrost se pri SMP gostote, B = 50 mT drastično poveča, in sicer za w = 85±5%. Glede na pozitiven učinek SMP na čiščenje odpadne vode, smo v nadaljevanju raziskave preučili vpliv SMP na tri bakterije odpadne vode ˗ Nitrosomonas europaea, Escherichia coli in Pseudomonas putida. Rezultati raziskave kažejo, da SMP (B = 17 mT) negativno vpliva na rast bakterij E.coli in P.putida, toda pozitivno na dehidrogenazno aktivnost in koncentracijo znotrajceličnega ATP. Inhibitorni vpliv na omenjeni bakteriji je bil največji pri njunih optimalnih temperaturah rasti.

V okviru doktorske naloge smo kot prvi dokazali, da SMP (B = 17 mT) pozitivno vpliva na oksidacijo amonijevega dušika pri bakteriji N.europaea. SMP prav tako pozitivno vpliva na mešano združbo amonijoksidirajočih bakterij v odpadni vodi, s povečanjem za w = 40% pod vplivom SMP in na čisto kulturo N.europaea s povečanjem števila bakterij po sedmih dneh izpostavitve SMP za w = 60%, v primerjavi s kontrolnimi vzorci. Iz rezultatov je razvidno, da lahko z ustrezno gostoto SMP, časom izpostavitve in temperaturo znatno povečamo odstranjevanje dušikovih spojin iz odpadne vode ter pospešimo rast enega izmed najpomembnejših nitrifikatorjev odpadne vode – Nitrosomonas europaea.

ZNANSTVENO MAGISTRSKO DELO Na Fakulteti za strojništvo Univerze v Mariboru je z uspehom zagovarjala svoje magistrsko delo: ● dne 17. maja 2013 Ramona IRGOLIČ z naslovom: »Ponovna uporaba obdelane odpadne vode v procesu barvanja« (mentorica: prof. dr. Alenka Majcen Le Marechal); DIPLOMSKE NALOGE Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv univerzitetni diplomirani inženir strojništva: dne 28. maja 2013: Dušan BREGANT z naslovom: »Razvoj in izdelava merilne naprave za analizo pretočnosti mikro prahov« (mentor: prof. dr. Igor Emri); Tilen NAGODE z naslovom: »Modeliranje in simulacija piezoelektričnega aktuatorja za hidravlični preklopni ventil« (mentor: izr. prof. dr. Niko Herakovič); Žiga TOMC z naslovom: »Analiza vplivnih parametrov in načrtovanje montažnega procesa skiroja« (mentor: izr. prof. dr. Niko Herakovič); dne 29. maja 2013: SI 76

Benjamin HUSKIĆ z naslovom: »Celovito produktivno vzdrževanje« (mentor: prof. dr. Marko Starbek, somentor: izr. prof. dr. Janez Kušar); Tomaž PIRNAT z naslovom: »Projektno vodenje naročil« (mentor: prof. dr. Marko Starbek, somentor: izr. prof. dr. Janez Kušar); Gregor TESKAČ z naslovom: »Projekt postavitve vrtine za oskrbo s termalno vodo« (mentor: prof. dr.

Marko Starbek, somentor: izr. prof. dr. Janez Kušar); Nejc VRHOVNIK z naslovom: »Zagotavljanje enakomerne površinske temperature brizgalnega trna z uporabo toplotne cevi« (mentor: prof. dr. Iztok Golobič); Andrej ZUPANČIČ z naslovom: »Analiza življenjskega cikla tehnologije rezanja z lednim abrazivnim vodnim curkom« (mentor: doc. dr. Franci Pušavec, somentor: asist. dr. Andrej Lebar); Tomaž ŽUŽEL z naslovom: »Modeliranje kinematike in realizacije vrtanja kvadratnih lukenj« (mentor: doc. dr. Franci Pušavec, somentor: prof. dr.

Janez Kopač); dne 30. maja 2013: Linda Barbara DROL z naslovom: »Uporaba tehnike povečevanja globinske ostrine slik pri karakterizaciji površin« (mentor: prof. dr. Peter Butala, somentor: doc. dr. Drago Bračun); Jan HOČEVAR z naslovom: »Sistem za shranjevanje in upravljanje z energijo v električnem vozilu« (mentor: prof. dr. Peter Butala, somentor: izr. prof. dr. Tomaž Katrašnik); Denis KREBELJ z naslovom: »Razvojno vrednotenje napetostno deformacijskega stanja v meglenki« (mentor: prof. dr. Marko Nagode); Jure POTOČNIK z naslovom: »Razvoj preskuševališča prirobničnih tesnil« (mentor: prof. dr.

Marko Nagode). * Na Fakulteti za strojništvo Univerze v Mariboru sta pridobila naziv univerzitetni diplomirani inženir strojništva: dne 24. maja 2013: Tomaž SKARLOVNIK z naslovom: »Prilagodljivost in optimizacija montaže kuhalnih aparatov glede na velikost serije« (mentor: doc. dr. Marjan Leber, somentor: izr. prof. dr. Borut Buchmeister); Andrej SKRT z naslovom: »Vpeljava novega produkta na trg pod blagovno znamko Gorenje« (mentor: doc. dr. Iztok Palčič).

Strojniški vestnik - Journal of Mechanical Engineering 59(2013)6, SI 75-77

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Na Fakulteti za strojništvo Univerze v Ljubljani je pridobil naziv magister inženir strojništva: dne 30. maja 2013: Sergio LOPEZ CAMPILLO z naslovom: »Model celovite učinkovitosti solarnega hlajenja / Model of overall efficiency of solar assisted cooling« (mentor: prof. dr. Sašo Medved);

Na Fakulteti za strojništvo Univerze v Mariboru je pridobil naziv univerzitetni diplomirani gospodarski inženir: dne 24. maja 2013: Sašo FLUHER z naslovom: »Uvedba proizvodnega procesa tabletiranja v podjetju Turno d.o.o.« (mentor: doc. dr. Iztok Palčič, somentor prof. dr. Duško Uršič).

* Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv diplomirani inženir strojništva: dne 10. maja 2013: Jaka BORŠTNAR z naslovom: »Termoekonomska analiza hlajenja objektov s solarno toplotno črpalko« (mentor: prof. dr. Alojz Poredoš); Urban JEZERŠEK z naslovom: »Izboljšanje merilnega postopka za merjenje neobdelanih lukenj komutatorjev« (mentor: izr. prof. dr. Ivan Bajsić); Žiga MRAZ z naslovom: »Sprememba tehnologije montaže v serijski proizvodnji elektromotorjev za avtomobilsko radarske sisteme« (mentor: prof. dr. Peter Butala); dne 13. maja 2013: Borut DEMŠAR z naslovom: »Vpenjalne glave za avtomatizirane CNC-stružnice« (mentor: prof. dr.

Janez Kopač). * Na Fakulteti za strojništvo Univerze v Ljubljani sta pridobila naziv diplomirani inženir strojništva (UN): dne 10. maja 2013: Blaž BABNIK z naslovom: »Optimiranje geometrijske oblike spoja med aluminijastim hladilnikom in termoplastičnim ohišjem LED žarometa za meglo« (mentor: izr. prof. dr. Jernej Klemenc); dne 13. maja 2013: Marko PŠENIČNIK z naslovom: »Avtomatsko sočelno varjenje cevi iz nerjavnega jekla« (mentor: prof. dr. Janez Tušek).

* Na Fakulteti za strojništvo Univerze v Mariboru sta pridobila naziv magister inženir strojništva: dne 22. maja 2013: Denis LEP z naslovom: »Dimenzioniranje nosilca hidroležaja« (mentor: prof. dr. Srečko Glodež, somentor: doc. dr. Janez Kramberger), Peter PLANINŠEK z naslovom: »Izdelava modela CNC-stružnice GF NDM16« (mentor: doc. dr.

Mirko Ficko). * Na Fakulteti za strojništvo Univerze v Mariboru je pridobil naziv diplomirani inženir strojništva (UN): dne 24. maja 2013: Gregor REŠEK z naslovom: »Dimenzioniranje nosilca izpušnega sistema motornega kolesa« (mentor: prof. dr. Srečko Glodež, somentor: doc. dr.

Janez Kramberger). * Na Fakulteti za strojništvo Univerze v Mariboru sta pridobila naziv diplomirani inženir strojništva: dne 24. maja 2013: Dražen BEKAVAC z naslovom: »Izdelava in optimizacija orodja za tlačno litje aluminijevih odlitkov« (mentorica: izr. prof. dr. Borut Buchmeister, somentor: doc. dr. Marjan Leber); Marjan MESARIČ z naslovom: »Mazanje orodij pri vročem iztiskovanju aluminijevih zlitin« (mentor: prof. dr. Ivan Anžel, somentor: doc. dr. Leo Gusel).

SI 77

Strojniški vestnik – Journal of Mechanical Engineering (SV-JME) Aim and Scope The international journal publishes original and (mini)review articles covering the concepts of materials science, mechanics, kinematics, thermodynamics, energy and environment, mechatronics and robotics, fluid mechanics, tribology, cybernetics, industrial engineering and structural analysis.

The journal follows new trends and progress proven practice in the mechanical engineering and also in the closely related sciences as are electrical, civil and process engineering, medicine, microbiology, ecology, agriculture, transport systems, aviation, and others, thus creating a unique forum for interdisciplinary or multidisciplinary dialogue.

The international conferences selected papers are welcome for publishing as a special issue of SV-JME with invited co-editor(s).

Editor in Chief Vincenc Butala University of Ljubljana Faculty of Mechanical Engineering, Slovenia Technical Editor Pika Škraba University of Ljubljana Faculty of Mechanical Engineering, Slovenia Editorial Office University of Ljubljana (UL) Faculty of Mechanical Engineering SV-JME Aškerčeva 6, SI-1000 Ljubljana, Slovenia Phone: 386-(0)1-4771 137 Fax: 386-(0)1-2518 567 E-mail: info@sv-jme.eu, http://www.sv-jme.eu Print DZS, printed in 450 copies Founders and Publishers University of Ljubljana (UL) Faculty of Mechanical Engineering, Slovenia University of Maribor (UM) Faculty of Mechanical Engineering, Slovenia Association of Mechanical Engineers of Slovenia

http://www.sv-jme.eu

59 (2013) 6

Chamber of Commerce and Industry of Slovenia Metal Processing Industry Association

Since 1955

Strojniški vestnik Journal of Mechanical Engineering

s

Journal of Mechanical Engineering - Strojniški vestnik

mož Mrvar, Jožef Medved, Janez Grum: alysis of Laser Coating Ceramic Components TiB2 and TiC y EN AW-6082-T651

no. 6 2013 volume 59

year

Cover: Evolution in programming software made robots suitable not only for pick and place operations, but for multiaxis machining as well. With reasonably appointed milling limitations robots can be used in every branch of industry. Some materials (steel, wood, foam and plastics) which were machined for research purposes in laboratories are presented in the cover.

Image Courtesy: Laboratory of Computer Aided Design (LECAD), Laboratory for cutting (LABOD), Faculty of Mechanical Engineering, University of Ljubljana

International Editorial Board Koshi Adachi, Graduate School of Engineering,Tohoku University, Japan Bikramjit Basu, Indian Institute of Technology, Kanpur, India Anton Bergant, Litostroj Power, Slovenia Franci Čuš, UM, Faculty of Mech. Engineering, Slovenia Narendra B. Dahotre, University of Tennessee, Knoxville, USA Matija Fajdiga, UL, Faculty of Mech. Engineering, Slovenia Imre Felde, Obuda University, Faculty of Informatics, Hungary Jože Flašker, UM, Faculty of Mech. Engineering, Slovenia Bernard Franković, Faculty of Engineering Rijeka, Croatia Janez Grum, UL, Faculty of Mech. Engineering, Slovenia Imre Horvath, Delft University of Technology, Netherlands Julius Kaplunov, Brunel University, West London, UK Milan Kljajin, J.J. Strossmayer University of Osijek, Croatia Janez Kopač, UL, Faculty of Mech. Engineering, Slovenia Franc Kosel, UL, Faculty of Mech. Engineering, Slovenia Thomas Lübben, University of Bremen, Germany Janez Možina, UL, Faculty of Mech. Engineering, Slovenia Miroslav Plančak, University of Novi Sad, Serbia Brian Prasad, California Institute of Technology, Pasadena, USA Bernd Sauer, University of Kaiserlautern, Germany Brane Širok, UL, Faculty of Mech. Engineering, Slovenia Leopold Škerget, UM, Faculty of Mech. Engineering, Slovenia George E. Totten, Portland State University, USA Nikos C. Tsourveloudis, Technical University of Crete, Greece Toma Udiljak, University of Zagreb, Croatia Arkady Voloshin, Lehigh University, Bethlehem, USA President of Publishing Council Jože Duhovnik UL, Faculty of Mechanical Engineering, Slovenia General information Strojniški vestnik – Journal of Mechanical Engineering is published in 11 issues per year (July and August is a double issue).

Institutional prices include print & online access: institutional subscription price and foreign subscription €100,00 (the price of a single issue is €10,00); general public subscription and student subscription €50,00 (the price of a single issue is €5,00). Prices are exclusive of tax.

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ISSN 0039-2480 © 2013 Strojniški vestnik - Journal of Mechanical Engineering. All rights reserved. SV-JME is indexed / abstracted in: SCI-Expanded, Compendex, Inspec, ProQuest-CSA, SCOPUS, TEMA. The list of the remaining bases, in which SV-JME is indexed, is available on the website.

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Strojniški vestnik - Journal of Mechanical Engineering is also available on http://www.sv-jme.eu, where you access also to papers’ supplements, such as simulations, etc.

Instructions for Authors All manuscripts must be in English. Pages should be numbered sequentially. The maximum length of contributions is 10 pages. Longer contributions will only be accepted if authors provide justification in a cover letter. Short manuscripts should be less than 4 pages. For full instructions see the Authors Guideline section on the journal’s website: http://en.sv-jme.eu/.

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THE FORMAT OF THE MANUSCRIPT The manuscript should be written in the following format: - A Title, which adequately describes the content of the manuscript.

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Meaning of symbols and units belonging to symbols should be explained in each case or quoted in a special table at the end of the manuscript before References.

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Tables should carry separate titles and must be numbered in consecutive numerical order in the text and referred to in both the text and the caption as

Table 1, Table 2, etc. In addition to the physical quantity, e.g. t (in italics), units (normal text), should be added in square brackets. The tables should each have a heading. Tables should not duplicate data found elsewhere in the manuscript.

Acknowledgement of collaboration or preparation assistance may be included before References. Please note the source of funding for the research.

REFERENCES A reference list must be included using the following information as a guide. Only cited text references are included. Each reference is referred to in the text by a number enclosed in a square bracket (i.e., [3] or [2] to [6] for more references). No reference to the author is necessary.

References must be numbered and ordered according to where they are first mentioned in the paper, not alphabetically. All references must be complete and accurate. All non-English or. non-German titles must be translated into English with the added note (in language) at the end of reference. Examples follow.

Journal Papers: Surname 1, Initials, Surname 2, Initials (year). Title. Journal, volume, number, pages, DOI code.

[1] Hackenschmidt, R., Alber-Laukant, B., Rieg, F. (2010). Simulating nonlinear materials under centrifugal forces by using intelligent crosslinked simulations. Strojniški vestnik - Journal of Mechanical Engineering, vol. 57, no. 7-8, p. 531-538, DOI:10.5545/sv-jme.2011.013.

Journal titles should not be abbreviated. Note that journal title is set in italics.

Please add DOI code when available and link it to the web site.

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[2] Groover, M.P. (2007). Fundamentals of Modern Manufacturing. John Wiley & Sons, Hoboken.

Note that the title of the book is italicized.

Chapters in Books: Surname 1, Initials, Surname 2, Initials (year). Chapter title. Editor(s) of book, book title. Publisher, place of publication, pages.

[3] Carbone, G., Ceccarelli, M. (2005). Legged robotic systems. Kordić, V., Lazinica, A., Merdan, M. (Eds.), Cutting Edge Robotics. Pro literatur Verlag, Mammendorf, p. 553-576.

Proceedings Papers: Surname 1, Initials, Surname 2, Initials (year). Paper title. Proceedings title, pages.

[4] Štefanić, N., Martinčević-Mikić, S., Tošanović, N. (2009). Applied Lean System in Process Industry. MOTSP 2009 Conference Proceedings, p.

422-427.

Standards: Standard-Code (year). Title. Organisation. Place.

[5] ISO/DIS 16000-6.2:2002. Indoor Air – Part 6: Determination of Volatile Organic Compounds in Indoor and Chamber Air by Active Sampling on TENAX TA Sorbent, Thermal Desorption and Gas Chromatography using MSD/FID. International Organization for Standardization. Geneva. www pages: Surname, Initials or Company name. Title, from http://address, date of access.

[6] Rockwell Automation. Arena, from http://www.arenasimulation.com, accessed on 2009-09-07.

EXTENDED ABSTRACT By the time the paper is accepted for publishing, the authors are requested to send the extended abstract (approx. one A4 page or 3.500 to 4.000 characters). The instructions for writing the extended abstract are published on the web page http://www.sv-jme.eu/ information-for-authors/.

COPYRIGHT Authors submitting a manuscript do so on the understanding that the work has not been published before, is not being considered for publication elsewhere and has been read and approved by all authors. The submission of the manuscript by the authors means that the authors automatically agree to transfer copyright to SV-JME and when the manuscript is accepted for publication. All accepted manuscripts must be accompanied by a Copyright Transfer Agreement, which should be sent to the editor. The work should be original by the authors and not be published elsewhere in any language without the written consent of the publisher.

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Authors can track the status of their accepted articles on http://en.svjme.eu/.

PUBLICATION FEE For all articles authors will be asked to pay a publication fee prior to the article appearing in the journal. However, this fee only needs to be paid after the article has been accepted for publishing. The fee is 300.00 EUR (for articles with maximum of 10 pages), 20.00 EUR for each addition page.

Additional costs for a color page is 90.00 EUR.

http://www.sv-jme.eu

59 (2013) 6

Strojniški vestnik Journal of Mechanical Engineering

Since 1955

Contents

Papers

358

Tomaž Petrun, Jože Flašker, Marko Kegl: A Theoretical and Numerical Study of an Additional Viscosity Term in a Modified Elasto-Plastic Friction Model for Wet Friction Clutch Simulations

367

Pingfa Feng, Chenglong Zhang, Zhijun Wu, Jianfu Zhang: Effect of Scratch Velocity on Deformation Features of C-plane Sapphire during Nanoscratching

375

Sedat Karabay, Kasım Baynal, Cengiz İğdeli: Detecting Groan Sources in Drum Brakes of Commercial Vehicles by TVA-FMEA: A Case Study

387

Peng Gao, Shaoze Yan, Liyang Xie, Jianing Wu: Dynamic Reliability Analysis of Mechanical Components Based on Equivalent Strength Degradation Paths

400

Tomaž Berlec, Janez Kušar, Lidija Rihar, Marko Starbek: Selecting the Most Adaptable Work Equipment

409

Iztok Palčič, Marc Pons, Andrea Bikfalvi, Josep Llach, Borut Buchmeister: Analysing Energy and Material Saving Technologies’ Adoption and Adopters

418

Yilmaz Kucuk: Modeling Nonlinear Viscoelastic Nanoindentation of PVAc at Different Unloading Rates

Journal of Mechanical Engineering - Strojniški vestnik

Vytautas Ostasevicius, Rimvydas Gaidys, Rolanas Dauksevicius,

Sandra Mikuckyte: 351 Study of Vibration Milling for Improving Surface Finish of Difficult-to-Cut Materials

6 year 2013 volume 59 no.