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60 (2014) 4

Strojniški vestnik Journal of Mechanical Engineering

Since 1955

Papers

213

Sergey N. Grigoriev, Victor K. Starkov, Nikolay A. Gorin, Peter Krajnik, Janez Kopac: Creep-Feed Grinding: An Overview of Kinematics, Parameters and Effects on Process Efficiency

221

Tine Seljak, Matjaž Kunaver, Tomaž Katrašnik: Emission Evaluation of Different Types of Liquefied Wood

232

Tatiana Minav, Henri Hänninen, Antti Sinkkonen, Lasse Laurila, Juha Pyrhönen: Electric or Hydraulic Energy Recovery Systems in a Reach Truck – A Comparison

241

Tomaž Bešter, Matija Fajdiga, Marko Nagode: Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters

250

Marija Blažić, Stevan Maksimović, Zlatko Petrović, Ivana Vasović, Dragana Turnić: Determination of Fatigue Crack Growth Trajectory and Residual Life under Mixed Modes

255

Jan Škofic, David Koblar, Miha Boltežar: Parametric Study of a Permanent-Magnet Stepper Motor’s Stepping Accuracy Potential

265

Tilen Thaler, Primož Potočnik, Janez Kopač, Edvard Govekar: Experimental Chatter Characterization in Metal Band Sawing

274

Marko Corn, Gregor Černe, Igor Papič, Maja Atanasijević-Kunc: Improved Integration of Renewable Energy Sources with the Participation of Active Customers

Journal of Mechanical Engineering - Strojniški vestnik

Contents

4 year 2014 volume 60 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. 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

Founding Editor Bojan Kraut

University of Ljubljana, Faculty of Mechanical Engineering, Slovenia

Editorial Office University of Ljubljana, Faculty of Mechanical Engineering SV-JME, Aškerčeva 6, SI-1000 Ljubljana, Slovenia Phone: 386 (0)1 4771 137 Fax: 386 (0)1 2518 567 info@sv-jme.eu, http://www.sv-jme.eu Print: Littera Picta, printed in 400 copies Founders and Publishers University of Ljubljana, Faculty of Mechanical Engineering, Slovenia University of Maribor, Faculty of Mechanical Engineering, Slovenia Association of Mechanical Engineers of Slovenia Chamber of Commerce and Industry of Slovenia, Metal Processing Industry Association President of Publishing Council Branko Širok University of Ljubljana, Faculty of Mechanical Engineering, Slovenia

Vice-President of Publishing Council Jože Balič University of Maribor, Faculty of Mechanical Engineering, Slovenia

Cover: The picture shows a highly-porous vitrified aluminum oxide grinding wheel designed for creep-feed grinding applications, such as grinding of turbine blades.

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 Mechanical Engineering, Slovenia Narendra B. Dahotre, University of Tennessee, Knoxville, USA Matija Fajdiga, UL, Faculty of Mechanical Engineering, Slovenia Imre Felde, Obuda University, Faculty of Informatics, Hungary Jože Flašker, UM, Faculty of Mechanical Engineering, Slovenia Bernard Franković, Faculty of Engineering Rijeka, Croatia Janez Grum, UL, Faculty of Mechanical 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 Mechanical Engineering, Slovenia Franc Kosel, UL, Faculty of Mechanical Engineering, Slovenia Thomas Lübben, University of Bremen, Germany Janez Možina, UL, Faculty of Mechanical 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 Mechanical Engineering, Slovenia Leopold Škerget, UM, Faculty of Mechanical 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 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. Delivery is included in the price. The recipient is responsible for paying any import duties or taxes. Legal title passes to the customer on dispatch by our distributor. Single issues from current and recent volumes are available at the current single-issue price. To order the journal, please complete the form on our website. For submissions, subscriptions and all other information please visit: http://en.sv-jme.eu/. You can advertise on the inner and outer side of the back cover of the magazine. The authors of the published papers are invited to send photos or pictures with short explanation for cover content.

Image Courtesy: Moscow State University of Technology “Stankin”, Russian Federation

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ISSN 0039-2480 © 2014 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 60(2014)4 Contents

Contents Strojniški vestnik - Journal of Mechanical Engineering volume 60, (2014), number 4 Ljubljana, April 2014 ISSN 0039-2480 Published monthly

Papers Sergey N. Grigoriev, Victor K. Starkov, Nikolay A. Gorin, Peter Krajnik, Janez Kopač: Creep-Feed Grinding: An Overview of Kinematics, Parameters and Effects on Process Efficiency Tine Seljak, Matjaž Kunaver, Tomaž Katrašnik: Emission Evaluation of Different Types of Liquefied Wood Tatiana Minav, Henri Hänninen, Antti Sinkkonen, Lasse Laurila, Juha Pyrhönen: Electric or Hydraulic Energy Recovery Systems in a Reach Truck– A Comparison Tomaž Bešter, Matija Fajdiga, Marko Nagode: Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters Marija Blažić, Stevan Maksimović, Zlatko Petrović, Ivana Vasović, Dragana Turnić: Determination of Fatigue Crack Growth Trajectory and Residual Life under Mixed Modes Jan Škofic, David Koblar, Miha Boltežar: Parametric Study of a Permanent-Magnet Stepper Motor’s Stepping Accuracy Potential Tilen Thaler, Primož Potočnik, Janez Kopač, Edvard Govekar: Experimental Chatter Characterization in Metal Band Sawing Marko Corn, Gregor Černe, Igor Papič, Maja Atanasijević-Kunc: Improved Integration of Renewable Energy Sources with the Participation of Active Customers

213 221 232 241 250 255 265 274


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 213-220 © 2014 Journal of Mechanical Engineering. All rights reserved. DOI:10.5545/sv-jme.2013.1547

Original Scientific Paper

Received for review: 2013-11-14 Received revised form: 2014-02-04 Accepted for publication: 2014-02-11

Creep-Feed Grinding: An Overview of Kinematics, Parameters and Effects on Process Efficiency

Grigoriev, S.N. – Starkov, V.K. – Gorin, N.A. – Krajnik, P. – Kopac, J. Sergey N. Grigoriev1 – Victor K. Starkov1* – Nikolay A. Gorin1 – Peter Krajnik2 – Janez Kopač2 1 Moscow State University of Technology “Stankin”, Russian Federation 2 University of Ljubljana, Faculty of Mechanical Engineering, Slovenia Grinding kinematics is one of the main mechanisms affecting the behaviour and efficiency of the creep-feed grinding process; it is thus essential in understanding the interplay of its parameters in the material removal. This paper presents an overview of non-traditional process parameters, such as the apparent area of the removed material, the grinding force engagement angle, the ratio of normal-to-tangential grinding force, as well as the ratio between the depth of cut and the wheel diameter. The kinematic aspects of creep-feed grinding processes are illustrated in three different case studies for creep-feed grinding of turbine blades, gears and broaches, using highly porous, vitrified, alumina-oxide wheels at low speeds. Details about the experimental work, especially with regard to analysis and validation, are not included. Based on the case studies, however, some practical guidelines for improving process efficiency in terms of productivity and quality are provided. Keywords: grinding, creep-feed, conventional, kinematics, parameters

0 INTRODUCTION High-performance grinding, particularly in the automotive and aerospace industries, requires enhanced processes that provide increased efficiency with respect to productivity, quality and costs [1]. High wheel speeds are usually employed in the pursuit of high-performance grinding, because such speeds enable utilizing larger depths of cut and hence achieve higher material removal rates, leading to increased productivity [1] without compromising quality. The different applications of grinding employing large depths of cut include (a) creep-feed grinding for extremely deep forms, and (b) high-efficiency deep grinding (HEDG) for extremely high removal rates and deep forms [2]. Creep-feed grinding has been used as early as high-speed grinding, dating back to mid-1960s. In creep-feed grinding, low workpiece speeds and large depths of cut are used. Typical creep-feed workpiece speeds are lower than 60 mm/ min [3]. In the early 1990s, however, HEDG emerged as a process that increased the efficiency of creepfeed grinding by using both higher wheel speeds and higher workpiece speeds. Tawakoli has shown that setting these two parameters at a unusually high level allows the realisation of high material removal rates and reduced grinding temperatures [4]. This paper presents an overview of three different applications of creep-feed grinding using highlyporous grinding wheels. In these case studies, the workpiece form is ground with large depths of cut, often even in a single grinding pass at the depth that can reach 10 or more millimetres, using workpiece speeds ranging from 40 to 500 mm/min, depending

on the wheel speed, type of material being ground and surface integrity requirements [5]. Even though the major increase of creep-feed grinding efficiency has been driven by HEDG and by high-performance grinding machines, it is possible to increase the efficiency of creep-feed grinding even at wheel speeds that are sometimes as low as 20 to 35 m/s [6]. This is an essential realization, since not all end-users in the industry have the possibility of utilizing HEDG due to its high costs. Creep-feed grinding requires efficient cooling, achieved by providing a useful flow of the coolant to the grinding zone in order to convect heat and to avoid thermal damage [7]. Fluid flow through the grinding zone can be enhanced by the use of porous grinding wheels. It has been shown that creep-feed grinding with porous aluminium oxide wheels can yield extremely low energy partitions (fraction of heat entering the workpiece) of only 3 to 7%, which are comparable to grinding applications with vitrified and electroplated CBN wheels [7]. This allows the material removal rate to be increased, since a highly porous grinding wheel generates less heat than a closed wheel; therefore, the amount of heat entering the workpiece is reduced. In this consideration, the highly-porous conventional grinding wheels are not obsolete and can be used for numerous creep-feed grinding applications. These wheels are also relatively inexpensive in comparison to CBN wheels and, therefore, are more cost efficient. The case studies included in this paper show satisfactory performance of these wheels with an open structure (ranging from 16 to 24). Special consideration is needed, however, because open structures means less bond, which can

*Corr. Author’s Address: Moscow State University of Technology “Stankin”, Vadkovskiy pereulok, d.1, GSP-4, Moscow, 127055, Russia, v.starkov@stankin.ru

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Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 213-220

lead to excessive wear, in overly aggressive grinding conditions [6]. The behaviour of creep-feed grinding processes is defined mostly by its kinematics, which depend on the set of selected process parameters and the wheel topography. The set parameters further determine the thermal aspects of the process (e.g. heat flux, energy partition, grinding temperature) and a number of process outputs, such as a grit contact time, grinding forces, wheel wear, etc. Creep-feed grinding differs from conventional, shallow-cut grinding, particularly with respect to thermal aspects. The mechanics of chip removal, ground surface generation, along with quality outputs, e.g. thermal damage to the workpiece surface and sub-structural layer are given in [5] and [6]. These investigations, carried-out in Russia, highlight some unique parameters of creepfeed grinding. For example, different analytical approaches were applied to explain why it is possible to grind efficiently at the two extremes of material removal rate without excessive thermal damage in the intermediate depth-of-cut range; one such approach to analyse creep-feed grinding has been to characterize the angle of inclination of the contact plane [8]. Nevertheless, there is a need to introduce additional parameters of creep-feed grinding, not only for process analysis but also to identify the levers to increase process efficiency. More than 50 years have passed since the development of the first kinematic models of grinding processes; there are now numerous basic models available for process analysis, e.g. grinding wheel topography, chip thickness, grinding forces, energy, and temperature, etc. [9] and [10]. Interestingly, a different set of basic models evolved in Russian grinding research, resulting in parameters that are practically unknown internationally. This unknown theoretical base, suitable for creep-feed grinding analysis, is hence described in this paper, along with straightforward interpretations to aid in the improving of process efficiency. More specifically, the paper introduces parameters, including (a) the apparent area of removed material, (b) the grinding force engagement angle, (c) the ratio of normal to tangential grinding force, and (d) the ratio between the depth of cut and wheel diameter, which are useful in setting-up the grinding system. For example, the combination of wheel diameter and depth of cut can be optimised in order to minimize the ratio of normal to tangential grinding force. Furthermore, different interpretations of kinematics and guidelines for increasing process efficiency are given in order to increase the usefulness of the presented work, especially with respect to practicing engineers. 214

Three case studies are given to illustrate the ranges of process parameters for different applications. One case study is in the gas turbine industry: grinding a shank and an attachment section of a turbine blade. The other two examples refer to gear grinding and the grinding of broaching tools. Note that the ambition of this paper is not to experimentally verify or analyse any of these applications in any particular detail, but to revisit the grinding kinematics in view of largely unknown parameters and to simulate the effects on grinding efficiency by different selections of these parameters. 1 KINEMATICS OF CREEP-FEED GRINDING In grinding, the kinematic relations between the wheel and the workpiece are typically analysed on the abrasive grit-cutting edge scale. The first analytical investigations of this type were made by Peklenik [11], who suggested that a limited number of kinematic cutting edges are actively engaged in material removal. Since then, numerous analyses of grinding kinematics, arising from consideration of the kinematic interactions between the grit and the workpiece, have been translated to the basic models of grinding wheel topography and chip thickness [9] and [10]. These basic models include parameters of grinding kinematics, such as the speed ratio (ratio between the wheel and the workpiece speed), as well as parameters of grinding geometry (e.g. depth of cut, equivalent wheel diameter) and parameters to quantify the topography (e.g. static density of the cutting edges) of the wheel surface [1]. However, kinematics can also be modelled on the macro scale [12].

vs O1

s Fn

B1 ae

O

ΡP

B

ds

Ft lg F

S A1

A

vw

Fig. 1. Illustration of surface (straight) grinding

For surface (straight) grinding operation, shown in Fig. 1, commonly used for creep-feed grinding, the

Grigoriev, S.N. – Starkov, V.K. – Gorin, N.A. – Krajnik, P. – Kopac, J.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 213-220

three main process parameters are (1) the wheel speed vs, (2) the workpiece speed vw, and (3) the depth of cut ae. The illustrated geometry corresponds to an up-grinding type of operation, in which the tangential directions of the wheel and workpiece motion are opposite. A cutting point in up-grinding begins its contact with the workpiece between Points A and A1 and ends at B1. The previous cutting point follows the same geometrical path shape but is displaced along the workpiece surface by the distance BB1, that corresponds to the feed-per-cutting point s, which is equal to the product of the workpiece speed and the time between successive cuts [13]. Penetration of the grinding wheel into the workpiece results in the apparent area of removed material S, which can be expressed as:

S = lg ⋅ s =

vw ⋅ ae ⋅ d s . (1) vs

Here, ds is the wheel diameter and lg the arc length AB = A1B1 of the cutting path as the wheel centre moves from O to O1, and the grit passes through the contact zone. The area S is inversely proportional to the speed ratio vs/vw and is increased with the rise in both the wheel diameter ds and the depth of cut ae. The geometric contact length, lg, disregarding the contribution of feed per cutting edge, can be expressed as [2]:

lg = ae⋅ d s . (2)

The grit contact time τ with the workpiece within the contact length (the contact time experienced by a grit during which Point A of the grinding wheel moves to Point B1) is given by the geometric contact length lg divided by the wheel speed vs:

τ=

lg vs

. (3)

Simultaneously, within the same contact time, the grit moves horizontally, as governed by the feed-percutting point s, limiting the apparent area of material removal by BB1 = AA1, whereas the geometric contact length limits the same area by AB = A1B1. Note that the wheel and grit deflection effects are not taken into account. The total force vector F generally increases with the apparent area of removed material S and, among others, depends on grinding geometry and kinematics. The other influencing factors on the magnitude of F are: (1) mechanical properties of workpiece material

(e.g. strength and hardness); (2) chip thickness; (3) wear-flat area (due to wear and dressing); and (4) contact conditions between the grit and the workpiece (i.e. contact stress and friction coefficient). The direction of vector F is determined by the grinding force engagement angle αP, calculated as:

αP =

lg

a 4 ⋅ 90° = 57.32 ⋅ e . (4) ds 2 π ds ⋅

Typical αP values for creep-feed grinding lie in the range between 1.8 to 12 degrees [6]. An additional parameter of the grinding kinematics is the cotangent of the grinding force engagement angle αP, which is the ratio of normal grinding force Fn to tangential grinding force Ft:

Fn = ctgα P . (5) Ft

Note that this parameter is the inverse of the more commonly used grinding force ratio, indicating the relationship of tangential force relative to the normal force [2]. Weal wear is not considered in the presented analysis of grinding kinematics; however, it should be mentioned that as the wheel wear progresses, the tangential force increases slightly, but the normal force increase is more drastic. 2 INTERPRETATION OF CREEP-FEED GRINDING KINEMATICS Grit contact time as a grain passes through the contact length τ is an parameter in process analysis, particularly in consideration of grinding temperatures [2]. Typical τ values for grinding processes range between 10-3 to 10-5 seconds (larger values for creepfeed grinding). In consideration of the fact that the time τ is remarkably short, we can treat the determined kinematic parameters as instantaneous, which should be applied for the analysis of the process. As mentioned earlier, it is possible to determine the number of kinematic cutting edges in contact with the workpiece per unit area [11]. The number of active cutting edges involved in material removal is proportional to the number of kinematic cutting edges. In conventional shallow-cut grinding with an 80-grit 12-structure aluminium oxide wheel (diameter ds=500 mm), approximately 3.3 abrasive grits can theoretically be in contact with the workpiece over the contact length of 1 mm during the grit contact time of τ [6]. In contrast, for a creep-feed grinding scenario using the same wheel and cutting depths in the range between 0.5 to 10 mm, a much larger number of grits,

Creep-Feed Grinding: An Overview of Kinematics, Parameters and Effects on Process Efficiency

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i.e. 50 to 200, could be actively engaged in material removal within the contact length between 15.8 to 70.7 mm [6]. In general, the number of kinematic cutting edges increases with the depth of cut. Moreover, the number of kinematic cutting edges increases with the workpiece speed and decreases with both the wheel speed and the wheel diameter [14]. Considering only the larger depths of cut (characteristic for creep-feed grinding), the associated higher number of kinematic cutting edges leads to a higher concentration of active cutting edges lying close to each other, which can cause high temperatures in the grinding zone and, hence, a risk of thermal damage in case a temperature equilibrium is established. However, this is not the case in creep-feed grinding. According to Malkin and Guo [5], creep-feed grinding with relatively fast workpiece speeds is characterized by the inclined heat source (tending in a direction from A to B1 in Fig. 1) associated with the large depth of cut. This leads to a situation where some heated material in the wedge ahead of the grinding zone is removed with the chips during the grinding process, thereby resulting in lower temperatures on the finished ground surface (area AA1 in Fig. 1). Similar observations have been made by others, who investigated the phenomena experimentally [4], [8], [14] and [15]. The grinding force engagement angle αP (Eq. (4)) increases with the depth of cut (while in contrast the Fn/Ft value declines), and decreases when utilizing grinding wheels with larger diameters. The latter interrelation is somewhat ambiguous, because grinding with a large wheel diameter generally yields large contact lengths. Therefore, it is necessary to investigate the ae/ds ratio, i.e. the ratio between the depth of cut and the wheel diameter, in greater detail. An increase in the ae/ds ratio causes the vector F to slightly move from the normal OA, towards the machined surface, to line BB1 parallel to it (straight surface of the workpiece). For example, if the depth of cut increases from ae = 1 to 10 mm, the grinding force engagement angle increases from αP = 2.56 to 8.11 degrees, or by the factor of 3.17. Simultaneously, the ratio between the normal and tangential grinding force, Fn/Ft, is decreased from 22.3 to 7. Note, however, that the Fn/Ft ratio also depends on the dressing conditions, the wheel sharpness and the penetration depth of the grit, and not solely on the grinding geometry. For example, a dull wheel would have a larger Fn/Ft ratio without any change in the grinding geometry. Nevertheless, these factors are not considered in the interpretation of grinding kinematics in focus here. 216

Additional interpretations can be made into grinding kinematics. The illustrated wheel engagement in creep-feed grinding, shown in Fig. 1, suggests that the limit of theoretical depth of cut is ae=ds/2. Correspondingly, the upper limits for the grinding force engagement angle αP lie between 40.53 to 57.32 degrees in case of ae/ds=0.5. Running the process at such extremes would be advantageous from the viewpoint of grinding kinematics and attainable material removal rates, but impossible to implement in practice due to mechanical limitations in machine tools and grinding wheels. In more realistic, commonly applied grinding scenarios, the ae/ds ratios for conventional shallow-cut grinding are between 10-6 and 10-4; ratios corresponding to the grinding force engagement angle αP in the range between 0.1 to 1.2 degrees. The applications of creep-feed grinding using conventional wheels typically employ ae/ds ratios of 10-3 to 10-2, and the grinding force engagement angle αP in the 1.8 to 12 degrees range. These typical values of ae/ds and αP can be used as criteria for determining whether the application refers to a creep-feed grinding operation. Finally, it should be noted that the ae/ds ratio is not related to wheel and workpiece speeds or, therefore, to the machining time. The ratio only refers to the grinding force engagement angle αP and the ratio of normal to tangential grinding force Fn/Ft. 3 ILLUSTRATIVE CASE STUDIES The kinematics of creep-feed grinding is illustrated in three different case studies related to grinding of (1) turbine blades, (2) gears and (3) broaches. The ranges of kinematic parameters used for comparison are presented in Table 1. Table 1. Ranges of kinematic parameters [6] Process (1) (2) (3) ds [mm] 500 300 100 ae [mm] 0.5 to 10 0.1 to 6 0.002 to 0.2 vw [mm/min] 90 to 400 200 to 2000 300 to 10000 vs [m/s] 28 to 30 35 25 Q’w [mm3/(mm·s)] 3.3 to 14.9 2.5 to 20 0.33 to 1 ae/ds 10-3 to 2∙10-2 3∙10-4 to 2∙10-2 2∙10-5 to 2∙10-3 S [mm2] 0.034 to 0.15 0.021 to 0.206 0.0033 to 0.011 αp [°] 1.81 to 8.11 1.05 to 8.11 0.26 to 2.56 Fn/Ft 7.01 to 31.53 7.01 to 54.56 23.31 to 214.9 aggr 7 to 30 10 to 25 9 to 30

Considering that the introduced parameters are neither widely known nor widely used in the analysis of grinding kinematics, it is useful to make a

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comparison to a more standard parameter used in the industry today, in particularly to the aggressiveness number, aggr, introduced by Badger [15]. This nondimensional parameter is a simplification of the maximum chip thickness model and can be calculated as: 1/ 2

 v  a  aggr = 10  w  e  . (6)  v  d  6

s

e

3.1 Creep-Feed Grinding of Turbine Blades Turbine blades made of nickel-base alloys can be ground using conventional wheels. The use of CBN grinding wheels, which have high thermal conductivity, is generally the preferred approach to grinding this type of material; nevertheless, the expense of grinding is high in such cases. Therefore, a conventional aluminium-oxide wheel, e.g. vitrified 80-grit, F-grade, 16-structure, 25A wheel (diameter 500 mm, width 25 mm) can be used for profiling of turbine blades. Wheels of this type are soft with an extremely open structure (high-porosity) that reduce the tendency for dulling. The grinding of the blade shank and the attachment section was done simultaneously using a profiled wheel, as shown in Fig. 2. Additional contours of the blade, the Z-profile, platforms, and blade roots were also ground [6].

normal and tangential grinding force is decreased from 31.53 to 7.01. Due to large depths of cut, ranging from 0.5 to 10 mm, the geometric contact lengths are also great, enabling enhanced cooling at the grinding zone; effective cooling is crucial for creep-feed grinding with highly porous wheels in order to avoid thermal damage. A 3.5% concentration emulsion was applied for cooling with a pressure of 12 bar and a flow rate of 200 l/min. Material investigations (not given here) showed an undamaged surface layer, with no noticeable metallographic changes in the microstructure, and with compressive residual stresses on the ground surface [6]. The improved efficiency refers to a reduction of the number of grinding passes that leads to fewer cycles of heating. This, combined with lower grinding temperatures, are the main advantages of creep-feed grinding in terms of thermal damage. 3.2 Creep-Feed Grinding of Gears Another example of creep-feed grinding refers to profiling of gear tooth flanks, shown in Fig. 3. Here, the grinding wheel machines a single flank in the direction of grinding per tooth gap. This process allows grinding of different moduli with an unchanged wheel width.

Fig. 2. Grinding of turbine blades

The apparent area of removed material S depends on the depth of cut, and ranges from 0.068 to 0.293 mm² when the depth of cut is increased by a factor of 20. The ae=10 mm depth of cut gives a specific material removal rate of Q’w=14.9 mm³/(mm·s), which is large for conventional wheels. The grinding force engagement angle αP increases proportionally with the depth of cut and ranges between 1.81 to 8.11 degrees. Simultaneously, the Fn/Ft ratio between the

Fig. 3. Profiling of a gear

Typical gears to be ground have moduli of 1 to 6 mm, 5 to 130 teeth, widths between 5 to 70 mm, and diameters of 50 to 600 mm. Grinding of gears normally requires several grinding passes. The maximal depth of cut for the first grinding pass depends on the gear module. For example, creep-feed grinding of a gear with a module between 3 to 5 mm

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enables employment of depths of cut ranging from 6 to 10 mm when using 300 to 350 mm wheels [6]. In terms of productivity, conventional grinding of gears yields Q’w between 2.5 to 20 mm³/(mm·s), which is associated with 0.017 to 0.033 ae/ds ratios. For this application, a typical wheel could be an 80-grit, G-grade, 14-structure, 25A aluminium oxide wheel (diameter 300 mm, width 20 mm). Again, this is a soft conventional wheel with very open structure, offering better grinding economics in comparison to CBN [6]. For further illustrating an industrial use of this process, two different process variants are discussed in [7]: (1) grinding with two passes, with 6 mm depth of cut for the first pass and 0.75 mm for the second pass; (2) grinding with five passes ranging from the maximum depth of cut of 2.4 mm for the first pass down to 0.1 mm in the last pass. The tested depths of cut were hence varied between 0.1 to 6 mm as further detailed in Table 1. In this case, we have the apparent area of removed material S ranging between 0.021 to 0.172 mm². The grinding force engagement angle αP increases simultaneously from 1.05 to 8.11 degrees. The first grinding variant with two grinding passes is more efficient than the second variant comprised of five grinding passes. Moreover, the overall decrease in the ratio between the normal and tangential grinding force Fn/Ft from 54.56 to 7.01 gave a sharp cut resulting in sufficient grinding quality. 3.3 Creep-Feed Grinding of Broaches The last case study of creep-feed grinding is a broaching tool application, shown in Fig. 4.

Fig. 4. Grinding of a broaching tool

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Broaching is used to produce internal slots, such as dovetails and turbine disk roots using a large broaching machine. Turbine disk broaches are large and can contain hundreds of cutting edges to produce the slots. The high cost and maintenance of such tools makes grinding appropriate not only for producing the tools but also the remanufacturing (retrofitting) of them. In a conventional broach grinding scenario, for example, a profiled, vitrified 80-grit, H-grade, 12-structure, 25A aluminium oxide wheel (diameter 100 mm, width 30 mm) can be used. This wheel is not as soft and opened as the other two wheels exemplified in this section. Nevertheless, the wheel still features an open structure and is soft; characteristics that are required for grinding of hardened steel (e.g. hardness of broaches is between 61 to 68 HRC). Grinding depths employed in this case study are not as large as in the previous examples. The depths range from 0.002 to 0.2 mm, giving a specific material removal rate up to Q’w = 1.67 mm³/(mm·s). This process uses a small wheel diameter and is considered as creep-feed grinding operation because the ae/ds ratio is greater than 10-3 to 10-2, according to Starkov [6]. The grinding force engagement angle αP increases proportionally with the grinding depth from 0.26 to 2.56 degrees. At the same time, the Fn/Ft ratio is decreased non-proportionally, from 214.9 to 23.31. 3.4 Guidelines for Improving Creep-Feed Grinding Process Efficiency The high temperatures in grinding can cause various types of thermal damage, affecting the integrity of a ground surface. The threshold for the onset of thermal damage largely depends on the material being ground. For example, high temperature nickelbase alloy tolerates grinding temperatures up to 1200 °C. For grinding of hardened steels, however, the temperature reached at the grinding zone should not exceed 723 °C (critical temperature of steel) to avoid detrimental metallurgical transformations. As a practical matter, it is therefore desirable to be able to quantify grinding temperature associated with metallurgical transformations that occur during creepfeed grinding. For this, temperature measurements by a thermocouple using a semi-dynamic method were carried out [6]. It was found experimentally that the increase of grinding depth makes the grinding temperature to rise sharply, and then, at the depth of 0.1 to 0.15 mm, the temperature rises more slowly, and, finally, becomes stable at the depth of 0.2 mm [16]. In this case, the grinding temperature remained

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constant during an increase of the depth of cut, i.e. from 0.2 to 0.5 mm, where the ae/ds ratio value equalled 6.7∙10-4. The temperature reached at the grinding zone became stable, meaning that higher ae/ ds ratios created advantageous conditions for both material removal and thermal aspects of the process [6]. Based on this reference example, it is possible to assume that the major guideline to enhance grinding efficiency in terms of avoiding thermal damage is to change the ae/ds ratio (for a given process kinematics and wheel topography). When this ratio is enlarged on account of increasing the depth of cut, ae, and, if possible, decreasing the grinding wheel diameter ds, suitable preconditions are created for grinding with higher grinding force engagement angles αp. In this scenario, the vector of applied cutting force shifts in the direction of removed material and more heat is evacuated from the grinding zone with chips. Fig. 5 shows the variation of the three discussed kinematic parameters when the depth of cut is increased from 0.5 to 10 mm, which corresponds to the ae/ds ratio in the range between 0.001 and 0.02.

decreases with larger wheel diameters. A compromise can be achieved by simultaneously minimizing ae/ ds and ds∙ae values. In practice, wheels with smaller diameters should be used while keeping the ae/ds ratio fixed. In this way, the creep-feed grinding efficiency can be improved, so that the benefits (e.g. low energy partitions [17] and avoiding of thermal damage) are achieved with lower depths of cut. The effect of wheel diameter ds (for different depths of cut ae) on grinding force engagement angle αp is shown in Fig. 6.

Fig. 6. Effect of wheel diameter ds on αp

Fig. 5. Variation of kinematic parameters with the depth of cut

The variations shown are characteristic for the grinding of turbine blades using a wheel with a ds = 500 mm diameter and an ae = 10 mm depth of cut, and the grinding of gears using a wheel with a diameter of ds = 300 mm and a depth of cut ae = 6 mm. Here, the ae/ds ratio equals 0.02 for both cases, meaning that the processes have the same αp and Fn/Ft values. Nevertheless, during the creep-feed grinding of blades, the apparent area of removed material S is 1.7 times higher in comparison with creep-feed grinding of gears, even though the workpiece speed vw was 2.2 times lower (90 vs. 200 mm/min). The apparent area of removed material is proportional to workpiece speed vw, wheel diameter ds and depth of cut ae, whereas the grinding force engagement angle αp

As mentioned earlier, characteristic αP values for creep-feed grinding are between 1.8 to 12 degrees. Based on the figure above, it is apparent that creepfeed grinding conditions are achieved earlier using smaller wheels. More specifically, the grinding wheel with ds = 100 mm diameter, achieves kinematic conditions for creep-feed grinding at the cutting depth of ae = 0.0986 mm. Similarly, grinding with a ds = 300 mm wheel, requires a depth of cut of ae = 0.296 mm, while during grinding using the wheel with ds = 500 mm, creep-feed grinding conditions are reached at ae = 0.493 mm. The workpiece speed vw is another key parameter for improving process efficiency, because it significantly affects the material removal rate (productivity) and the heat transfer into the workpiece, which are quantities reliant on the apparent area of removed material S. At the same time, higher workpiece speeds aid in more heat remaining in the path of the advancing S with less time for it to be conducted into the workpiece, which in turn leads to a reduction of the grinding temperature. In so doing, it is practical to use higher workpiece speeds in both creep-feed grinding and shallow-cut grinding with lower depths of cut.

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4 CONCLUSIONS This paper overviews the non-traditional parameters (reflecting different understanding of the functional potentials of basic grinding models) for characterizing the kinematics of creep-feed grinding and investigating their effects on process efficiency. Kinematics has been derived and analysed using parameters, such as apparent area of removed material S, grinding force engagement angle aP, ratio of normal to tangential grinding force Fn/Ft, as well as the ratio between the depth of cut and the wheel diameter ae/ ds. These parameters evolved over many years in Russian grinding research and are largely unknown internationally. The useful ranges of parameters are presented in illustrative examples of creep-feed grinding of turbine blades, gears and broaches. These case studies suggest that process efficiency can be increased when using highly-porous aluminium-oxide wheels run at low speeds. The benefits of creep-feed grinding, such as increased material removal rate and reduced risk of thermal damage (grinding burn), can be achieved, e.g. once the values of grinding force engagement angle are in the 1.8 to 12 degrees range. It has been shown that the work-piece speed is the key parameter for improving process efficiency. Furthermore, when grinding with a particular wheel diameter, the depth of cut should increase in order to maximize the grinding force engagement angle while minimizing the ratio of normal to tangential grinding force. The interpretation of grinding kinematics suggests that grinding with smaller wheel diameters is beneficial, because creep-feed grinding benefits can be achieved at lower depths of cut. The introduced parameters can thus be used to select the proper combination of wheel diameters and grinding parameters for different process applications. 5 REFERENCES [1] Kopac, J., Krajnik, P. (2006). High-performance grinding - a review. Journal of Materials Processing Technology, vol. 175, p. 278-284, DOI:10.1016/j. jmatprotec.2005.04.010. [2] Rowe, W.B. (2009). Principles of Modern Grinding Technology. William Andrew Publishing, Norwich. [3] Andrew, C., Howes, T.D., Pearce, T.R. (1985). Creep Feed Grinding. Holt, Rinehart & Winston, Eastbourne. [4] Tawakoli, T. (1993). High Efficiency Deep Grinding. VDI-Verlag and Mechanical Engineering Publications, London. [5] Malkin, S., Guo, C. (2007). Thermal analysis of grinding. Annals of the CIRP - Manufacturing

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Technology, vol. 56, no. 2, p. 760-782, DOI:10.1016/j. cirp.2007.10.005. [6] Starkov, V.K. (2007). Shlifovanie vysokoporistymi krugami [Grinding with highly porous grinding wheels]. Mashinostroenie, Moscow. (in Russian) [7] Starkov, V.K., Ryabtsev, S.A., Gorin, N.A. (2012). Povyshenie effektivnosti protsessov glubinnogo shlifovaniya [Efficiency improvement in deep grinding processes]. Moscow State University of Technology “Stankin”, Moscow. (in Russian) [8] Peklenik, J. (1957). Ermittlung von geometrischen und physikalischen Kenngrößen für die Grundlagenforschung des Schleifens [Determination of geometrical and physical parameters for basic grinding research]. Doctoral Thesis, TH Aachen, Aachen. [9] Rowe, W.B. (2001). Thermal analysis of high efficiency deep grinding. International Journal of Machine Tools and Manufacture, vol. 41, no. 1, p. 1-19, DOI:10.1016/ S0890-6955(00)00074-2. [10] Tönshoff, H.K., Peters, J., Inasaki, I., Paul, T. (1992). Modelling and simulation of grinding processes. Annals of the CIRP – Manufacturing Technology, vol. 41, no. 2, p. 677-688, DOI:10.1016/S0007-8506(07)63254-5. [11] Brinksmeier, E., Aurich, J.C., Govekar, E., Heinzel, C., Hoffmeister, H.-W., Klocke, F., Peters, J., Rentsch, R., Stephenson, D. J., Uhlmann, E., Weinert, K., Wittmann, M. (2006). Advances in modeling and simulation of grinding processes. Annals of the CIRP - Manufacturing Technology, vol. 55, no. 2, p. 667-696, DOI:10.1016/j. cirp.2006.10.003. [12] Drazumeric, R., Krajnik, P., Vrabic, R., Meyer, B., Butala, P., Kosel, F., Kopac, J. (2010). Modelling of grinding gap macro geometry and workpiece kinematics in throughfeed centreless grinding. Journal of Materials Processing Technology, vol. 210, no. 1, p. 104-109, DOI:10.1016/j.jmatprotec.2009.08.006. [13] Malkin, S., Guo, C. (2008). Grinding Technology: Theory and Applications of Machining with Abrasives, 2nd ed. Industrial Press, New York. [14] Klocke, F., König, W. (2005). Fertigungsverfahren 2: Schleifen, Honen, Läppen [Manufacturing Processes 2: Grinding, Honing, Lapping]. 4th ed. Springer Verlag, Berlin, Heidelberg. [15] Badger, J. (2009). Factors affecting wheel collapse in grinding. Annals of the CIRP – Manufacturing Technology, vol. 58, no. 1, p. 307-310, DOI:10.1016/j. cirp.2009.03.048. [16] Silin S.S., Hrulkov, V.A., Lobanov, A.V., Rykunov, N.S. (1984). Glubinnoe shlifovanie detalei iz trudnoobrabatyvaemykh materialov [Deep grinding of difficult-to-machine materials]. Mashinostroenie, Moscow. (in Russian) [17] Jin, T., Stephenson, D.J. (2003). Investigation of the heat partitioning in high efficiency deep grinding. International Journal of Machine Tools and Manufacture, vol. 43, no. 11, p. 1129-1134, DOI:10.1016/S0890-6955(03)00123-8.

Grigoriev, S.N. – Starkov, V.K. – Gorin, N.A. – Krajnik, P. – Kopac, J.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 221-231 © 2014 Journal of Mechanical Engineering. All rights reserved. DOI:10.5545/sv-jme.2013.1242

Original Scientific Paper

Received for review: 2013-05-26 Received revised form: 2013-11-30 Accepted for publication: 2013-12-20

Emission Evaluation of Different Types of Liquefied Wood Seljak, T. – Kunaver, M. – Katrašnik, T. Tine Seljak1,2,* – Matjaž Kunaver1 – Tomaž Katrašnik2 1 Centre of Excellence Polimat, Slovenia 2 University of Ljubljana Faculty of Mechanical Engineering, Slovenia

After initial studies, further research work on the combustion properties of second generation biofuels, obtained through solvolysis in polyhydroxy alcohols is oriented towards different types of liquefied wood that exhibit several favorable properties. In this study, different types were obtained by altering the reactant ratios of the fuels. These were focused on increased wood content and elevated pH value that would increase the techno-economic attractiveness of the fuel. Three different types of fuels were tested in a laboratory scale gas turbine, and evaluated through CO, THC and NOx emissions measurements, while varying multiple operating parameters. To achieve sufficient atomization quality, the high viscosity of the fuels was reduced by preheating to 100 °C. To speed up the droplet evaporation process and additionally to resemble conditions present in commercially available systems, high temperatures of primary air were employed by the use of exhaust gas heat regenerator. CO and THC emissions were found to be highly dependent on wood content and turbine inlet temperature, whereas with partial neutralization of the fuel this dependency was less pronounced and only NOx concentrations were influenced by altered elemental composition of the fuel. Results indicate it is possible to maintain successful combustion in microturbines even with fuels that exhibit higher pH value and reduced reactivity and with fuels containing higher amounts of lignocellulosic biomass. Keywords: biomass, fuel, gas turbine, emissions, waste to energy, solvolysis

0 INTRODUCTION Increasing interest in biofuels and fuels produced from waste materials is mainly driven by the goal to replace fossil fuels in specific applications. Presently, ethanol and biodiesel have the largest share as they closely resemble physical and chemical properties of the petroleum distillates and thus require only minor modifications of the existing spark ignition (SI) and compression ignition (CI) engine technology, whereas some engines are already suitable for direct use of these fuels as shown in [1]. Production of biodiesel is based mainly on the rapeseed and soybean feedstock, and ethanol is mainly obtained through fermentation of sugar cane and corn starch and thus both compete with food feedstock [2]. Therefore primary materials for the second generation biofuels are frequently selected from a large pool of lignocellulosic feedstock. Although lignocellulosic materials can be used in external combustion systems for small CHP applications [3], different processes are available for depolymerization of relatively complex molecules of lignin, cellulose and hemicelluloses that constitute lignocellulosic materials. Through these processes lignocellulosic materials are converted to gaseous or liquid fuels and are applicable directly in internal combustion engines. Besides production of cellulosic ethanol which features low efficiency as mainly cellulose is converted to the fuel [4], several thermochemical procedures can also be utilized to decompose lignin molecules. Although thermochemical procedures generally yield

heavier biofuels, they are still promising as some of the products can be directly used in low and medium speed diesel engines and gas turbines [5]. Currently the main research focus is on pyrolysis oils (liquid fraction of pyrolysis process) which are extensively reviewed in [6], and possibilities for their exploitation are presented in [5]. Almost any polymer material can be used for production of pyrolysis oils, although tires and biomass are the most widely used feedstocks [7]. The physical and chemical properties of pyrolysis oils are dependent on the feedstock composition and the type of the pyrolysis process. Generally, lignocellulosic waste of different origins yields oils with high viscosity (15 to 35 cSt at 40 °C) and low calorific value (~20 MJ/kg) [5]. Severe reaction conditions required for pyrolysis with temperatures exceeding 500 °C [6] usually dictate the choice of expensive materials and advanced process equipment. Depolymerization of biomass constituents can also be realized through liquefaction in different media, e.g. water and different alcohols. Required process temperatures and pressures with water are in the vicinity of critical point (280 to 370 °C and 100 to 250 bar) [8]. The advantage of direct liquefaction becomes apparent with the use of the phenol or multifunctional alcohols, since process conditions are significantly less severe in this case. This process could therefore be viable on small and medium scale where the initial investment cost is often a key factor for profitability. When converting biomass, both processes, i.e. pyrolysis and liquefaction in multifunctional alcohols,

*Corr. Author’s Address: Centre of Excellence PoliMat, Tehnološki park 24, SI-1000 Ljubljana, Slovenia, tine.seljak@polimat.si

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yield fuels with low pH value and high viscosity. In pyrolysis oils, low pH value is a consequence of high acid content; up to 10% hydroxyacetaldehyde, roughly 5% acetic acid and around 3% formic acid. The consequent pH value is usually between 2 and 3. In products of direct liquefaction in multifunctional alcohols, high acidity arises no only from acidic decomposition products as in pyrolysis oils but also from the use of acids as catalysts (H2SO4, TsOH). Besides acidity, these fuels exhibit relatively high viscosity which arises from relatively large molecules obtained through depolymerization of lignin and cellulose. For successful use in combustion applications, preheating is usually required to reduce the viscosity which is required to achieve sufficient atomization quality. Conditions during preheating combine high temperature and low pH value; this presents highly corrosive environment requiring careful selection of materials used in fuel and injection system to avoid failures due to corrosion of components. Liquefied wood (LW) was suggested as a fuel in [9], although the presented process featured an additional upgrading step with catalytic hydrotreatement to reduce the viscosity and elevate the heating value. Although, studies in [9] and [10] indicated the use of hydrotreated LW as a fuel, no data on combustion of such an upgraded type of LW was found during extensive literature survey. In contrast, authors of this paper presented primary results on combustion of non-upgraded LW in [11], which indicates that catalytic hydrotreatement of the LW as proposed in [9] and [10] is not always necessary. Instead, viscosity can be lowered by altering reactant ratios as presented in [12]. In addition, this measure can be succeeded by preheating of the fuel and thus viscosity can be sufficiently reduced to reach acceptable CO and THC emissions. Following a successful demonstration of combustion capabilities of the LW in [11] this study focuses on analyses of combustion and emission formation phenomena of different types of LW. This research is motivated by the fact that efficient exploration of different LW types extends the knowledge base on the influencing and governing mechanisms of combustion and emission formation of different fuel types and significantly contributes to the successful application of this innovative biofuel. Low price and availability of biomass is the main driver for increasing biomass content to boost the economic attractiveness of one LW type. Additionally, the LW type with more neutral pH value will be tested on its combustion characteristics. The 222

main motive for this lies in reduced reactivity of the partially neutralized LW exhibiting increased storage stability and stability at high temperatures. This is the consequence of acid catalyst neutralization which then shifts the temperature threshold of liquefaction reactions higher and prevents solid residue formation during preheating phase [13]. Preliminary indications exist that elevated pH value could also reduce the corrosivity of such fuel and in this way reduce the process equipment and fuel system material costs. The influence of altered chemical composition and product reactivity on combustion performance and exhaust emissions will be evaluated through emission measurements. Experiments are performed in an experimental gas turbine, designed exclusively for combustion analyses of heavy fuels at different operating regimes. 1 MATERIALS In the results section, baseline measurements were performed by firing diesel fuel, compliant with the EN 590 standard. Fuel production processes of the three types of the LW are described below. 1.1 Fuel Processing LW was produced through aforementioned solvolysis of spruce wood in polyhydroxy alcohols. Although solvolysis in various media is a well-known process, its subtype, glycolysis, is still a relatively unknown procedure for fuel production. Liquefaction of isolated cellulose and lignin in multifunctional alcohols was already studied in [14] to [16], whereas studies investigating liquefaction of naturally combined lignin and cellulose in woody biomass were conducted in [9], [10] and [17]. In all of the above cases, lignocellulosic material was added to the glycols mixture and heated for sufficient time in temperature interval between 120 to 250 °C. The reaction rates were increased substantially by the use of acid catalysts. Details of the liquefaction process are provided in Table 1. Efficiency of conversion of initial material (wood) to liquid products depends mainly on: Acidity: A high concentration of acid catalyst results in increased rate of recondensation reactions after the highest degree of depolymerization is already achieved. If the reaction conditions are maintained, products are prone to formation of the solid residue which influences both the efficiency of conversion and process complexity due to filtering demands. Formation of residue occurs as a consequence of recombination of liquefaction products [13]. On

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the other hand, too low acid concentration needs to be compensated by higher process temperature and prolonged liquefaction time [17]. The amount of acid catalyst should therefore be carefully determined as it plays a crucial influence on liquefaction efficiency.

Spruce wood with particle size up to 3 mm was added to the solvent. The mixture was heated to 180 °C and constantly mixed for 180 minutes. The analyzed LW compositions are explained in the following sections.

Table 1. Reaction parameters for fuel production

1.1.1 LW Type no. 1

Process Temperature Pressure Catalyst Residence time Feedstock Solvent

Solvolysis in acidified polyhydroxy alcohols 180 °C Atmospheric TsOH (tosylic acid CH3C6H4SO3H) 180 min, mixing Spruce wood flour Glycerol, di-ethylene glycol

Biomass to solvent ratio: The efficiency of conversion significantly decreases if the biomass to solvent ratio exceeds 1. At ratios below 1, the biomass content only has a minor influence on conversion efficiency. Furthermore, higher biomass contents tend to generate products with higher viscosities which could be a limiting factor for successful utilization in heat engines. Liquefaction time: Short reaction times result in low biomass to liquid conversion efficiency. The reason for this lies in only partial depolymerization of the lignin and the cellulose. Contrarily, when reaction time is prolonged beyond the optimal point, solid residue starts to form as a consequence of recondensation reactions taking place between degradation products of cellulose and lignin, as mentioned above. This clearly indicates the importance of accurately defined reaction time in combination with upper two parameters. Considering the above limitations, three different types of LW (presented also in Fig. 2) were produced in a 200 L batch reactor and tested in a gas turbine. Based on feedstock availability, two main constituents of the fuel were selected as spruce wood and glycerol. Di-ethylene glycol was also added to lower viscosity of the fuel [11]. Due to the wide availability of lignocellulosic waste and glycerol (large amounts of residual glycerol are produced by trans esterification of vegetable and animal fats to biodiesel), the content of these two components was maximized in accordance to limitations of the process, process conditions and end product physiochemical properties. Wood and glycerol content is therefore limited to 33%, which is in line with current stage of research and relatively novel approach to fuel production. The solvent was thus prepared with identical shares of glycerol and di-ethylene glycol to ensure sufficiently low viscosity of the fuel. 3% TsOH acid was added as a catalyst.

36.38% glycerol; 36.38% diethylene glycol; 2.25% TsOH; 25% wood. This type was previously tested in [11] and it is used here to show the influence of LW composition on exhaust emissions. It yields relatively low viscosity of 80 cSt at 100 °C (see Fig. 1) and low pH value. The evaluation of emissions characteristics in the measured points was repeated in the series of experiment presented in this study to avoid the influence of deviations in ambient conditions and minor changes in the experimental setup. Table 2. Elemental composition of LW with 25% wood content Value

Property

Value

C [wt.%] 47.60 H [wt.%] 7.98 N [wt.%] 0.19 Stoichiometric ratio Density [kg/L] LHV [MJ/kg] Viscosity at 100 °C

Property

S [wt.%] O [wt.%] pH value

0.89 43.34 2.5 6.8:1 1.30 20.2 80 cSt

1.1.2 LW Type no. 2 32.33% glycerol; 32.33% diethylene glycol; 2.00% TsOH; 33% wood. This type features lower feedstock costs. The increased proportion of wood is still within the allowed limitations of the liquefaction process, however it raises the viscosity of the fuel, which is nearly 2 times higher (171 cSt at 100 °C) than for aforementioned type with 25% wood content, as shown in Fig. 1. This is the consequence of higher amount of lignin and cellulose degradation products with higher molecular weights. The viscosity of D2 diesel fuel is also shown for the purpose of comparison. The elemental composition and thus also the heating value of LW type no. 1 are very similar to the ones of the LW type no. 2. This is mainly due to the almost identical elemental ratio of O/C and H/C of glycols and wood.

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the addition of 0.71% of 25% ammonia solution, the pH value of LW type no. 1 was increased from 2.5 to 5.5 to mimic diesel fuel.











 





  

  



 



    



 







 

   





  

 

 

Fig. 1. Viscosity of D2 and LW with 33 and 25% wood content











1.1.3 LW Type no. 3 36.38% glycerol; 36.38% di-ethylene glycol; 2.25% TsOH, 25% wood, 0.71% NH4OH. The basis for this LW type is the reactant ratio used in LW type no. 1, partially neutralized by the addition of ammonium hydroxide. Partial neutralization of the acid catalyst has a beneficial effect on corrosivity of the fuel which effectively broadens the selection of fuel system materials to lower grade stainless steels. In addition, partially neutralized acid catalyst shifts the liquefaction reaction threshold to higher temperatures and by this increases the stability of the fuel. Undesirable effects that are experienced due to the presence of alkali metals in the fuels for the turbine engines (i.e. ash and slug formation) were the main driver for avoiding other bases such as NaOH or KOH and thus favoring NH4OH. Table 3. Elemental composition of LW with 25% wood content and elevated pH value Property Value C [wt.%] 47.52 H [wt.%] 8.00 N [wt.%] 0.34 Stoichiometric ratio Density [kg/L] LHV [MJ/kg] Viscosity at 100 °C

Property S [wt.%] O [wt.%] pH value

Value 0.89 43.26 5.5 6.8 : 1 1.30 20.2 80 cSt

The addition of ammonium hydroxide slightly altered the elemental composition of the fuel due to added nitrogen, however, the molecular composition remained almost unchanged, thus the viscosity was the same as for LW type no. 1 (80 cSt at 100 °C). By 224





Fig. 2. Composition of different tested LW types

2 METHODS Combustion research was performed on an experimental turbine presented in [11]. For the purpose of this study, the fuel preheating system was upgraded with pre-filtering elements, designed to withstand the increased viscosity of LW type no. 2. Adaptation to increased viscosity was also performed on the piping and pumps. This was of particular importance to enable cold starting of the fuel system, where sufficiently low pressure losses were required (viscosities at room temperatures were roughly 3000 cSt for LW type no. 1 and 14000 cSt for LW type no. 2). The system was therefore capable of handling heavy and corrosive fuels and able to achieve required range of fuel flows (14 to 19 kg/h) and supply pressures (2.3 to 2.7 bar). Piping connection to the gas turbine was achieved through a three-way valve. This three-way valve allows utilization of the diesel fuel while starting and stopping the gas turbine. Data were collected at three different turbine inlet temperatures (TIT) for each LW type; 750, 800 and 850 °C. These TIT values were achieved by altering the fuel mass flow. TIT was therefore a steered parameter, whereas air flow, EQR, pressure ratio and primary air temperature were the dependent variables. Based on the conclusions of the study in [11], fuel preheat temperature was kept constant at 100 °C. Higher preheating temperatures exhibited proneness to formation of carbon deposit on nozzle discharge surfaces due to accelerated thermal degradation of the fuel. This causes an off center spray pattern, which results in carbon buildup on combustion chamber

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walls and thus in excessive CO emission and impaired operational stability, being highly undesirable. The experimental system was equipped with the following. The fuel preheater had K-type thermocouples, PT-100 RTDs, coriolis mass flow meter and capacitive pressure sensors to closely monitor the conditions that fuel was exposed to. The gas turbine used K-type thermocouples, capacitive pressure sensors, laminar flow meter and emission measuring hardware to characterize the operating point and combustion process. An exhaust gas sample (~3 L/min) was continuously pumped through a heated line (190 °C) to FID, NDIR and CLD detectors to avoid any condensation of hydrocarbons or water vapor. Measured emissions values were postprocessed and translated from measured volume concentrations to concentrations corrected to 15% O2 in exhaust gasses by Eq. (1). This was done to remove the ambiguities linked to different EQR in different operating points. This is also a widespread practice to specify the permissible emission levels [18].

∅i 15% O2 exh = ∅i

EQR15% O2 exh EQR i

. (1)

3 RESULTS AND DISCUSSION The main scope of the study was to determine the influence of different LW types on emission trends and underlying formation mechanisms, basing on acquired thermodynamic and exhaust gas composition data. Interdependencies between known fuel compositions and gaseous emissions were then defined and explained. To provide benchmark thermodynamic and emissions results, conventional diesel fuel was used. First, the emissions and thermodynamic parameters of each LW type are analyzed separately to enlighten the interaction between the emission formation phenomena and chemical and physical properties of different types of LWs. This will be followed by a comparison of emission concentrations for all LW types with added results of a benchmark diesel fuel. 3.1 LW Type no. 1 It is clearly discernible from Fig. 3 that CO and THC emissions substantially decrease with increasing TIT, whereas NOx emissions increase with TIT. This opposing trend can be explained by the molecular composition of the fuel. One of the main wood constituents is the lignin which in comparison to

cellulose contains high number of cyclic structures in its basic monomers [19]. Generally, the liquefaction process depolymerizes lignin into its basic monomersmonolignols (i.e. p-coumaryl alcohol, coniferyl alcohol, sinapyl alcohol [16]) which still contain an aromatic ring that usually requires high activation energy to form combustion supporting radicals. Thus, high dependency of CO emissions on TIT is mainly related to the fact that large amount of species with high autoignition temperature do not reach their temperature threshold to dissociate or are dissociated too late in the primary zone, causing an insufficient residence time of the mixture. Increasing TIT is thus clearly beneficial in reducing CO and THC emissions - the overall temperature field in combustion chamber is shifted towards higher temperatures which slightly postpones the reaction quenching and effectively helps to reduce the emissions of intermediate combustion species. An additional mechanism that contributes to reduced CO emissions at high TIT is increased turbulence in the primary zone, which is predominantly caused by higher volumetric flows of the air (Fig. 4). Increased turbulent kinetic energy increases the rate of spray evaporation and mixture preparation. Despite a relatively low LHV and high viscosity , the CO emissions can be considered as relatively low for this fuel. This is to a large extent related to the high oxygen content of LW. It can substantially increase the mixture homogeneity if sufficient dissociation is achieved as 22% of stoichiometric oxygen is available through the fuel bound oxygen atoms. This favors formation of OH radicals, which are the essential species for CO oxidation reactions [20]. High oxygen content is perceived also through low stoichiometric ratio of LW which is only 6.8. Mechanisms of THC emissions are similar to those of CO and therefore usually follow the same trends. They consist of only partially dissociated fuel molecules and are therefore a preceding step before CO formation occurs. They most likely originate from highly temperature resistant hydrocarbons and undesirable reaction quenching in cold spots of the combustion chamber. Unlike CO and THC, NOx concentrations slightly increased with increasing TIT. This is in-line with Zeldovich NOx formation mechanism, where higher temperatures in primary zone cause the N2 molecule to dissociate into atomic state and then further oxidize to NO and/or NO2. With its high activation energy, the separation of N from N2 occurs only at very high temperatures (over 1600 °C) [21] and is therefore favored only in the immediate vicinity of the flame

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zone. However, it is believed that the major component of NOx emissions in gas turbines is the Fenimore NOx mechanism, which originates from reactions of HCN and CN radicals with oxygen. The Fenimore mechanism features high rates in the areas exhibiting relatively low temperature (~700 °C) and fuel-rich conditions [20]. Its sensitivity to TIT is thereby linked to EQR elevation at high TIT and thus richer conditions in primary combustion chamber zone. With larger quantity of delivered fuel, the amount of FBN (fuel bound nitrogen) is also increased. An increased amount of FBN contributes to higher NOx emissions although this effect is not visible in the presented charts as the correction to 15% O2 is applied to the measured data. The overall perceivable NOx increase is therefore a combination of two mechanisms Zeldovich and Fenimore NOx. These two are believed to have a major influence and this trend is observed also if no correction to the measured data is applied.



 



  





















 



 





  



 















     

       

 



 

  





  

  

Fig. 3. Emissions, corrected to 15% O2 for LW with 25% wood content

 

Fig. 4. Operating parameters of experimental system firing LW with 25% wood content

226

3.2 LW Type no. 2 The increased wood content (by 8%) in this type of LW is manifested mainly through altered molecular composition. The elemental composition and thus also heating value remains unchanged due to very similar elemental compositions of glycols and lignocellulosic biomass. Perceivable differences in LHV and elemental ratio would be visible only if very large changes in reactant ratios would be made. As a larger amount of monolignols is contained in LW type no. 2, increased share of cyclic hydrocarbon molecules could slightly increase ignition resistance of the LW. Generally, a higher number of molecules with larger molecular mass in comparison to LW type no. 1 with 25% wood content also lead to higher viscosity (80 cSt for 25% wood content and 171 cSt for 33% wood content) which significantly influences atomization ability. Decreased evaporation rate is the consequence of larger initial droplet diameters and altered evaporation curve of the fuel. These two mechanisms are the main reason for late formation of the mixture and thus a main driver of a significant increase of CO and THC emissions compared to those of LW type no. 1 presented in section 3.1. Larger droplet diameter contributes to the increased spray penetration depth and lower evaporation rates. Both mechanisms extend combustion into the secondary combustion chamber zone. Here, lower temperatures and lower local EQR could impede final oxidation stage of carbon and hydrocarbon molecules leading to increased CO and THC emissions. As mentioned above, another aspect of altered molecular composition is its cyclic hydrocarbon content. It is anticipated that even if the viscosity of LW with 33% wood content would be the same as for LW with 25% wood content, CO and THC emissions would still be higher due to increased content of monolignols and their high ignition resistance. Elevated CO and THC emissions are therefore a consequence of altered molecular composition which influences the combustion process in two ways: a) by impairing the atomization process and b) by increasing ignition resistance of the mixture. NOx emissions of LW type no. 2 are lower by approximately 20 ppm compared to those of LW type no. 1 at lower TIT (Fig. 5). In the upper TIT end, this reduction is slightly lower; roughly 10 ppm. The sensitivity of LW composition on NOx reduction mechanisms is therefore declining with increasing TIT and thus the influence of all three mechanisms on this sensitivity is analyzed subsequently. Focusing on FBN, its content should be roughly 8% higher as the only source of nitrogen is wood feedstock

Seljak, T. – Kunaver, M. – Katraťnik, T.




 



  





















 



 





  



 



Fig. 5. Emissions, corrected to 15% O2 for LW with 33% wood content

Focusing on the upper explanations it can be summarized that the main drivers for reduced NOx concentrations of the LW with increased content of wood are: 1) the combustion chamber temperature. 2) the flow field which is significantly influenced by different fuel properties and influences NOx formation mechanisms, and 3) NOx formation paths as explained above. This, however, does not alter the global thermodynamic parameters as they closely resemble those exhibited by LW type no. 1 as is evident from Fig. 6. Since water content was not analyzed, some of NOx reduction could also be attributed to higher water content of LW type no. 2, which can be related to a higher share of a lignocellulosic material in the fuel.













     

       

 



 

  





  

which is in the case of analyzed LW type increased by 8%. Contribution to cumulative NOx is therefore in the range of 4 to 8 ppm. The exact numbers were acquired through equilibrium equations for complete combustion assuming a 1/3 and 2/3 efficiency of FBN conversion to NOx; these rates were chosen based on conversion efficiencies of FBN determined in the next section. Partial reduction of NOx emissions when comparing them to those of LW with 25% wood content arises from aforementioned postponed combustion and lower flame temperatures due to lower local EQR as described above. The reduction of NOx concentrations is therefore a consequence of suppressed Zeldovich mechanism due to lower temperatures and also suppressed Fenimore mechanisms due to lower EQR ratios. Considering that Fenimore mechanism has a relatively low activation energy and is maintained even in the areas where temperatures are as low as 700 °C [20], the main cause for reduction of Fenimore NOx component is most likely lower local EQR ratio and not the lower temperature as in the case of the Zeldovich component.

  

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Fig. 6. Operating parameters of experimental system firing LW with 33% wood content

3.3 LW Type no. 3 Thermodynamic parameters of the experimental turbine remained unchanged also with LW type no. 3 as visible in Fig. 8, thus providing comparable conditions to LW type no. 1 and no. 2. When observing emissions concentrations in Fig. 7, the hypothesis that addition of ammonia and change in pH value of the fuel will not affect CO and THC emissions is clearly confirmed, since the emissions concentrations and their trends are almost the same as those with LW type no. 1. The reason for this lies in the negligible influence of partial neutralization on molecular composition: a) atomization quality is not affected by change in viscosity or surface tension as these properties are not affected by neutralization and b) the addition of ammonia does not alter the evaporation curve of the LW. The local EQR ratios of the mixture, air temperatures and oxidation kinetics of hydrocarbon molecules are therefore very close to those encountered with the LW type no. 1. More pronounced differences are perceivable in NOx concentrations which are elevated by roughly 30 ppm over the entire TIT interval. This nearly constant shift to higher values supports the hypothesis that NOx emissions trends are the same as those of LW type no. 1, although the absolute value of emissions is elevated. This increase in NOx emissions is mainly related to conversion of FBN to NOx. This is reasoned by the fact that this mechanism is usually almost insensitive to EQR and temperature field, whereas it is clearly supported by the elevated nitrogen share in the fuel. There are strong indications that Zeldovich and Fenimore contributions to the overall NOx emissions are nearly identical for the LW type nos. 1 and 3. This is reasoned by the fact that TIT and global EQR of

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these two LW types are identical, and furthermore nearly identical CO and THC emissions also indicate very similar temperature and concentration field within the combustion chamber. Therefore it is believed that the difference in NOx emissions is the consequence of increased amount of FBN to NOx conversion. Calculation revealed that 0.1 mol/L of nitrogen was added to the LW with 25% wood content to produce LW type no. 3. This effectively increased the share of FBN almost by a factor of 2. As stated in [20], conversion efficiency from FBN to NOx is around 2/3 for some types of fuels (propane with 2500 ppm addition of methylamine).

  

 























 





  



 















     

        



 

  





  

  

Fig. 7. Emissions, corrected to 15% O2 for LW with 25% wood content and elevated pH value



 

Fig. 8. Operating parameters of experimental system firing LW with 25% wood content and elevated pH value

This can be compared with the conversion rate of the FBN to NOx when firing LW type no. 3, assuming that FBN to NOx is the only contributing mechanism to the difference in NOx between LW type no. 1 and 3. In this case it can be calculated that conversion efficiency of FBN to NOx when firing LW in aforementioned conditions is roughly 1/3 (i.e. 33%). Clearly, nitrogen bound in wood decomposition 228

3.4 Influence of Fuel Type on Emissions



 



products or nitrogen added through ammonium hydroxide could exhibit different combustion kinetics and thus also different conversion efficiencies to NOx. This could cause a large discrepancy between predicted and actual FBN conversion efficiency so the proposed 33% FBN to NOx efficiency should only be used as a rough estimate under similar conditions to those employed in this study. It was, however, already proven in [21] that high O/N ratio in fuels could reduce FBN conversion efficiency. This thesis is supported by the high oxygen content of LW and the measured data.

In this section, emissions for each of the tested fuels are given. The results are grouped by emission species to illuminate the influence of fuel type on each of the emission species. Thermodynamic parameters of operation with diesel fuel were very close to those of all LW types to assure comparable conditions in combustion chamber in terms of airflow, temperature and pressure. In Fig. 9, CO emissions are shown. Trends are in line with phenomena explained above and can be mainly attributed to: a) viscosity of the fuels (with LW type no. 2 having the highest, followed by LW type no. 1 and no. 3 and significantly lower viscosity of diesel fuel); b) density of the fuels (with all LW types having higher density than diesel fuel by roughly a factor of 1.5); c) molecular composition of the fuels (with LW type no. 2 having higher content of cyclic hydrocarbons than LW type no. 1 and no. 3 and all LW types having higher molecular weights than diesel fuel). Atomization of LW is most likely notably impaired by a) and penetration depth is further increased by b). The combination of these two facts reduces surface to mass ratio of the droplets which is most likely together with c) the reason for reduced evaporation rate of the fuel. Besides decreased volatility due to heavier molecular weights, c) also influences the combustion process due to cyclic hydrocarbon content with its high ignition resistance. The CO emissions are therefore the highest with LW type no. 2, followed by LW type no. 1 and LW type no. 3. The latter two exhibit similar concentrations due to similar properties under a), b) and c). With diesel fuel, very low concentrations were achieved due to more favorable physical and chemical characteristics of diesel in comparison to all LW types allowing for fast

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mixture formation and very low rate of mixture escape into colder parts of combustion chamber.

  



       





 





 







phenomena as formation of CO. Again, very low concentrations were measured with diesel fuel, as the density, viscosity and volatility favour much faster mixture formation and thus a lower rate of combustion reaction quenching in cold parts of the combustion chamber (namely cooling and dilution air). NOx emissions are presented in Fig. 11, with emissions of LW no. 3 being the highest, followed by LW type no.1 and diesel fuel. The lowest concentrations are observed with LW type no. 2. The results support the aforementioned influences of FBN and temperature field on NOx formation mechanisms, with high FBN content promoting NOx formation as in the case of LW type no. 3 and reduced NOx formation rate with LW type no. 2 due to postponed mixture formation and thus lower flame temperatures.

Fig. 9. CO emissions for different LW types 

          

  



3 CONCLUSIONS







 





  







Fig. 10. THC emissions for different LW types

  



       







 





  







Fig. 11. NOx emissions for different LW types

Similar trends are visible for THC emissions in Fig. 10, confirming the hypothesis that formation of unburned hydrocarbons is influenced by the same

Three different types of lignocellulosic biofuel were tested in a gas turbine to investigate the effects of increased wood content and altered pH value of the fuel on the emissions of CO, THC and NOx. Results indicated that combustion process is strongly affected by content of lignocellulosic material, while the effect of partial neutralization is isolated to influence on NOx emissions. The emission characteristics of the LW types were as follows. LW type no. 1: It exhibited relatively high CO and THC emissions, highly dependent on TIT in comparison to emissions of a diesel fuel. In literature, this was attributed to specific molecular composition and cyclic hydrocarbon content as well as viscosity and density. NOx emissions showed increasing trend over TIT where increase in Zeldovich and Fenimore mechanisms of NOx formation is believed to have a major role. LW type no. 2: CO and THC emissions increased in comparison to LW type no. 1 as a consequence of higher wood content. The reason for this is altered molecular composition of the fuel, resulting in viscosity increase and subsequent deterioration of atomization quality. NOx emissions are slightly reduced due to altered flow and temperature field, caused by reduced mixture formation rate and consequent local EQR ratios. LW type no. 3: CO and THC emissionswere very similar to LW type no. 1. However, partial neutralization was found to influence NOx emissions. An increase of 30 ppm was visible over entire operating range of the turbine which was accounted to increased FBN originating from ammonia-borne nitrogen. The conversion of FBN to NOx was found

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to be roughly 33%. CO and THC emissions were unchanged in comparison to LW type no. 1, confirming the hypothesis that partial neutralization does not influence the oxidation kinetics of hydrocarbon molecules. Diesel fuel: Absolute values of emissions were significantly lower (by an order of magnitude) for CO and THC. This is in line with physical and chemical characteristics of different LW types which have unfavorable influence on the combustion process mainly because of delayed mixture formation and influence of cyclic hydrocarbons on combustion kinetics. NOx emissions are comparable to LW type no. 1. Despite the elevated CO and THC emissions, LW type no. 2 exhibits promising trends as these emissions significantly reduce with increasing TIT. Moreover, they would most likely further reduce if more a advanced gas turbine with higher TIT were used. An adapted combustion chamber geometry could further help to prevent excessive mixture escape into the secondary combustion chamber zone. Additionally, the low impact of added base in LW with 25% wood content on combustion and emissions could greatly improve the applicability of such fuel in gas turbines by reducing the costs for fuel processing systems and possibly filtration methods without impeding combustion performance and CO emissions, while moderately increasing NOx emissions. 4 ACKNOWLEDGEMENTS The authors acknowledge the financial support from the Ministry of Education, Science, Culture and Sport of the Republic of Slovenia through the contract no. 3211-10-000057 (Centre of excellence Polymer Materials and Technologies), University of Ljubljana for support through Innovative Scheme, contract no. 323 and finally to Slovenian Research Agency for support through contract no. L2-5468. 5 NOMENCLATURE D2 Diesel fuel LW Liquefied wood Carbon monoxide CO THC Total hydrocarbons Nitrous oxides NOx LW Liquefied wood H2SO4 Sulfuric acid TsOH Tossylic acid LHV Lower heating value NH4OH Ammonium hydroxide 230

FBN FID NDIR CLD TIT ∅i

Fuel bound nitrogen Flame ionization detector Nondispersive infrared Chemiluminiscence detector Turbine inlet temperature Measured concentrations

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Original Scientific Paper

Received for review: 2013-11-28 Received revised form: 2014-01-09 Accepted for publication: 2014-01-17

Electric or Hydraulic Energy Recovery Systems in a Reach Truck– A Comparison

Minav, T. – Hänninen, H. – Sinkkonen, A. – Laurila, L. – Pyrhönen, J. Tatiana Minav1,* – Henri Hänninen1 – Antti Sinkkonen1 – Lasse Laurila2 – Juha Pyrhönen2 1 Aalto University, School of Engineering, Department of Engineering Design and Production, Finland 2 Lappeenranta University of Technology, LUT Energy, Finland In this paper, electric and hydraulic regeneration methods of recovering potential energy from an electro-hydraulic forklift truck are studied. Two similar forklift setups equipped with either electric or direct hydraulic energy storage are compared. In the first setup, the forklift lifting system is controlled directly with an electric servomotor drive. The servomotor drives a hydraulic pump capable of also operating as a hydraulic motor during lowering motion. In the second setup, the hydraulically operated forklift is equipped with an energy recovery system consisting of pressure accumulators for storing energy and a hydraulic digital valve package for precise leakage free flow control. This paper describes the arrangements of the experimental setups. The results of the proposed systems are then compared from the energy efficiency point of view. Energy-savings ratios for electric and hydraulic test systems were calculated for different fork velocities and payloads. Keywords: digital flow control unit, electric energy recovery, energy storage, forklift, hydraulic energy recovery, hydraulics, lead-acid battery, hydraulic accumulator, permanent magnet synchronous machine, reach truck, supercapacitor

0 INTRODUCTION Globally, energy efficiency and energy savings have become important practical research topics in nonroad mobile machinery [1] and [2]. In [3] and [4] energy saving lifting hydraulic systems and control techniques [5] have already been suggested. However, energy savings are still very important in non–road mobile machine applications, e.g. in excavators [6] and [7] and in machines operated purely by accumulatorstored electric energy [8] and [9]. In order to reduce the energy consumption of a machine, either the efficiencies of the components have to be improved or energy that is otherwise lost in the process has to be utilized by regeneration. In many cases the latter can be accomplished most advantageously by reusing the kinetic or potential energy of the machine or its subsystem [10] and [11]. Depending on the system and process in question, the utilization energy recovery can lead to significantly lower overall energy consumption and, with mobile machines, to longer operating times. [12] and [13] When considering regeneration of energy, the work cycles in which forklifts often operate include bidirectional material or payload transfers, which provide an opportunity for efficient recovery of potential energy. In this study, reach trucks (a subtype of forklifts) are modified to allow energy recovery from the payload of the mast’s lifting/lowering function. In the case of the other functions of the machine, there is no potential energy to be recovered, and kinetic energy levels are too low for any feasible recovery system. 232

There are several base technologies on which to build a recovery system; thermal, mechanical (i.e. fly-wheel or counterweight based recovery systems), electric or hydraulic. This study focuses on the last two types. A well-established method of recovering energy in mobile working machines is an electric recovery system. This system type usually consists of an electric motor/generator, an inverter, possibly a DC/DC converter, a battery, and also, in some cases, an electric double-layer capacitor (EDLC) [14]. The advantages of this kind of system are control flexibility, compactness, efficient control, and fairly high, energy efficiency [15]. With regard to the mast operations of a hydraulic reach truck, another well-known option is to use a direct or indirect hydraulic recovery system. In an indirect hydraulic storage system consisting of a hydraulic motor-pump, a controllable hydraulic pumpmotor, and a hydraulic accumulator, the flexibility of control is as good as that of electric recovery. This system first converts the hydraulic energy into mechanical and then back to hydraulic energy, requiring as many conversions as the electric storage system. In this paper, however, the direct hydraulic recovery system is compared with the electric recovery system. The direct hydraulic recovery system removes the need for energy conversions from the hydraulic to the electric form in the recovery phase and vice versa in the regenerating phase. A direct application of the hydraulic accumulator has more limiting factors than the indirect recovery system [16] and [17]. The utilization of this type of system requires two flow control edges in the recovery phase in order to maintain controllability

*Corr. Author’s Address: Aalto University, School of Engineering, Espoo, Finland, Tatiana.minav@aalto.fi


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of mast velocity. This is achieved by utilization of a digital valve package (DFCU). This study compares electric and direct hydraulic recovery systems with each other in terms of energy efficiency. The operational characteristics of both systems are also analyzed. 1 TEST SETUPS This section describes the studied system setups, which includes a description of energy evaluation and work cycle. As two different setups were used, an electric recovery setup was located at the Lappeenranta University of Technology, and a hydraulic recovery setup at Aalto University. 1.1 Electric Recovery System Setup The original non-regenerative AC electric drive and the hydraulic system of the Humanic HS-16F5400 forklift were replaced with the schematics shown in Fig. 1. The electric motor servo drive directly controls the fixed displacement hydraulic pump speed and thereby the position of the hydraulic cylinder piston instead of a traditional proportional valve. The twoway normally closed poppet valve is used as a safety valve, which prevents the load from dropping in the case of a failure. For lifting, the hydraulic pump produces a flow depending on the rotational speed of the servomotor. While lowering a mass, the potential energy forces the hydraulic machine to rotate as a motor, and the electric machine acts as a frequencyconverter-controlled generator [18]. The converter controls the generator torque and actively rectifies the generated electric energy to the DC link. Because of the relatively short lowering period (around 10 s), recharging of conventional lead acid batteries is considered inefficient [19]. For energy measuring purposes, a brake resistor was used as the “energy storage”. At the moment, super capacitors seem to be the most suitable solution for fast recharging. The measured super capacitor charge-discharge cycle efficiency of 99% [4] will be used to estimate the cycle efficiency of the future system equipped with electrical energy storage. In [4], the measured voltage and current signals of the forklift electric recovery setup were used for the super capacitor efficiency measurements. In the forklift electrical energy recovery test setup, a control program was created to control both the electrical and hydraulic parts of the forklift system [20]. The instrumentation of the system covers measurement devices for pressures, rotational speed,

torque, load position, phase voltages and currents, and DC voltage and DC current. The energy consumption in this paper was calculated from the measured current in the DC link. Measurements were carried out utilizing dSpace-measurement software. The Converter software was used to measure the rotating speed of the permanent magnet synchronous machine (PMSM) and to estimate the motor torque. Two S-10 pressure sensors manufactured by WIKA were installed to measure the pressures at the pump outlet and between the 2/2-valve and cylinders. Yokogawa PZ4000 Power analyzers with a sampling time of 10 μs were used to measure the phase voltages and currents. The speed and height of the fork were measured by a wire-actuated encoder SGW/SGI from SIKO. HITEC Zero-Flux B 2000 current sensors were used. The accuracy of the sensors can be considered acceptable for these test purposes.

DC

Fig. 1. Electric and hydraulic circuits of the main lift function with energy regeneration from potential energy; the experimental system consists of single-acting cylinder (I free lift zone, II second cylinder zone), two-way normally closed poppet valve, pressure relief valve, hydraulic pump/motor, oil tank, permanent magnet synchronous motor/generator, phase voltage and phase current probes, frequency converter and brake resistor Rbrake, DC voltage and DC current probes

1.2 Hydraulic Recovery System Setup The hydraulic recovery setup is based on a fairly similar truck model, the Humanic HX-16. The simplified hydraulic system of the forklift is shown in Fig. 2. The main components in the studied energy recovery system are the pressure accumulators and the digital flow control unit (DFCU). The DFCU consists of two individually adjustable control edges

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qV = Cq A

qV,accu

qV

qV,tank

Fig. 2. Simplified circuit diagram of the hydraulic storage test system

qV = K ⋅ ∆p z . (2)

The corresponding values of constants K and z are individually identified for each passage in the DFCU and adapted to the controller. The identification was done by manually finding the values for the said constants in Eq. (2) to match the measured pressure difference – flow curve of each passage. This approach has been found to be a viable method for taking into to account both the turbulent losses in the orifice and in the valve, in addition to the partly laminar losses in the flow paths of each passage [21]. The recovered energy is utilized in the following lift phase by directing it to the pump inlet, thus reducing the pressure difference over the pump and thereby decreasing the power needs of the electric motor driving the pump. 234

DFCU

2∆p , (1) ρ

is modified to:

The instrumentation of the test bench covers measurement devices for pressures, flows, temperatures, rotational speed, torque, load forces, load position, battery voltage, and current. In this study, however, the flow sensors were bypassed to achieve the full recovery potential of the system. The current transducer used was an LEM DH 500 B420L B and the load position was measured with a Waycon SX120-6000-420A-SA draw wire sensor. The energy consumption reduction results given in this paper were calculated from the measured battery voltage and current drawn from it. Measurements were carried out using Matlab/Simulink xPC Target software, which also included the controller for the DFCU.

Volumes: 4 litres each

containing five poppet-type on/off-valves each, all paired with differently sized orifices. The individual adjustability of the control edges is needed for the dynamic division of volume flows between the accumulator package and the tank. A hydraulic accumulator is a device that stores pressurized hydraulic fluid with an internal nitrogen gas volume enabling the energy storing. The accumulators, manufactured by Hydroll with nominal size of 4 liters, used in this study are of a piston-type, which consist of oil and gas chambers separated from each other with a piston. The pre-load pressure level of the gas chamber determines the maximum energy content of the accumulator and affects the efficiency of the recovery. Thus, for efficient operation, the pressure level must be adjustable. The altering of preload pressure to a higher value between tests, when needed, was done utilizing an external gas container. For operating the DFCU, a cost-function-based controller was built to determine which valves of the DFCU (both control edges) are to be opened and which closed in order to simultaneously perform the charging of the accumulators and provide the required lowering of velocity. The controller calculates the flow through the control edges using the data from the pressure transducers. For these calculations, the equation for turbulent flow through an orifice given by:

1.3 Description of Performed Tests The experimental setups were tested with payloads of 0, 500, and 1000 kg at different motor speeds. The velocities of the forks in both cases were set to 0.2, 0.3, and 0.4 m/s. The travel distance had to be limited to 1.6 m due to the maximum measurement time of 10 seconds with a sampling time of 10 µs of the Yokogawa PZ4000 Power analyzer in the electric recovery setup. The measurements were made in the free lift zone using the first cylinders of the telescopes, resulting in a low tare in this case. This was done in order to attain better correspondence between the pressure levels of the two systems, since the moving structural masses of the masts are fairly similar in free lift zone, but differ greatly in second cylinder zone. For the electric recovery setup, lifting and lowering motions were measured separately due to

Minav, T. – Hänninen, H. – Sinkkonen, A. – Laurila, L. – Pyrhönen, J.


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limitations of used Power analyzer. For the hydraulic recovery setup, measurements for each measurement point were performed in a single cycle consisting of a continuous sequence of lifting, lowering and lifting phases. The first lift phase in the cycle is executed without assistance from the accumulators and the latter lift phase with the assistance of energy recovered during lowering phase. 1.4 Detailed Information on Test Platforms Details on the main components of both test platforms are presented in Table 1. Table 1. Main differences between platforms Test setup

Parameter Theoretical volumetric displacement of the pump [m3/rev] Motor Converter Piston: crosssectional area of the free lift cylinder [m2] Maximum stroke of free lift cylinder [m]

2.2 Energy Efficiency

Electric recovery

Hydraulic recovery

13.3·10-6, manufactured by Erker 10 kW CFM112M PMSM manufactured by Sew-eurodrives ACSM1-04x4) x5)−046A-4 by ABB

19·10-6, manufactured by Parker

MHI 16A70-04020, A001

0.0026

0.0033

0.88

1.35

14 kW IM by Danaher

2 EVALUATION OF ENERGY UTILIZATION Before investigating any energy-saving system, it is necessary to discuss how to evaluate the utilization of energy in an electro-hydraulic forklift. This section introduces the definitions used for the evaluation of energy utilization in the test setups: 2.1 Efficiency Efficiency as a function of time η(t) is normally defined as a ratio between the output (Pout) and input (Pin) powers:

η (t ) =

efficiency of a motor without time dependence. Measuring of the rated efficiency of an electric motor normally takes several hours in order to reach the thermal equilibrium of the machine before defining the efficiency. In the case of a limited linear movement, however, it is very difficult to apply this definition of efficiency as there is no steady state in the operation. Even when abandoning the need to reach thermal equilibrium there is only a few seconds of constant speed operation at some “constant” efficiency, and thus, measuring efficiency becomes very difficult. In the case of a forklift, we are actually not very interested in the instantaneous efficiency of the system but in the ratio of the total output to input energies, which is how energy efficiency is defined.

Pout (t ) . (3) Pin (t )

Even though the efficiency η(t) is a function of time, it is normally measured while trying to keep Pout and Pin as constant as possible in a static situation in order to be able to obtain, for example, the rated

The energy efficiency ηenergy(t) for a time interval [t1, t2] is defined as: t2

∫P

out

ηenergy (t ) =

t1 t2

(t )dt

∫ P (t )dt in

=

Eout , (4) Ein

t1

where Eout is the total output energy and Ein is the total input energy of the system during the time interval starting at t1 and ending at t2. Energy efficiency should normally be regarded as a comprehensive term taking the whole life cycle of the system into account. Therefore, the cycle efficiency is defined for each test setup. The calculation of energy efficiency is described in detail in [18] for the electric energy recovery setup and in [16] for the hydraulic recovery setup. 2.3 Energy-Saving Ratio In order to compare the different test setup efficiencies, the energy-saving ratio Γs is defined as:

Γs =

Eold − Enew , (5) Eold

where Eold is the energy consumption of the forklift without energy recovery and Enew is the energy consumption of the forklift with energy recovery. This ratio Γs describes how much energy can be saved when energy recovery is used. The energy consumption Eold of the electric drive forklift without energy recovery and Enew energy

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consumption of the forklift with energy recovery for an electric recovery test setup is defined as:

Ecycle_old = Emot / (ηSC ⋅ηinv ), (6)

Ecycle_new = Eold − Ebrake ⋅ηSC , (7)

where Emot is the input energy from the electric motor, ηinv is the inverter efficiency, Ebrake is the recovered energy, and ηSC is the discharge efficiency of the supercapacitor, assuming that the charge efficiency is equal to the discharge efficiency. Therefore, the energy-saving ratio Γs can be defined for an electric recovery test setup as:

Γs =

t2

(

)

2 Ebrake = ∫ ibrake ⋅ Rbrake dt , (9) t1

t2

Emotor = ∫ ( ia ua + ibub + icuc ) dt. (10) t1

as:

For the hydraulic recovery test setup, Γs is defined

Γs =

Eunassisted − Eassisted , (11) Eunassisted

where Eassisted is the energy consumption (calculated from the measured electric power drawn from the battery pack) with the hydraulic assistance on and Eunassisted is the energy consumption with the assistance off. The energy consumptions for both the assisted and unassisted cases are defined as:

t1

t1

t0

t0

E = ∫ Pdt = ∫ (U ⋅ I )dt , (12)

where U and I are voltage and current, respectively. Since the energies are calculated from discrete measurements, Eq. (12) is discretized to:

∆E = ∑ P ⋅ ∆t = ∑U ⋅ I ⋅ ∆t , (13)

where Δt is the sampling interval. 236

This section reports the results obtained from the measurements described in the previous section. Tables 2 and 3 show the energy saving ratios for different fork’s velocities and payloads for the electric and hydraulic recovery test setups, respectively. Fig. 3 below shows a graphical comparison of the energy saving ratios of the electric and hydraulic setup (Tables 2 and 3 (single pre-load pressure results)). Table 2. Energy-savings ratios for different speeds and payloads for the electric recovery setup

Ebrake ⋅ηSC Ebrake ⋅ηinvηCDSC = . (8) Ecycle_old Emot

The ηCDSC is the charge-discharge efficiency of the supercapacitor [4]. The inverter efficiency is assumed to be constant and equal to 95%. The calculation of Ebrake and Emotor is described in detail in [19] for the electric energy recovery setup:

3 RESULTS

Fork’s velocity [m/s] 0.4 0.3 0.2

0 0%* 0%* 0%**

Load [kg] 500 17% 25% 34%

1000 25% 31% 36%

*

Indicates motoring mode generating mode; because of its small value, the energy-savings ratio was rounded to zero. ** Indicates

Fig. 3 and Table 2 demonstrate the results obtained using only the free lift zone of the Humanic forklifts. This results in a low tare as the telescopic system is not lifted at all. In first case, the hydraulic recovery system is (Table 3, top section) optimized for each load and roughly for the travel distance. In the second case it is optimized to a one load/distance, and therefore is not optimal in any way for the other loads. Optimization for velocities was not conducted for any case. It should be noted that the energy saving ratio is high where an optimized hydraulic accumulator was used. In Fig. 3, the hydraulic accumulator was optimized for the 500 kg payload and 1.6 m height. The direct hydraulic energy storage results in a good energy saving ratio in such a case. However, when the same settings are used for the 1000 kg mass, the electric energy storage system outperforms the hydraulic storage system by 5 to 14% with corresponding velocities of 0.4 to 0.2 m/s (compare Tables 2 and 3). In Table 3, the optimized pre-load pressure for the hydraulic accumulator gave good results for the 1000 kg case in the hydraulic storage system. However, if this setting is used, there will be no recovery in the case of a 500 kg mass, which shows the vulnerability of the direct hydraulic storage system when using varying loads and heights.

Minav, T. – Hänninen, H. – Sinkkonen, A. – Laurila, L. – Pyrhönen, J.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 232-240

Table 3. Energy saving ratios for different speeds and payloads for the hydraulic recovery setup Fork’s velocity [m/s]

Load [kg] 0 500 Optimal pre-load pressure for each load [%] 0.4 0 26 0.3 0 30 0.2 0 31 Single pre-load pressure (optimized for 500 kg) [%] 0.4 0 26 0.3 0 30 0.2 0 31

1000 41 42 45 20 20 22

Electric Setup 36 34

Hydraulic Setup 31

Electric Setup Hydraulic Setup 31 30 Electric Setup Hydraulic Setup 26

25

25 22 20

20

0

0 0,2

0

0

0

0,3 V [m/s]

1000 kg

0 kg

500 kg

1000 kg

0 kg

500 kg

1000 kg

0 kg

500 kg

500 kg

1000 kg

0 kg

500 kg

1000 kg

0 kg

1000 kg

0 kg

500 kg

17

0 0,4

Fig. 3. Comparison of the energy-savings ratios in % for the electric and hydraulic tests systems, (for hydraulic test setup results for a single pre-load pressure setup are shown), where V is the fork’s velocity

4 DISCUSSION In this study, the target was to produce similar operating conditions for two slightly different forklift trucks. Considering the restraints and similarities of the two systems, the simplest solution was to run tests in the free lift zone, and to limit the travel distance to 1.6 m. The relatively short travel distance favored the hydraulic setup by a few percent points in terms of saving ratios, since the hydraulic accumulator capacity in these tests is limited to 16 liters. The selection of the free lift zone instead of the second cylinder zone affected both of the systems by decreasing the achievable savings ratios. 4.1 Results of the Electric Energy Storage System The results shown in Table 2 for the energy savings ratio of the electric recovery test setup seem low. The

energy-savings ratio increases slightly with increasing load and decreasing speed. With a 0 kg load there was no recovery observed, and during lowering the electric machine was working in the motoring mode instead of generating mode. The maximum energy savings ratio in the free lift zone is 36%. This is significantly less than earlier measurements achieved when operating in the second lifting zone of the forklift where the tare of the system is high and the relative hydro-mechanical losses of the system are lower [20]. According to [20], the maximum energy-saving ratio reached with this same Humanic forklift was 53% when operating in the second lifting zone of the telescope with a 920 kg load. The mass of the moving parts of the mast is larger, which thereby enables the electric drive components to operate closer to their nominal values. There is significantly more potential energy to recover in the second cylinder zone than in the free lift zone. 4.2 Results of the Hydraulic Storage System The measured energy savings ratios of the hydraulic recovery circuit ranged from 0 to 45%. With no load, the system pressure levels remain too low for energy recovery, as in the previous case. This is due to the fact that the total flow losses of the system are in the range of cylinder pressure. By introducing a load, the recovery system becomes effective. Using loadoptimized pre-load pressures in the gas chambers of the accumulators, the measured energy saving ratios ranged from 26 to 31% and from 41 to 45% for loads of 500 and 1000 kg, respectively. Considering the applicability of such a system with load optimized pre-load pressure, it is evident that the loads should remain relatively constant for sufficient durations. Such cases are numerous, for example, warehouses and material handling tasks in industry. In mixed goods (variable loads) warehouses, it would not be advisable to alter the pre-load pressure between each lifting/lowering phase (because of the additional energy consumption in the form of pressurized gas) and, therefore, in most cases the pre-load pressure setting would not be optimal. The measurements indicate that when using a single preload pressure, optimized for a 500 kg load, the savings ratios with a 1000 kg load would drop significantly (to approximately 20%) for the whole velocity range when compared with those using the optimal preload pressure. The effect of pre-load pressure setting on the effectiveness of the energy recovery system is analyzed in greater detail in [22]. In order to devise a hydraulic recovery system that performs better in mixed load situations the research group at Aalto

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University will design and construct an alternative recovery circuit based on a hydraulic transformer. Even though this new system is expected to have a lower peak efficiency (than the optimized direct system), simulations [13] suggest that it has a better overall efficiency. 4.3 Differences in the Two Previous Systems It was observed that in the free lifting zone, the hydraulic recovery setup with its optimized pressure settings showed better results compared with the electric recovery setup. The less efficient behavior of the electric setup can be partly explained by the short travel distance (1.6 m), short time (max 5 s) during which generator can recover the electric energy from the potential energy in the electric recovery setup and lower tare, which does not enable the electric drive components to operate closer to their nominal and efficient values. 4.4 Operational Characteristics The hydraulic recovery system was designed to leave the operational characteristics unchanged and this was also achieved. The system can be used by adopting one of two different operation strategies in terms of pre-load pressure optimisation. Firstly, one could optimize the pressure level permanently for one load. This would be easier from the operator point of view, but the efficiency would not be optimal in most cases. The second strategy can be implemented if the parameters (load and height) for lifting and lowering are known in advance, if they remain relatively constant, and if their quantity is sufficient. In this strategy, the pre-load pressure is optimized to match the known upcoming cycle. This would allow the recovery system to operate at optimal efficiency, but as a drawback it would add an additional work phase, thus consuming energy. The electric recovery system showed very good controllability of the hydraulics side. There are no limitations in the electric recovery setup as long as the electric energy storage selected is large enough to receive the largest possible amount of recovered energy during a single lowering action. With increasing electric machine torque (proportional to the sum of the payload and tare), the system efficiency increases because the electric drive components operate closer to their nominal and most efficient values [23] and [24]. 238

4.5 Other Observations To modify a conventional forklift to recover potential energy, the following actions are required: For electric recovery: the control valve has to be replaced with a two-way normally closed valve; the traditional single-acting hydraulic pump has to be replaced with a hydraulic machine working in both directions; an energy storage, e.g. a supercapacitor bank, has to be added for storing the recovered energy, and the control software of the electric motor has to be updated. Lead-acid batteries can also be used, but supercapacitors have a higher charging efficiency. Current electric recovery setup is operated with a high voltage up to 900 V. It is considered a dangerous voltage level for mobile working machines. In the future, a detailed comparison with a low voltage 48V safe setup will be studied. However, we anticipate that it might have similar results from the energy-savings point of view as the electric energy recovery system evaluated in this paper. For direct hydraulic recovery: the hydraulic circuit has to be enhanced with an additional (leakage free) flow control unit and regeneration valve; a hydraulic accumulator(s) has to be added, a pump has to be altered to a type allowing pressurization of the inlet, and the software has to be updated to control the directions of oil flow. An indirect hydraulic energy recovery system consisting of two controllable hydraulic machines and a hydraulic accumulator could also be implemented, but its behaviour was not studied here. However, we anticipate that it might have similar capabilities as the electric energy control system evaluated in this paper. 5 CONCLUSIONS The presented work concentrated on analysing the possibilities of using energy regeneration in electrohydraulic forklift systems. The measurements showed that energy recovery from potential energy is possible in both hydraulic and electric energy storage applications. According to the results, the maximum energy-savings ratio for the free lift zone with optimized hydraulic accumulator parameters was 45% using the direct hydraulic recovery setup. In practice, however, the direct system suffers from the need to control the pre-load pressure of the hydraulic accumulator or the requirement to select a fixed value for the pre-load pressure, and, as a result, significantly lower average values may be obtained. In this test, the best energy-savings ratio of the electric recovery setup was 36%. This result is disappointingly low compared

Minav, T. – Hänninen, H. – Sinkkonen, A. – Laurila, L. – Pyrhönen, J.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 232-240

with the previous results, obtained when operating in the second cylinder zone of the same truck [19]. It can be concluded that the test arrangement favoured the direct hydraulic recovery system, but it also shows that the electric drive system has numerous advantages. The latter does not require any pre-load settings and tunings of the energy storage for a specific load or lifting height. Therefore, the direct hydraulic recovery approach is impractical in cases where the lifting and lowering range and the mass vary. In such cases, an electric or indirect hydraulic energy recovery system should be considered instead. 6 ACKNOWLEDGEMENTS Hydraulic recovery studies; This study is connected to the MIDE/HybLab project, funded by Aalto University. The cooperation of Jyri Juhala, M.Sc. at the Department of Engineering Design and Production of the School of Engineering of Aalto University is highly appreciated. Electric recovery studies; The research was enabled by the financial support of Tekes, the Finnish Funding Agency for Technology and Innovation and FIMA (Forum for Intelligent Machines), the European Union, the European Regional Development Fund, and the Regional Council of South Karelia. The research was carried out at the Institute of Energy Technology, Department of Electrical Engineering, Lappeenranta University of Technology, Lappeenranta, Finland. 7 REFERENCES [1] Liang, X. (2002). On Improving Energy Utilization in Hydraulic Booms. PhD thesis, Espoo. [2] Lin, T., Wang, Q., Hu, B., Gong, W. (2010). Development of hybrid powered hydraulic construction machinery. Journal Automation in Construction, vol. 19, no. 1, p. 11-19, DOI:10.1016/j.autcon.2009.09.005. [3] Nyman, J., Rydberg, K. (2001). Energy saving lifting hydraulic systems. Proceedings of the 7th Scandinavian International Conference on Fluid Power, Linköping. [4] Nyman, J., Bärnström, J., Rydberg, K. (2003). Use of accumulators to reduce the need of electric power in hydraulic lifting systems. Proceedings of the 8th Scandinavian International Conference on Fluid Power, Tampere. [5] Ming, X., Bo, J., Guojin, C., Jing, N. (2013). Speed-control of energy regulation based

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[16] Sinkkonen, A., Kauranne, H., Hänninen, H., Pietola, M. (2011). Analysis of energy balance in electrohydraulic forklift. Proceedings of the 12th Scandinavian International Conference on Fluid Power, Tampere. [17] Juhala, J., Kauranne, H., Kajaste, J., Pietola, M. (2009). Improving energy efficiency of work machine with digital hydraulics and pressure accumulator. Proceedings of the 11th Scandinavian International Conference on Fluid Power, Linköping. [18] Minav, T. Immonen, P., Laurila, L., Vtorov, V., Pyrhönen, J., Niemelä, M. (2011). Electric energy recovery system for a hydraulic forklift - theoretical and experimental evaluation. IET Electric Power Applications, vol. 5, no. 4, p. 377-385, DOI:10.1049/iet-epa.2009.0302. [19] Charging Lead Acid. (2012). from http:// batteryuniversity.com/learn/article/charging_ the_lead_acid_battery, accessed on 10.01.2012. [20] Minav, T. (2011). Electric-Drive-Based Control and Electric Energy Regeneration in

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a Hydraulic System. PhD thesis, Lappeenranta University of Technology, Lappeenranta. [21] Linjama, M., Huova, M., & Karvonen, M. (2012). Modelling of flow characteristics of on/ off valves. The 5th Workshop on Digital Fluid Power, Tampere. [22] Hänninen, H., Kajaste, J., Pietola, M. (2012). Optimizing hydraulic energy recovery system of reach truck. Fluid Power and Motion Control, Bath, p.109-121. [23] Minav, T., Laurila, L., Pyrhönen, J., Vtorov, V. (2011). Direct pump control effects on the energy efficiency in an electro-hydraulic lifting system. Journal International Review of Automatic Control, vol. 4, no. 2, p. 235-242. [24] Minav, T., Laurila, L., Pyrhönen, J. (2012). Permanent magnet synchronous machine sizing: effect on the energy efficiency of an electro-hydraulic forklift. IEEE Transactions on Industrial Electronics, vol. 59, no. 6, p. 24662474, DOI:10.1109/TIE.2011.214868.

Minav, T. – Hänninen, H. – Sinkkonen, A. – Laurila, L. – Pyrhönen, J.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 241-249 © 2014 Journal of Mechanical Engineering. All rights reserved. DOI:10.5545/sv-jme.2013.1348

Received for review: 2013-07-31 Received revised form: 2013-11-15 Accepted for publication: 2013-12-11

Original Scientific Paper

Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters Bešter, T. – Fajdiga, M. – Nagode, M. Tomaž Bešter* – Matija Fajdiga – Marko Nagode

University of Ljubljana, Faculty of Mechanical Engineering, Slovenia Air spring manufactures use constant amplitude tests for the quality validation of air springs. The tests are very simple and the only information we get from them is that a spring is adequate if it passes the test and inadequate if it does not. One of the objectives of this article is to use these tests to make life predictions based on the standardised load spectrum. This prediction is made with force as the damage parameter. The second objective is to determine if it is possible to use experimental results obtained at one control parameter, e.g. force, to make life predictions for another control parameter, e.g. stress. With equations it is proved that such transformation is possible. Keywords: vehicle suspension, air spring, load spectrum, dynamic tests, fatigue life

0 INTRODUCTION

1 STANDARD LOAD SPECTRUM

Air spring assembly consists of a piston, bellows, a bed plate and a bumper (Fig. 1). Pressure inside the bellows and the piston shape determines air spring characteristics. An air spring can have progressive spring characteristic, which is most suitable for transport vehicles that are loaded with various loads during exploitation [1].

For transport vehicles, standard load spectra were determined [2] and [3] which define dynamic wheel force Fz,dyn on transport vehicles. Standard load spectrum has 1.5×108 load cycles, which corresponds to 500000 km driving distance with 300 load cycles per kilometre. Standardised load spectra have three driving modes: straight driving, cornering and braking. Based on wheel force measurements during exploitation, standard load spectrum was determined with the level crossing method. Dynamic force Fz,dyn in standardized load spectra is expressed with dynamic load factor nz which represents ratio between dynamic and static load nz = Fz,dyn / Fz,sta. Dynamic load factor nz depends on number of load cycles N and driving mode (Table 1, Fig. 2). For arbitrary static load Fz,sta dynamic load Fz,dyn can be determined with following equation: Fz ,dyn = nz ⋅ Fz ,sta . (1) Load ratio:

Fig. 1. Air spring

Air spring manufacturers have been testing springs with various static and dynamic experiments in order to verify spring quality. Dynamic tests usually have a constant amplitude and require a certain number of load cycles without critical damage on the air spring. If the spring successfully endures the test, critical damage on the spring does not occur and hence test results do not give exact information about fatigue life. In air spring fatigue tests, critical damage usually occurs on air spring bellows. In this article, the possibilities of obtaining SN-curves with modified constant amplitude tests will be examined. To make fatigue life predictions, loads on air springs and appropriate SN-curves must also be determined.

RF =

Fmin , (2) Fmax

where Fmin is minimal force, and Fmax is maximal force. During exploitation this ratio is changing. Load ratio range for all driving modes has been presented in Table 2. When deformations are small and materials have linear characteristic, load ratio is equal to stress ratio:

R = RF = Rσ =

σ min , (3) σ max

where σmin and σmax are minimum and maximum stress. For high cycle fatigue, life calculation stress is

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

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usually used as the control parameter. When load ratio equals stress ratio, force can be used as the control parameter as well. On air springs, large deformations occur during exploitation and air spring bellows are made of polymer materials with nonlinear stress strain diagram [4] to [7]. Due to the geometric and material nonlinearities, Eq. (3) is not valid for air springs in general. This means fatigue life calculations with different control parameters will have different results. To compare fatigue life calculations with different parameters, SN-curves and Goodman diagrams would have to be determined for all control parameters in question. At least four experiments must be performed to design one Goodman diagram (Fig. 3). In this article, it will be shown how the Goodman diagram for one control parameter could be converted into the Goodman diagram for another control parameter that is not linearly dependent from the former parameter.

2 DYNAMIC EXPERIMENTS AND FATIGUE LIFE In the standardized load spectrum, load forces on wheels are defined. Wheel forces are transferred to air springs by suspension mechanism. As air spring forces are easily calculated, the easiest way to calculate fatigue life is to use force as a control parameter for the determination of SN-curves and the Goodman diagram. As load ratios of the standard load spectrum are in the range between 0 and 1, experiments for the determination of SN-curves and Goodman diagram must be in the same range. For constant amplitude loads Fa1R1, Fa2R1, Fa1R2 and Fa2R2 the number of load cycles when critical damage occurs: N1R1, N2R1, N1R2 and N2R2 must be experimentally determined (Fig. 3). Those results allow to calculate SN-curves slope:

k R1 =

log N a1R1 − log N a 2 R1 , (4) log Fa1R1 − log Fa 2 R1

kR 2 =

log N a1R 2 − log N a 2 R 2 . (5) log Fa1R 2 − log Fa 2 R 2

Table 1. Standard load spectra [2] Driving mode

Dynamic load factor nz

Load spectrum parameters Hg = 1.5×108 H He SHAPE Linear 0.96·Hg H·10-6 distribution

Straight driving Cornering Outer wheel Inner wheel

1.5 0.4

0.04·Hg

50

Breaking

2.0

5·105

104

2.0

Normal distribution Normal distribution

If fatigue limit is assumed to be at ND = 2×106, amplitude fatigue strength can be calculated for both dynamic factors: 1

 N  kR1 FaDR1 = Fa1R1  D  , (6)  N1R1 

 N  kR 2 = Fa1R 2  D  . (7)  N1R 2 

1

Straight driving Cornering Breaking Total

2.0

nz

1.5

FaDR 2

1.0

From amplitude forces and load ratio, medium forces are calculated (for development of Eqs. (8) and (9) see the appendix):

0.5

0.0

101

102

103

104

N

105

106

Fig. 2. Standard load spectra Table 2. Load ratio ranges for various driving modes Driving modes Straight driving Cornering Breaking

242

RF 0 to 1 0.27 to 1 0.5 to 1

107

108

FmDR1 =

FaDR1 (1 + R1 )

, (8)

(1 − R1 ) F (1 + R2 ) , (9) FmDR 2 = aDR 2 (1 − R2 )

and the gradient of the Goodman diagram:

M=

FaDR 2 − FaDR1 . (10) FmDR 2 − FmDR1

When the gradient of the Goodman diagram is known, equivalent amplitude load Fa1R1 with load ratio R1 can be calculated (Fig. 4, for development of the Eq. (11) see the appendix): Bešter, T. – Fajdiga, M. – Nagode, M.


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Fig. 3. Design of Goodman diagram

Fa1R1 = Fa1Ri

 1 + M   1 + M 

( Ri + 1)   ( Ri − 1)  . (11) ( R1 + 1)   ( R1 − 1) 

Fig. 4. Equivalent loads based on Goodman diagram [8]

Air spring manufacturers use various constant amplitude dynamic tests to validate quality and endurance of their air springs [9] and [10] . Two durability tests have been used in our research. In the first durability test, an air spring is loaded with the displacement amplitude ±50 mm, at the frequency 3.3 Hz. The air spring must endure 2×106 load cycles without critical damage. In the second durability test, the displacement amplitude is ±75 mm at the frequency 2 Hz, with the air spring having to endure 106 load cycles without critical damage. Both tests are stopped when the required number of load cycles is reached unless critical damage occurs first. If springs are tested until the critical damage occurs, two points on the SN-curve can be obtained. Air spring

characteristic was measured (Fig. 5) to determine the appropriate force amplitude for any displacement amplitude.

Fig. 5. Air spring characteristic

In the first durability test, the minimum required number of load cycles is the same as the fatigue limit. In the second durability test the minimum required number of load cycles is relatively near the fatigue limit. It is reasonable to use tests with higher amplitudes for determination of the remaining SNcurve points in order to reduce test time. With two modified existing and two additional constant amplitude tests, two SN-curves and a Goodman diagram can be determined. This enables the transformation of any load to an equivalent load with the load ratio RF of a known SN-curve. An SN-curve can be approximately determined with the tensile strength force FM and one dynamic test in the high cycle fatigue range because SN-curves with various load factors intersect near the tensile strength

Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters

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spectrum (Table 3). The results show significant differences between accumulation rules. When material has stress in the linear region of the stress strain diagram and deformations are small, the stress ratio is equal to the load ratio. Air springs are made of material with nonlinear stress strain diagram and are submitted to large deformations in most load cycles of the standard load spectrum. To evaluate the significance of those nonlinearities, life predictions with stress as the damage parameter should be made. Table 3. Fatigue life

Fig. 6. Approximate SN-curves [11]

 

force (Fig. 6). An air spring’s maximum displacement is limited, therefore it is not possible to load a spring with a high enough load to cause the bellows to burst. The air spring bellows tensile strength cannot be directly determined, but it is possible to evaluate the tensile strength if pressure is increased in the spring’s bellows until it bursts and the maximum force is used as the tensile strength [11]. Once SN-curves are determined, fatigue life can be calculated with one of the damage accumulation rules [8] and [11]. The most frequently used linear accumulation rules are the original and elementary Palmgren-Miner rule and the Haibach rule (Fig. 7). In all linear accumulation rules, damage di is equal to the inverse number of load cycles when critical damage occurs. Total damage is:

n

d =∑ i =1

1 . (12) N iR

In the described manner, damage was calculated for an air spring loaded with standardized load

Numumber of load cycles Relative number of load cycles

Basic PalmgrenMiner rule

Elementary PalmgrenMiner rule

Haibach rule

9020073

3670064

5155576

1

0.40688

0.57157

3 LIFE PREDICTION WHEN DYNAMIC EXPERIMETS HAVE DIFFERENT LOAD RATIOS 3.1 Determination of the Goodman Diagram from Four Dynamic Experiments with Different Load Ratios A pair of experiments with equal load ratios when force is used as the control parameter (Fig. 3) will not have equal stress ratios (Fig. 8) due to the nonlinear relation between force and stress. It is not possible to determine SN-curves and the Goodman diagram directly with four experiments that have four different stress ratios. It will be shown that it is possible to solve the problem with a system of equations. The solution of this system can give all the data needed to determine the SN-curves and the Goodman diagram

Fig. 7. SN-curves for original Palmgren-Miner, elementary Palmgren-Miner and Haibach rule [8]

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Fig. 8. Determination of the Goodman diagram from four dynamic tests

even if only four experiments with different stress ratios are available. Goodman diagram gradient M depends on two fatigue limit medium and amplitude stresses σaDR11, σaDR21, σmDR11 and σmDR21 (Table 4). To determine those stresses, two SN-curve gradients kR11 and kR21 are needed. To determine those gradients, it is necessary to obtain equivalent stresses with stress ratios R11 and R21: σm2R11, σm2R21, σa2R11 and σa2R21. To calculate equivalent stresses the gradient of the Goodman diagram M is needed. There are eleven unknowns. Eleven equations can be set, hence the unknowns can be determined by the equation system solution:

1

 N  kR11 σ aD , R11 = σ a1R11  D  , (19)  N1R11 

 N  kR 21 σ aDR 21 = σ a1R 21  D  , (20)  N1R 21 

1

σ mDR11 =

σ aDR11 (1 + R11 ) , (21) (1 − R11 )

σ mDR 21 =

σ aDR 21 (1 + R21 ) , (22) (1 − R21 )

M=

σ aDR 21 − σ aDR11 . (23) σ mDR 21 − σ mDR11

σ m 2 R11 = −

(σ a 2 R12 − M σ m 2 R12 ) ⋅ ( R11 + 1) , (13) ( ( R11 − 1) + M ( R11 + 1) )

σ m 2 R 21 = −

(σ a 2 R 22 − M σ m 2 R 22 ) ⋅ ( R21 + 1) , (14) ( ( R21 − 1) + M ( R21 + 1) )

It is difficult to determine stresses on air spring bellows because large deformations and material nonlinearity must be considered. Despite these difficulties, precise enough stress-strain analyses were made [12] to [14]. As stresses in air spring bellows can be determined, known quantities in the system of equations are medium and amplitude stresses (Table 4). With part of equations being nonlinear, the system of equations does not have a simple analytical solution, but it can be solved numerically.

σ a 2 R11 =

σ m 2 R11 (1 − R11 ) , (15) ( R11 + 1)

σ a 2 R 21 =

σ m1R 21 (1 − R21 ) , (16) ( R21 + 1)

k R11 =

log N1R11 − log N 2 R11 , (17) log σ a1R11 − log σ a 2 R11

k R 21 =

log N1R 21 − log N 2 R 21 , (18) log σ a1R 21 − log σ a 2 R 21 Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters

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Table 4. Determination of the Goodman diagram from four dynamic tests– used quantities

σa1R11 [MPa] σa2R12 [MPa] σa1R21 [MPa] σa2R22 [MPa] σm1R11 [MPa] σm2R12 [MPa] σm1R21 [MPa] σm2R22 [MPa]

Known quantities Amplitude stress at Fa1R1 [N] and Fm1R1 [N]

Unknown quantities Amplitude fatigue strength at R11

Amplitude stress at Fa2R2 [N] and Fm2R2 [N]

σaDR11 [MPa] σaDR21 [MPa] σmDR11 [MPa] σmDR21 [MPa]

Medium stress at Fm1R1 [N] and Fa1R1 [N]

kR11

Medium stress at σaDR21 [MPa] SN-curve gradient at R11

Medium stress at Fm2R1 [N] and Fa2R1 [N]

kR21

SN-curve gradient at R21

Medium stress at Fm1R2 [N] and Fa1R2 [N]

σa2R11 [MPa] σa2R21 [MPa] σm2R11 [MPa] σm2R21 [MPa]

Equivalent amplitude stress for σa2R12 [MPa] at R11

Amplitude stress at Fa2R1 [N] and Fm2R1 [N] Amplitude stress at Fa2R1 [N] and Fm2R1 [N]

Medium stress at Fm2R2 [N] and Fa2R2 [N]

R11

Stress ratio at Fa1R1 [N] and Fm1R1 [N]

R12

Stress ratio at Fa2R1 [N] and Fm2R1 [N]

R21 R22 N1R11 N2R12 N1R21 N2R22 ND

Stress ratio at Fa2R1 [N] and Fm2R1 [N] M Stress ratio at Fa2R2 [N] and Fm2R2 [N] Number of load cycles when critical damage occurs at Fa1R1 [N] and Fm1R1 [N] Number of load cycles when critical damage occurs at Fa1R1 [N] and Fm1R1 [N] Number of load cycles when critical damage occurs at Fa1R1 [N] and Fm1R1 [N] Number of load cycles when critical damage occurs at Fa1R1 [N] and Fm1R1 [N] Fatigue limit

3.2 Determination of the Goodman Diagram from Three Dynamic Experiments with Different Load Ratios Although it is possible to determine the Goodman diagram from four dynamic tests with different dynamic factors, it is easier to use just three dynamic tests, but we have to omit the calculation in the first step to determine σaD1, kR1 and M. Equivalent stress σa12, σm12 and σa13, σm13 (Fig. 9, Table 5) can be calculated using following equations: 1

σ aD1  N D  k1 =  , (24) σ a1  N1 

σ aD1  N D  k1 =  , (25) σ a12  N 2 

σ aD1  N D  k1 =  . (26) σ a13  N 3 

Amplitude fatigue strength at RR21 Medium stress at σaDR11 [MPa]

Equivalent amplitude stress for σa2R22 [MPa] at R21

Equivalent medium stress for σm2R12 [MPa] at R11

Equivalent medium stress for σm2R22 [MPa] at R21 Goodman diagram gradient

σ m12 =

σ a12 (1 + R1 ) , (29) (1 − R1 )

σ m13 =

σ a13 (1 + R1 ) . (30) (1 − R1 )

If Eqs. (29) and (30) are inserted in Eqs. (27) and (28), equations for σa12 and σa13 can be written using σa2, σa3, σm2, σm3, R1 and M, where M is the only unknown:

σ a12 =

(σ a 2 − M σ m 2 ) (1 − R1 ) , (31) (1 − R1 ) − M (1 + R1 )

σ a13 =

(σ a 3 − M σ m3 ) (1 − R1 ) . (32) (1 − R1 ) − M (1 + R1 )

1

1

As σa12 and σa13 are not known in Eqs. (25) and (26), they must be expressed with known quantities. Based on the Goodman diagram equations can be written.

σ a12 − σ a 2 = M (σ m12 − σ m 2 ) , (27)

σ a13 − σ a 3 = M (σ m13 − σ m 3 ) . (28) Medium stress can be written:

246

If logarithms from Eqs. (24), to (26) are calculated and Eqs. (31) and (32) are used for σa12 and σa13, a system of three equations and three unknowns is obtained:

log σ aD1 − log σ a1 =

1 ( log N D − log N1 ) , (33) k1

 (σ − M σ m 2 ) (1 − R1 )  log σ aD1 − log  a 2 =  (1 − R ) − M (1 + R )  1 1   1 = ( log N D − log N 2 ) , (34) k1

Bešter, T. – Fajdiga, M. – Nagode, M.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 241-249

Fig. 9. Determination of the Goodman diagram from three dynamic tests Table 5. Determination of the Goodman diagram from three dynamic tests – used quantities

σa1 [MPa] σa2 [MPa] σa3 [MPa] σm1 [MPa] σm2 [MPa] σm3 [MPa] R1 R2 R3 N1 N2 N3 ND

Known quantities Amplitude stress with stress ratio R1

k1

Amplitude stress with stress ratio R2

M

Goodman diagram gradient

Amplitude stress with stress ratio R3

σaD1 [MPa] σmD1 [MPa] σa12 [MPa] σa13 [MPa] σm12 [MPa] σm13 [MPa]

Amplitude fatigue strength at R1

Medium stress with stress ratio R1 Medium stress with stress ratio R2 Medium stress with stress ratio R3 Stress ratio at σa1 and σm1 Stress ratio at σa2 and σm2 Stress ratio at σa3 and σm3

Amplitude fatigue strength at R1 Equivalent amplitude stress for σa2 at R1 Equivalent amplitude stress for σa3 at R1

Equivalent medium stress for σm2 at R1 Equivalent medium stress for σm3 at R1

Number of load cycles when critical damage occurs at σa1 and σm1 Number of load cycles when critical damage occurs at σa2 and σm2 Number of load cycles when critical damage occurs at σa3 and σm3 Fatigue limit

 (σ − M σ m 3 ) (1 − R1 )  log σ aD1 − log  a 3 =  (1 − R ) − M (1 + R )  1 1   1 (35) = ( log N D − log N 3 ) . k1 From Eq. (33) logσaD1 can be calculated:

Unknown quantities SN-curve gradient at R1

log σ aD1 = log σ a1 +

1 ( log N D − log N1 ) , (36) k1

and inserted into Eqs. (34) and (35). Thus two equations with two unknowns are obtained:

N1 1  log k1  N2

 (σ a 2 − M σ m 2 ) (1 − R1 )    = log σ a1 , (37)  + log    (1 − R1 ) − M (1 + R1 ) 

 (σ a 3 − M σ m 3 ) (1 − R1 )  N1  1  = log σ a1. (38)  log  + log  k1  N3   (1 − R1 ) − M (1 + R1 ) 

There is no simple analytical solution for Eqs. (37) and (38) but it is possible to calculate k1 and M numerically. When k1 is calculated, fatigue limit σaD1 can be calculated with Eq. (36). With the Goodman diagram gradient M, any stress σa1Ri with arbitrary stress ratio Ri can be transformed to equivalent stress σa1R1 with stress ratio R1 and known SN-curve gradient k1:

Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters

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σ a1R1 = σ a1Ri

 1 + M   1 + M 

( Ri + 1)   ( Ri − 1)  . (39) ( R1 + 1)   ( R1 − 1) 

This enables to calculate equivalent stress for any load, e.g. σa12, σa13, etc. (Fig. 9). 4 CONCLUSION

6 APPENDIX Development of the medium force equation (Eqs. (8) and (9). Analogue equations are used for medium stress (Eqs. (29) and (30)). FD max – maximum force FD min – minimum force FmDRi – medium force with load ratio Ri FaDRi – amplitude force with load ratio Ri Ri =

Unchanged durability tests cannot be used to make fatigue life predictions, because unchanged tests are stopped when the required number of load cycles is reached. If those tests are performed until critical damage is reached and additional constant amplitude tests are made, two SN-curves, Goodman diagram and fatigue life prediction can be made. Fatigue life prediction was made with force as damage parameter using three damage accumulation rules. Various damage accumulation methods gave significantly different results. The elementary Palmgren-Miner rule and the Haibach modification take into account even loads that are smaller than the fatigue limit therefore those rules give shorter fatigue life predictions than the basic Palmgren-Miner rule. No tests were made with loads smaller than the fatigue limit therefore no estimation about the accuracy of damage accumulation rules in this range can be made. Due to geometric and material nonlinearities, fatigue life calculations with force as the control parameter may be significantly different than fatigue life calculations with stress as the control parameter. The comparison of life predictions using different control parameters would be very time consuming and expensive as we would have to experimentally determine Goodman diagrams for all control parameters in question. With the equations we proved it is possible to transform the Goodman diagram for one control parameter, e.g. force, into the Goodman diagram for another control parameter, e.g. stress. The transformation of the Goodman diagram to arbitrary control parameters makes further research of the influence of control parameters on fatigue life much quicker and less expensive.

FD min FmDRi − FaDRi = , FD max FmDRi + FaDRi

Ri ⋅ FmDRi + Ri ⋅ FaDRi = FmDRi − FaDRi ,

Ri ⋅ FmDRi − FmDRi = − Ri ⋅ FaDRi − FaDRi ,

FmDRi ( Ri − 1) = − FaDRi ( Ri + 1) ,

( Ri + 1) = F (1 + Ri ) . ( Ri − 1) aDRi (1 − Ri )

FmDRi = − FaDRi

Development of the equivalent amplitude force equation (Eq (11), Fig. 4). Analogue equation is used for equivalent amplitude stress (Eq. (39)). Fa1Ri – amplitude load with load ratio Ri Fa1R1 - equivalent amplitude load with load ratio R1 Fm1Ri – medium load with load ratio Ri Fm1R1 - equivalent medium load with load ratio R1

M=

Fa1R1 − Fa1Ri , Fm1R1 − Fm1Ri

Fa1R1 − Fa1Ri = M ( Fm1R1 − Fm1Ri ) ,

 ( R + 1) − F  , Fa1R1 − Fa1Ri = M  − Fa1R1 1   ( R1 − 1) m1Ri  

Fa1R1 + M ⋅ Fa1R1

( R1 + 1) = F − M ⋅ F , m1Ri ( R1 − 1) a1Ri

 ( R + 1)  = F + M ⋅ F ( Ri + 1) , Fa1R1 1 + M 1  a1Ri  ( R1 − 1)  a1Ri ( Ri − 1) 

 1 + M   1 + M 

5 ACKNOWLEDGMENT The authors appreciate the support provided by the Veyance Technologies Europe, d.o.o.

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Bešter, T. – Fajdiga, M. – Nagode, M.

Fa1R1 = Fa1Ri

( Ri + 1)   ( Ri − 1)  . ( R1 + 1)   ( R1 − 1) 


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7 REFERENCES [1] Pirnat, M., Savšek, Z., Boltežar, M. (2011). Measuring dynamic loads on a foldable city bicycle. Strojniški vestnik - Journal of Mechanical Engineering, vol. 57, no. 1, p. 21-26, DOI:10.5545/sv-jme.2009.149.21-26, 2011. [2] Neugebauer, J.R., Grubisic, V., Fischer, G. (1989). Procedure For Design Optimization And Durability Life Approval Of Truck Axles And Axle Assemblies. SAE Technical Paper, 892535. [3] Heuler, P., Klätschke, H. (2005). Generation and use of standardised load spectra and load–time histories. International Journal of Fatigue, vol. 27, no. 8, p. 974990, DOI:10.1016/j.ijfatigue.2004.09.012. [4] Oman, S., Nagode, M., Fajdiga, M. (2009).The material characterization of the air spring bellow sealing layer. Materials & Design, vol. 30, no. 4, p. 1141-1150, DOI:10.1016/j.matdes.2008.06.035. [5] Shaw, M.T., MacKnight, W.J. (2005). Introduction to Polymer Viscoelasticity, 3rd ed. John Wiley & Sons, Hoboken, DOI:10.1002/0471741833. [6] Tschoegl, N.W., Knauss, W.G., Emri, I. (2002). The effect of temperature and pressure on the mechanical properties of thermo- and/or piezorheologically simple polymeric materials in thermodynamic

equilibrium – A critical review. Mechanics of Time-Dependent Materials, vol. 6, no. 1, p. 53-99, DOI:10.1023/A:1014421519100. [7] Gent, A. (2012). Engineering With Rubber. Carl Hanser Verlag, Munich, DOI:10.3139/9783446428713. [8]  LMS Falancs Theory Manual version 2.9. (2000). LMS Durability Technologies, Leuven. [9]  Air Spring - Instaltion Requirement (2008). Mack Trucks, Allentown. [10]  Durability-Flex Life (2009). Goodyear Tire and Rubber Company, Green. [11] Ellyin, F. (1997). Fatigue Damage, Crack Growth And Life Predicition. Chapman&Hall, London. [12] Oman, S., Fajdiga, M., Nagode, M. (2010). Estimation of air-spring life based on accelerated experiments. Materials and Design, vol. 31, no. 8, p. 3859-3868, DOI:10.1016/j.matdes.2010.03.044. [13] Hackenschmidt, R., Alber-Laukant, B., Rieg, F. (2011). Simulating nonlinear materials under centrifugal forces. Strojniški vestnik - Journal of Mechanical Engineering, vol. 57, no. 7-8, p. 531-538, DOI:10.5545/ sv-jme.2011.013. [14] Oman, S., Nagode, M. (2013). On the influence of the cord angle on air-spring fatigue life. Engineering Failure Analysis, vol. 27, p. 61-73, DOI:10.1016/j. engfailanal.2012.09.002.

Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters

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Original Scientific Paper

Received for review: 2013-08-01 Received revised form: 2013-12-06 Accepted for publication: 2013-12-11

Determination of Fatigue Crack Growth Trajectory and Residual Life under Mixed Modes

Blažić, M. – Maksimović, S. – Petrović, Z. – Vasović, I. – Turnić, D. Marija Blažić1,* – Stevan Maksimović1 – Zlatko Petrović3 – Ivana Vasović2 – Dragana Turnić4 1 Military

Technical Institute, Serbia Goša, Serbia 3 University of Belgrade, Faculty of Mechanical Engineering, Serbia 4 University of Belgrade, Faculty of Civil Engineering and Architecture, Serbia 2 Institute

This paper considers the determination of the crack growth trajectory and residual life for the two-dimensional structural elements under mixed modes. To study crack growth behaviour, a specimen with two holes and a crack between them is considered. This crack is defined as achieving a mixed-mode I/II crack growth trajectory. To produce a crack growth trajectory under cyclic loads with an initial crack between the two holes, MTS servo-hydraulic system is used. Crack growth trajectory is defined using numerical simulations via finite elements. The results of the numerical simulations by finite elements are compared with experimental results. Residual life along the ‘curve’ mixed-mode crack growth trajectory is determined numerically and experimentally. The crack growth trajectory obtained via the presented numerical simulation and residual life are compared with own experimental results. Keywords: fatigue crack growth, mixed modes, residual life estimation, finite element simulation

0 INTRODUCTION The design considerations of aircraft structures based on damage tolerance approach often require the prediction of mixed-mode fatigue crack growth. In this approach, the propagation path of a crack in a part is an essential aspect for the fatigue life simulation using the methodology of fracture mechanics. However, most of the existing approaches are limited to the mode-I fatigue crack growth cases (e.g. [1] to [4]). These approaches are generally based on correlations between the fatigue crack growth rate (da/dN) and the range of the mode-I stress intensity factor (ΔKI). The commonly used fatigue crack growth rate equation is [5]:

da dN

n = C ( ∆K ) , (1)

involving the experimentally determined constants C and n may not be adequate, because they are restricted to cracks running in a straight line. In the damage tolerance approach, the propagation path of a crack in a part is an essential aspect for fatigue life simulation using fracture mechanics methodology. For these cases, the cracks do not propagate in the direction normal to the applied load; these models need the stress intensity factor history along the crack path. Interesting attempts to predict the angle of crack propagation, as well as the fatigue crack growth rate for mixed-mode cracks, are divided into two categories. The first incorporates the methodologies that consider the stress or the strain as the fatigue crack growth driving force, e.g. the maximum tangential 250

stress (MTS) criterion [6] and [7], the tangential stress factor and tangential strain factor [8], the maximum tangential strain criterion [9], etc. The second category contains the methodologies that recognize the material strain energy density as the fatigue crack growth driving force, e.g. the minimum strain energy density (S) criterion [10] to [12], the dilatational strain energy density (T) criterion [13] and [14], etc. In previously mentioned works, the distribution of the total or the dilatational elastic strain energy density around the crack tip is evaluated along a circular core or the elastic-plastic boundary region before the crack extension. It is postulated that the mixed-mode fatigue crack propagates along a direction defined by a minimum for the total strain energy density [10] to [12] or by a maximum for the dilatational component of the strain energy density [13] and [14]. However, these postulations are mostly based on hypothetical approaches [15]. The accuracy of their predictions depends on several parameters, including the material ductility [13], load mixities [15], etc. In this paper, the development of a method supported by a better physical basis is attempted. To this scope, the tendency of the elastic stress field to minimize the accumulated elastic strain energy (e.g. [9] to [15]) (not the energy density) is taken into account. The proposed methodology differs from the previous methodologies in the following points: a) The factor controlling the mixed-mode crack propagation is the accumulated energy, while in the above works [10] to [16] it is the accumulated energy density; and b) The criterion for the prediction of the path of the mixed-mode fatigue crack propagation is

*Corr. Author’s Address: Military Technical Institute, Ratka Resanovića 1, 11000 Belgrade, Serbia, vti@vti.vs.rs


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 250-254

the value of the accumulated elastic strain energy after the crack propagation, which incorporates the resulted new stress distribution due to the crack increment. In contrast in the aforementioned works of [10] to [15], the criterion for the crack path prediction is the energy density before the crack growth. In order to verify the computation procedure shown in this work, experimental tests have been undertaken. The main scientific contribution of this paper is the developed computation method for residual fatigue life estimation along to curve mixed-mode crack growth trajectory that is verified with own experiments. 1 DETERMINATION OF THE CRACK GROWTH TRAJECTORY With the stress and strain fields around the crack-tip, fracture parameters for mixed-mode problems are calculated to predict the crack propagation path of the plate with crack. For this purpose, fracture parameters such as KI, KII are used. Having the fracture parameters, a criterion is needed to predict the crack growth direction in a mixed-mode problem. Several criteria have already been proposed for this purpose. Previous research [15] and [16] shows that there are no significant differences between the obtained crack trajectories based on various crack propagation criteria. Using stress as a parameter, the (MTS) criterion was presented by Erdogan and Sih [6]. This criterion states that a crack propagates in a direction corresponding to the direction of maximum tangential stress along a constant radius around the crack-tip. Using the Westergaurd stress field in the polar co-ordinates and applying the (MTS) criterion, Eq. (2) is obtained to predict the crack propagation direction in each incremental step [16]. The fracture toughness for a brittle material is usually measured in a pure mode-I loading conditions, noted by KIC. For a general mixed-mode case, we need a criterion to determine the angle of incipient propagation with respect to crack direction, and a critical combination of stress intensity factors that lead to crack propagation. Various criteria have been proposed by researchers of mixed-mode crack propagation, including the maximum energy release rate, the minimum strain energy density criteria, the maximum circumferential tensile stress, etc. The maximum energy release rate was demonstrated by Erdogan and Sih [6] by assuming the Griffith theory as a valid criterion for crack growth. Based on this theory, the crack propagates in the direction for which the elastic energy release rate per

unit crack extension becomes maximal. In this case, the crack begins to grow when the energy release reaches a critical value [6]. The minimum strain energy density theory, proposed by Sih [6], postulates that a crack propagates when the strain energy density at a critical distance reaches a minimum value. The numerical implementation of this theory can be seen in [9] to [15]. The maximum circumferential tensile stress theory was presented by Erdogan and Sih [6] based on the state of stress near the crack tip. Based on the maximum circumferential tensile stress, the hoop stress reaches its maximum value on the plane of zero shear stress. Assuming that the size of plastic zone at the crack tip is negligible, we can use the singular term solutions of stress at the crack tip to determine the crack propagation angle, where the shear stress becomes zero. The crack propagation angle θ0 can be expressed by using the angle between the line of crack and the crack growth direction, with the positive value defined in the anti-clockwise direction, as:  K Θo = 2 tan -1  I  4 K II   K Θo = 2 tan -1  I  4 K II 

2  1  KI   for K 〉 0, + 8   II  4  K II   2  1  KI   for K 〈 0. (2) + + 8   II  4  K II  

-

To initiate crack propagation, the maximum circumferential tensile stress σ must reach critical value. This results in an expression for the equivalent stress intensity factor SIF in mixed-mode condition as:

K eq = K I cos3

Θ0 3 Θ - K II cos 0 sin Θ0 . (3) 2 2 2

However, when the plastic zone size cannot be ignored, it is necessary to use the stress state at a material-dependent finite distance from the crack tip. 2 COMPUTATION AND EXPERIMENTAL RESULTS In this work, two types of problems are considered: i) determination of the crack growth trajectory and ii) estimation of the residual life along the ‘curve’ mixed-mode crack growth trajectory. 2.1 Crack Growth Trajectory To illustrate determination of the crack growth trajectory under mixed modes I/II, a duraluminum plate with two holes and an initial crack under

Determination of Fatigue Crack Growth Trajectory and Residual Life under Mixed Modes

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tension load Fy are considered, as shown in Fig. 1. To determine stress intensity factors KI and KII, the Msc/ Nastran software code [17] is used here. In Fig. 2, the finite element model with stress distributions of the cracked specimen is shown.

To validate the computation procedure for determination of the crack growth trajectory, an experimental test is included; it was carried out using a servo-hydraulic MTS system, Figs. 4 and 5.

Fig. 1. Geometry of specimen for modelling of crack growth trajectory

Fig. 4. Specimen in servo hydraulic MTS system

Fig. 2. Stress distributions of cracked specimen using finite elements (Fy = 60000 N)

To predict the crack growth direction in a mixedmode problem, in this analysis, the MTS criterion [15] and [16] is used in combining Msc/Nastran code [17]. Combining finite elements for determination of the stress intensity factors and MTS criterion, the computation crack trajectory is obtained, Fig. 3.

Fig. 5. Experimentally determined of crack growth trajectory

Fig. 6 illustrates good agreement between computation crack growth trajectories and those of the experiments. 2.2 Residual Life Estimation

Fig. 3. Computation crack growth trajectory

252

Here, the residual life of cracked structural element, Fig. 6. is considered, numerically and experimentally. To determine the computation of residual life for this structural element Eq. (1) is used, along the mixedmode crack growth trajectory. For that purpose, an analytic formula for equivalent stress intensity factor Keq is necessary. For determination of the analytic formula of Keq, discrete values of the stress intensity factors along the crack trajectory are used. Discrete values of SIF’s are given in Table 1.

Blažić, M. – Maksimović, S. – Petrović, Z. – Vasović, I. – Turnić, D.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 250-254

Table 1. Discrete values of SIF’s along mixed-mode crack growth trajectory a [mm] KI [daN/mm3/2] KII [daN/mm3/2] θi [o]

a0 = 3 37.6 21 45

a0 +a1 = 7 90.8 9.5 32

a0 + ... + a2 = 9.5 123.5 2.2 5.9

a0 + ... + a3 = 11.5 162.5 -1 0.7

a0 + ... + a4 = 12.9 177 2.5 3.2

a) Fig.7. Comparisons computation with experimental crack growth trajectory

3 CONCLUSIONS b) Fig. 6. Comparison computation with experimental crack growth trajectory: a) complete crack growth trajectory, b) right part trajectories

Using discrete values of SIFs from Table 1 and the relation for equivalent SIF in the next form [21]:

1/ 4

K eq =  K I4 + 8 K II4   

, (4)

we can obtain analytic formulae for the stress intensity factor along the crack growth trajectory, in accordance to Fig. 6, in the next form: 3

2

Keq = – 2E+07a + 820616a – 5217.1a + 20.311, (5) in which a is the crack length along crack trajectory. To determine the residual life of the cracked structural component, the analytic formula Eq. (5) has been used in Paris’s law, Eq. (1). Paris’s constants for considered steel (1.7225) are C = 0.00000000058, n = 2.57. Specimens are tested under cyclic load of constant amplitude in which σmax = 250 MPa and σmin = 25 MPa. The crack length versus number of loading cycles is shown in Fig. 6. The experimentally determined number of cycles before failure is Nexp = 32200 cycles, as shown in Fig. 7. Residual life estimation under mixed-mode crack growth is computed from point a0 to point a4 in accordance Fig. 6b.

During the service of various structures, including those of aircraft, crack directions are not often normal to the loading direction. In such practical cases, the direction of crack growth is not obvious. Tests to predict the fatigue crack growth trajectory for mixedmode cracks are not only costly, but they also do not explain how each structural component in a complex structure could be optimized with another’s so that the fatigue life of the overall structure can be predicted within reasonable limits for establishing the periods of inspection. In this work, a computation procedure to predict the direction and the growth rate of a mixed mode fatigue crack, using mode-I and mode-II data from finite elements, has been attempted. Computation of the crack growth trajectory is compared with experimental results. Good agreement between computation and experimental trajectories has been obtained. The residual fatigue life along mixed mode crack growth trajectory has been determined analytically and experimentally. Good agreement between residual life estimation with experiment has also been obtained. 4 ACKNOWLEDGMENTS The authors would like to thank the Ministry of Education and Science of Serbia for financial support under project number OI 174001 and TR 35045.

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5 REFERENCES [1] Journet, B., Ohrloff, N., Pavlou, D., Pantelakis, S., Scra, L., Poole, D., Smart, J. (1993). Investigation on Al-Li Alloys for Damage Tolerant Applications. B/E3250 project, task 4 (Flight Simulation), Final Report DCR/M-60365/F-93, AEROSPATIALE, Suresnes. [2] Pantelakis, S., Kermanidis, T., Pavlou, D. (1995). Fatigue crack growth retardation assessment of 2024T3 and 6061-T6 aluminum specimens. Journal of Theoretical and Applied Fracture Mechanics, vol. 22, no. 1, p. 35-42, DOI:10.1016/0167-8442(94)00046-4. [3] Pavlou, D.G. (1998). The influence of the crack tip plastic zone strain hardening on the metal high-cyclefatigue behavior. Proceedings of the 5th International Conference on Structures under shock and impact. Thessaloniki, Southampton: Computational Mechanics Publications, p. 633-647. [4] Pavlou, D.G. (2000). Prediction of fatigue crack growth under real stress histories, Engineering Structures, vol. 22, no. 12, p. 1707-1713, DOI:10.1016/S01410296(99)00069-3. [5] Paris, P.C., Erdogan, F. (1963). A critical analysis of crack propagation Laws. Transaction ASME, Journal of Basic Engineering, vol. 85, no. 4, p. 528-533, DOI:10.1115/1.3656900. [6] Erdogan, F., Sih, G.C. (1963). On the crack extension in plates under plane loading and transverse shear. Transaction ASME, Journal of Basic Engineering, vol. 85, p. 519-525, DOI:10.1115/1.3656897. [7] Gdoutos, E.E. (1984). Problems of Mixed Mode Crack Propagation. Martinus Nijhoff Publishers, Haag, p. 187-200, DOI:10.1007/978-94-009-6189-0_9. [8] Stamenkovic, D., Maksimović, K., NikolićStanojević, V., Maksimovic. S., Stupar, S., Vasović, I. (2010). Fatigue life estimation of notched structural components. Strojniški vestnik - Journal of Mechanical Engineering, vol. 56, no. 12, p. 846-852. [9] Maksimovic, S., Maksimovic, K. (2012). Improved computation method in residual life estimation of structural components, theoretical and applied mechanics. Special Issue of Theoretical and Applied Mechanics, vol. 40, (S1), p. 223-246, DOI:10.2298/ TAM12S1247M. [10] Sih, G.C., Barthelemy, B.M. (1980). Mixed-mode fatigue crack growth predictions. Engineering Fracture Mechanics, vol. 13, no. 3, p. 439-451, DOI:10.1016/0013-7944(80)90076-4. [11] Khan, S.M.A, Khraisheh, M.K. (2000). Analysis of mixed mode crack initiation angles under various loading conditions. Engineering Fracture Mechanics, vol. 67, no. 5, p. 397-419, DOI:10.1016/S00137944(00)00068-0.

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[12] Blažić, M., Maksimović, K., Assoul, Y. (2011). Determination of stress intensity factors of structural elements by surface cracks. 3th Serbian Congress Theoretical and Applied Mechanics, Vlasina Lake, p. 374-383. [13] Theoharis, P.S., Andrianopoulos, N.P. (1982). The T-criterion applied to ductile fracture. International Journal of Fracture, vol. 20, no. 4, p. R125-R130, DOI:10.1007/BF01130617. [14] Qian, J., Fatemy, A. (1996). Mixed mode fatigue crack growth: a literature survey. Engineering Fracture Mechanics, vol. 55, no. 6, p. 969-990, DOI:10.1016/ S0013-7944(96)0071-9. [15] Boljanović, S., Maksimović, S. (2011). Analysis of the crack growth propagation process under mixedmode loading. Engineering Fracture Mechanics, vol. 78, no. 8, p. 1565-1576, DOI:10.1016/ j.engfracmech.2011.02.003. [16] Shafique, Khan, S.M.A., Marwan, K.K. (2000). Analysis of mixed mode crack initiation angles under various loading conditions. Engineering Fracture Mechanics, vol. 67, no. 5, p. 397-419, DOI:10.1016/ S0013-7944(00)00068-0. [17]  Msc/NASTRAN software code- Theoretical Manuals, (1994). The MacNeal Schwendler Corporation, Los Angeles. [18] Rusinski, E., Moczko, P., Pietrusiak, D., Przybyłek, G., (2013). Experimental and numerical studies of jaw crusher supporting structure fatigue failure. Strojniški vestnik - Journal of Mechanical Engineering, vol. 59, no. 9, p. 556-563, DOI:10.5545/sv-jme.2012.940. [19] Pavlou, D.G., Labeas, G.N., Vlachakis, N.V., Pavlou, F.G. (2003). Fatigue crack propagation trajectories under mixed-mode cyclic loading. Engineering Structures, vol. 25, no. 7, p. 869-875, DOI:10.1016/ S0141-0296(03)00018-X. [20] Tanata, K. (1974). Fatigue crack propagation form a crack inclined to the cyclic tensileaxis. Engineering Fracture Mechanics, vol. 6, no. 3, p. 499-507, DOI:10.1016/0013-7944(74)90007-1. [21] Petrašinović, D., Boško, R., Petrašinovic, N. (2012). Extended finite element method (XFEM) applied to aircraft duralumin spar fatigue life estimation. Tehnički vjesnik – Technical Gazzete, vol. 19, no. 3, p. 557-562. [22] Maksimovic, S., Vasović, I., Maksimović, M., Đurić, M., (2013). Some aspects of the damage tolerance analysis of the LASTA training aircraft structures. Scientific Technical Review, vol. 63, no. 2, p. 70-74. [23] Jovicic, G., Živković, M., Maksimović, K., Đorđević, N. (2008). The crack growth analysis on a real structure using the X-FEM and EFG methods. Scientific Technical Review, vol. 58, no. 2, p. 21-25.

Blažić, M. – Maksimović, S. – Petrović, Z. – Vasović, I. – Turnić, D.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 255-264 © 2014 Journal of Mechanical Engineering. All rights reserved. DOI:10.5545/sv-jme.2013.1526

Received for review: 2013-10-29 Received revised form: 2014-01-20 Accepted for publication: 2014-02-11

Original Scientific Paper

Parametric Study of a Permanent-Magnet Stepper Motor’s Stepping Accuracy Potential Škofic, J. – Koblar, D. – Boltežar, M. Jan Škofic1 – David Koblar2 – Miha Boltežar3,* 1 Iskra-Mehanizmi,

Slovenia Slovenia 3 University of Ljubljana, Faculty of Mechanical Engineering, Slovenia 2 Domel,

This paper deals with a large-scale parametric study of geometrical influences on the stepping accuracy of a small, claw-poled, permanentmagnet stepper motor. Even though the main focus is on the claw-poles, other parameters, such as the permanent-magnet height and the stack air gap, are varied in order to obtain a complete insight into the importance of the individual details. Compactly presented results can help engineers considerably lower the times needed to find the optimum design for the developed motor. Keywords: stepper motors, permanent-magnet motors, finite-element methods, magnetostatics, torque, movement simulation Highlights • Proposed modular construction of the motor. • Large-scale parametric study. • Altered geometry PM (permanent-magnet) modeling approach. • Upgraded system of equations for simulating the rotational and axial movements of the rotor. • Representation of geometrical influences on stepping accuracy of the motor.

0 INTRODUCTION Small, claw-poled, permanent-magnet (PM), stepper motors [1] to [3] are commonly used in many home appliances, industrial applications and the automotive industry due to their positioning abilities and their relatively low cost. The performance of the motor is defined by the materials used and its geometrical details, particularly the stator poles. However, optimizing the geometry in order to achieve a higher torque can sometimes affect the stepping accuracy and the dynamics of the rotor’s movement. Optimization of the driving torque and minimization of the detent torque are common tasks in the development of prototype steppers; therefore, many published patents and investigations can be found that relate to these topics [4] to [6]. An informative investigation was published in 2006 by Liu et al. [5], where a study of different ratios of teeth bases and different teeth heights was made. Because of the constant tooth surface used in the analysis, the holding and detent peak torques presented are very informative. Another study by Liu et al. [6] was carried out to investigate the influence of the stator stack gap. The experiment showed improved stepping when the gap was increased. In order to obtain a clearer insight into the influence of multiple geometrical parameters on the driving torque, detent torque and stepping accuracy, an extensive study is presented in this paper, using

the full-scale 3-D FEM and a custom code for the stepping simulation. The results are organized to help engineers quickly find the most influential parameters to optimize their motors. Such information can prove itself highly useful in industry where short deadlines and high investments do not allow for a time-consuming research, simulations, prototype manufacturing and testing. 1 PROTOTYPE MOTOR To ensure that the effects of the investigated parameters are as “isolated” as possible, a new construction was employed for the stator. This modular concept also allows the simple manufacturing of parts, assembly and disassembly of the motor (Figs. 1 and 2). This is very important, because it saves valuable resources, while only four parts have to be changed in the assembly when investigating pole-shape design and stator-rotor air gap. Because only the poles, stator rings and permanent magnets are made of magnetic material, the geometry of the motor is very simple/ basic. The minimum necessary “errors” in the stator are, therefore, six symmetric circular-pattern holes for mounting/positioning and two small holes for the winding leads. Manufacturing and assembly errors [7] that always occur in real life can be easily tested by replacing the stack-separating washers or changing the position or the shape of the six positioning holes.

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

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two phases of the stator, a 0.35mm copper-alloy wire is used for a winding with 160 turns. The stator is formed from cold-rolled, low-carbon strip steel with an inner diameter of 23 mm and an outer diameter of 42 mm. The motor is designed to be driven with 0.8 A and ~12 V.

Fig. 1. Modular PM stepper motor

Fig. 3 and Table 1 present the parameters and their values that are investigated in this paper. The tooth base width (PA) is defined by an angle, whereas all the other parameters, such as tooth height (H), vertical chamfer (VC), horizontal chamfer (HC), stack separation (S) and permanent-magnet height (MH), are defined in millimeters. The analysis is designed to have one main version/prototype (V01) with PA = 15°, H = 5 mm, VC = 3 mm, HC = 1 mm, S = 0.2 mm, MH = 12 mm, from which all other versions derive. For example, version V02 is identical to V01, with the exception of the tooth-width parameter, which changes to PA = 14°. We decided to use this approach in order to emphasize the effect of the parameter when optimizing a base prototype. The parameters are kept well within the feasible manufacturing range.

Fig. 3. Investigated parameters Table 1: Investigated parameters PA [°] H [mm] VC [mm] HC [mm] S [mm] MH [mm]

/ 4.4 V05 / 0.7 V12 / /

/ 4.7 V06 2.5 V09 0.85 V13 0.1 V16 /

15 V01 5 3 1 0.2 12

14 V02 13 V03 12 V04 5.3 V07 5.6 V08 / 3.5 V10 4 V11 / 1.15 V14 1.3 V15 / 0.3 V17 0.4 V18 / 12.8 V19 13.6 V20 14.4 V21

2 THEORETICAL BACKGROUND 2.1 FEM Calculation Details

Fig. 2. Exploded view of the motor

The low-cost, bipolar, claw-poled, PM, stepper motor investigated in this work (Fig. 2) has 48 poles, corresponding to a 7.5° step angle. In each of the 256

The current in the winding generates a magnetic field around the leads/coils. The direction of this rotational field depends on the direction of the current flow. The newly created magnetic flux runs on the stator metal (with specific B-H curve) and polarizes the claw poles. Each pole now attracts a rotor segment of the opposite polarization, thereby generating useful torque. Since there is a lot of interest in design influences [8] to [10], a 3D FEM in Ansys [11] was used to calculate these torques and forces in multiple rotor positions. The torque-displacement curves contain a lot of information about the performance of the motor. The peak torque and zero-crossing locations determine the strength and stepping accuracy of the motor. Because the main interest is the stepping accuracy, the shape of the curve is also important. Calculating the torque with the FEM is a very accurate (often used for different model verifications [12]) and flexible [13] way to calculate the torque behavior. Influences such as the temperature, the teeth bending stresses [13] and the manufacturing errors can be easily simulated.

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Fig. 4. Magnetic flux density profile of a PM; a) typical segment approach, b) alternative geometrical approach, c) measured average magnetic flux density at approximately 0.15 mm away from the magnet surface

Since a low-cost motor is being analyzed here, the PM rotor creates a specific challenge. An Nd-Fe-Co-B magnetic powder was mixed with a temperaturestable epoxy resin and pressed into a mold to form an economical, effective and magnet [14]. However, at this stage the magnet is yet to be magnetized in the magnetizing machine to create the radially magnetized segments. Because there is a considerable amount of resin, air pockets and optionally fibers in the magnet, the energy product, residual induction and coercive force all change respectively to volumetric loading of the magnetic powder (specified by the manufacturer). The magnetized magnet also has a sinusoidal magnetic flux distribution along its tangential direction (Fig. 4c), which is in slight conflict with the result of the simulation if a simple radial segments of the PM ring are used (Fig. 4a). Due to these problems with the need of scaling the magnetic material properties and obtaining sinusoidal shaped magnetic density profile (Fig. 4c) a simple method for changing the geometry of the PM was developed [15]. The segment’s geometry and volume are changed to suit the volume portion of the magnetic material and also the inner and outer PM diameters. Such a model therefore keeps the original diameters, the thickness of the wall and the very important rotor-stator air gap [16]. The shape of the elements is not explicitly defined due to the variation of the volumetric loading of magnetic powder and angular width of the segment. For the first approximation the sine shape can be used. Typically less than 4 iterations are needed to obtain the satisfactory shape that corresponds to desired volume. The magnetic flux density of the elliptically

shaped segments (Fig. 4b) resembles the sinusoidal distribution more closely than the typical approach (Fig. 4a). Using the FEM we calculated five torquedisplacement curves: the detent and four driving curves. These driving torque-displacement curves are defined by the driving current energizing only one phase of the motor ([I1 = 0.8 A, I2 = 0 A] = c16, [I1 = 0 A, I2 = 0.8 A] = c32, [I1 = –0.8 A, I2 = 0 A] = c48, [I1 = 0 A, I2 = –0.8 A] = c64). The reconstruction of the other curves represented in any stepping mode can be made using Eq. (1) to (4), where c16, c32, c48 and c64 are the calculated primary torque vectors, cX is the reconstructed torque vector, n is the nth vector element, I1 is the current in the first phase and I2 is the current in the second phase. For 0 A < I1 < 0.8 A and –0.8 A < I2 < 0 A:

cX ( n ) = c64 ( n ) ⋅ I 2 ( cX ) + c16 ( n ) ⋅ I1 ( cX ) . (1) For 0.8 A > I1 > 0 A and 0 A < I2 < 0.8 A:

cX ( n ) = c16 ( n ) ⋅ I1 ( cX ) + c32 ( n ) ⋅ I 2 ( cX ) . (2) For 0 A > I1 > –0.8 A and 0.8 A > I2 > 0 A:

cX ( n ) = c32 ( n ) ⋅ I 2 ( cX ) + c 48 ( n ) ⋅ I1 ( cX ) . (3) For –0.8 A < I1 < 0 A and 0 A > I2 > –0.8 A:

cX ( n ) = c 48 ( n ) ⋅ I1 ( cX ) + c64 ( n ) ⋅ I 2 ( cX ) . (4)

The reconstruction of all the torque-displacement curves for the desired stepping mode is fast and simple, but also very informative. Any small errors in the primary torque-displacement curves can have a significant effect on others that have not been calculated with the FEM, resulting in a compromised stepping accuracy. The primary curves were computed with FEM and the use of B-H curve for the metal, so they include the detent torque and saturation effects. Because the reconstruction expressions (Eq. (1) to (4)) produce torque curves that are only a linear combination of the primary ones, we can expect some minor discrepancies compared to a FEM calculated curves. The calculated FEM curves are static (magnetostatics), therefore the removed material from the magnet (FEM model) does not present any drawbacks in terms of the change of mass and inertia. Realistic mass and inertia will be used as constants in an independent dynamics simulation.

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2.2 Stepping Simulation The stepping simulation was performed with the help of a developed system of equations [15] that uses torque-displacement curves and the forcedisplacement curves obtained, as described in section 2.1, to simulate both the rotational and axial [17] movements of the rotor. Once the FEM results for torque-displacement and force-displacement curves are obtained and stored, any movement simulation can be quickly calculated using Eq. (5) to (9):

di1  di  = u1 − R ⋅ i1 − M ⋅ 2 + ω ⋅ TC1,0 ( Θ )  / L1 , (5) dt  dt 

di2  di  = u2 − R ⋅ i2 − M ⋅ 1 + ω ⋅ TC0,1 ( Θ )  / L2 , (6) dt  dt  dω 1 = ⋅ [T ( i1 , i2 , Θ ) − Br ⋅ ω − sgn (ω ) ⋅ (Te + Ta + Tz )], (7) dt J 1 dz = ⋅ [ Fz ( i1 , i2 , Θ ) − Ba ⋅ z − dt mr

T − sgn ( z ) ⋅  g ⋅ mr ⋅ µ a + e  − csw ⋅ z ], (8) rs  

dΘ = ω , (9) dt

where i1, i2 are the currents in both phases, u1, u2 are the voltages, R is the winding resistance, L is the winding inductance, J is the rotor inertia, Br is the rotational movement damping constant, Ba is the axial movement damping constant, M is the mutual inductance of the windings, TC1,0 (θ) is the torque curve calculated with the FEM, when only one winding is fully energized (note that it must be calculated with 1 A winding current or normalized to 1 A to satisfy unit requirements), T(i1,i2,θ) is the torque defined by Eq. (1) to (4) for the current rotor position, Te is the friction torque generated by the rotor’s eccentricity, Ta is the extra friction torque added to the motor (gearbox, encoder, etc.), Tz is the friction torque caused by the rotor’s axial movement Eq. (10), z is the rotor’s axial velocity, mr is the mass of the rotor, Fz(i1,i2,θ) is the axial force for the current rotor position (defined in a similar way to the torque T(I1,I2,θ)), g is the acceleration due to gravity (in current application the motor’s axis lays in horizontal plane), μr is the coefficient of friction between the plain brass bearing and the stainless-steel shaft, csw is the spring-washer constant, and z is the axial displacement of the rotor. Spring washers, located between rotor and plain bearings (Fig. 2), help to maintain the axial position of 258

the rotor, and so contribute to the good vibro-acoustic aspects of the product. The axial movement of the rotor, governed by Eq. (8), causes the compression of either the upper or lower spring washer. Compression creates a friction torque TZ that affects the rotation. In Eq. (10) rbf represents the mean radius of the bearing flange and μa is the axial friction coefficient.

TZ = csw ⋅ z ⋅ rbf ⋅ µ a , (10)

Te = mr ⋅ re ⋅ ω 2 ⋅ µr ⋅ rs . (11)

The friction torque caused by the mass eccentricity of the rotor Te is described by Eq. (11), where re is the eccentricity and rs is the shaft radius. The mass eccentricity should not be confused with geometrical run-out of the rotor. The run-out would affect the magnetic field and change the FEM calculated torque-displacement and force-displacement curves [18], [19]. For the motors constructed without spring washers, Eq. (8) should be altered to incorporate the effects of a collision between the rotor and the bearings. 3 MODEL VALIDATION In order to validate the proposed simulation method, a simple experiment allowing the measurement of the rotational displacement of the shaft was made [15]. The shaft of the motor was coupled with a Scancon miniature encoder using 10-7 kgm2 of inertia and 0.5 Nmm of friction torque. The encoder has a resolution of 7500 lines per rotation, which is then multiplied by four in the measuring software. The data acquisition was performed with a 20 kHz sampling frequency. The motor was driven with an AMIS-30623, bipolar, stepper-motor chip based on LIN communication (automotive CAN standard sub frame). Fig. 5 shows 12 consecutive steps in ½-stepping mode [2], one each tenth of a second. Theoretically, the step angles should be 3.75° and all equal. In this case the problematic stator geometry causes unwanted anomalies in the torque curves, which affect the motor’s stepping abilities. A combination of short and long steps is a common problem when micro-stepping a claw-poled, PM, steppers. For that particular reason Fig. 5 shows ½-stepping instead of full-stepping mode. Comparison of the simulation (Fig. 5a) and measurement (Fig. 5b) of a stepping accuracy test in ½-stepping mode reveals an adequate quality of the simulation (typically manufactured motors have a tolerance of ±5% non-accumulative error). 12 consecutive steps are enough to see the motors poor

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stepping abilities. The schematics of experimental setup, additional measurements and model validation is available in previous work [15]. 50

Angle [°]

40 30 20 10 0

0

0.2

0.4

0.6 0.8 Time [s] a) Simulated 1/2 stepping

1

1.2

Fig. 6. Detent torque variation when changing the teeth width (parameter PA); V01 = 15, V02 = 14, V03 = 13, and V04 = 12°

50

Angle [°]

40 30 20 10 0

0

0.2

0.4

0.6 0.8 Time [s] b) Measured 1/2 stepping

1

1.2

Fig. 5. a) simulation and b) measurement comparison of the ½-stepping

4 SIMULATION RESULTS 4.1 FEM Calculated Torque Properties An investigation of the FEM calculated torquedisplacement curves not only lets us characterize the peak detent and driving torque, but also the zerocrossings and zero-crossing gradients. The gradient defines the motor’s ability to secure the step position. A steep curve is important in applications where a high operating friction is expected, because the positioning will be much more accurate. The gradient also affects the oscillating frequency of the rotor. The shape of the curve will affect the rotor’s high-speed rotational movement, because a high torque ripple causes the rotor to run more roughly. Figs. 6 to 11 are showing the detent torque-displacement curves for each prototype. Each individual figure represents the effect of the variation of only one parameter. Because prototype V01 is the base motor, its data is shown in all the figures. The situation is similar for Figs. 12 to 17, where the plotted driving torque curves were calculated with I1 = 0.8 A and I2 = 0 A.

Fig. 7. Detent torque variation when changing the teeth height (parameter H); V05 = 4.4 , V06 = 4.7 , V01 = , V07 = 5, and V08 = 5.6 mm

Fig. 8. Detent torque variation when changing the teeth vertical chamfer (parameter VC); V09 = 2.5, V01 = 3, V10 = 3.5, and V11 = 4 mm

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Fig. 9. Detent torque variation when changing the teeth horizontal chamfer (parameter HC); V12 = 0.7, V13 = 0.85, V01 = 1.0, V14 = 1.15, and V15 =1.3 mm

Fig. 10. Detent torque variation when changing the stator stack separation (parameter S); V16 = 0.1, V01 = 0.2, V17 = 0.3, V18 = 0.4

Fig. 11. Detent torque variation when changing the PM height (parameter MH); V01 = 12, V19 = 12.8, V20 = 13.6, and V21 = 14.4 mm

260

Fig. 12. Driving torque variation when changing the teeth width (parameter PA); V01 = 15, V02 = 14, V03 = 13, and V04 = 12°.

Fig. 13. Driving torque variation when changing the teeth height (parameter H); V05 = 4.4, V06 = 4.7, V01 = 5, V07 = 5.3, V08 = 5.6 mm

Fig. 14. Driving torque variation when changing the teeth vertical chamfer (parameter VC); V09 = 2.5, V01 = 3, V10 = 3.5, V11 = 4 mm

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According to the data in Figs. 6 to 17 the most influential parameter is the tooth base width (Fig. 6), for the detent characteristics, and the tooth height (Fig. 13), for the driving torque characteristics. Naturally, the peak driving torque has a lot to do with the change of the surface area of the teeth. Interestingly, the vertical chamfer has a major influence on the detent (Fig. 8), but much less influence on the driving torque (Fig. 14). Increasing the permanent-magnet height (MH) seems to have a positive effect on lowering the detent and increasing the peak driving torque. Table 2. Static torque properties Fig. 15. Driving torque variation when changing the teeth horizontal chamfer (parameter HC); V12 = 0.7, V13 = 0.85, V01 = 1.0, V14 = 1.15, V15 = 1.3 mm

Fig. 16. Driving torque variation when changing the stator stack separation (parameter S); V16 = 0.1, V01 = 0.2, V17 = 0.3, and V18 = 0.4 mm

Fig. 17. Driving torque variation when changing the PM height (parameter MH); V01 = 12, V19 = 12.8, V20 = 13.6, and V21 = 14.4 mm

V 01 02 03 04 05 06 07 08 09 10 11 12 13 14 15 16 17 18 19 20 21

PA [°] 15 14 13 12 15 15 15 15 15 15 15 15 15 15 15 15 15 15 15 15 15

H VC HC S MH [mm] [mm] [mm] [mm] [mm] 5 3 1 0.2 12 5 3 1 0.2 12 5 3 1 0.2 12 5 3 1 0.2 12 4.4 3 1 0.2 12 4.7 3 1 0.2 12 5.3 3 1 0.2 12 5.6 3 1 0.2 12 5 2.5 1 0.2 12 5 3.5 1 0.2 12 5 4 1 0.2 12 5 3 0.7 0.2 12 5 3 0.85 0.2 12 5 3 1.15 0.2 12 5 3 1.3 0.2 12 5 3 1 0.1 12 5 3 1 0.3 12 5 3 1 0.4 12 5 3 1 0.2 12.8 5 3 1 0.2 13.6 5 3 0 0.2 14.4

Ca1 [°] 81.23 79.42 75.35 66.68 82.68 82.18 79.53 77.01 81.71 79.79 76.36 71.13 77.33 83.35 84.53 81.34 81.14 81.02 81.71 82.057 82.17

Ca2 [°] -86.80 -86.93 -87.09 -87.23 -86.26 -86.56 -86.91 -87.07 -86.48 -86.95 -87.14 -86.88 -86.86 -86.56 -86.28 -86.74 -86.77 -86.77 -86.69 -86.67 -86.72

T D [Nmm] [Nmm] 92.22 3.11 90.89 4.80 87.90 6.62 84.17 7.81 77.41 4.03 85.82 3.57 95.44 2.95 94.46 3.03 91.72 2.73 92.33 4.45 91.94 6.34 91.86 5.33 92.96 4.59 90.40 3.05 87.71 4.31 92.46 3.26 91.94 3.00 91.59 2.87 94.61 2.64 96.26 2.55 97.10 2.61

Fig. 18. Graphical presentation of the static torque properties

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R 1/1 [°]

0.45 0.4 0.35 0.3

R 1/2 [°]

The simulation of the stepping was carried out by considering a real test setup, where the shaft of the motor is coupled with a miniature encoder. Therefore, the added inertia and friction of the encoder and the coupling errors are included in the simulation parameters. Every version of the motor was tested in five different regimes (1/1, 1/2, 1/4, 1/8, 1/16 stepping), where the virtual driver directed 18 consecutive steps every tenth of a second (allowing the rotor to settle down after each step) with V = 12 V and Irun = 0.8 A. The steps were then analyzed and presented in the form of a statistical range (SR), which is the difference between the maximum and minimum step in each stepping mode. The theoretical step size in the 1/1 stepping mode is 7.5°, in the 1/2 mode it is 3.75°, in the 1/4, 1.875°, and so on.

1.25 1 0.75 0.5 2.4 2 1.6 1.2

R 1/4 [°]

4.2 Stepping Simulation Results

most sine-like driving torque-displacement curves. Every deviation from ideal sine should be reflected in distorted reconstructed curves that define the static stepping potential. If we follow the individual parameter vertically (from 1/1 to 1/16 mode) in Fig. 19, the theory is confirmed as the prototype V15 with the smallest difference between Ca1 and Ca2 (and therefore the least-distorted torque-displacement curve) proves to be the most accurate motor in microstepping positioning.

R 1/8 [°]

The results are summarized in Table 2 and visually in Fig. 18, where Ca1 is the angle of torquedisplacement zero-crossing on the left of the peak, Ca2 is the angle of the torque-displacement zerocrossing on the right of the peak, the T peak driving torque and the D peak detent torque.

01 02 03 04 05 06 07 08 09 10 11 12 13 14 15 16 17 18 19 20 21

SR 1/1 [°] 0.306 0.292 0.254 0.266 0.495 0.385 0.373 0.351 0.403 0.278 0.359 0.271 0.274 0.448 0.317 0.319 0.331 0.318 0.288 0.309 0.305

SR 1/2 [°] 0.525 0.599 1.307 0.938 0.594 0.428 0.533 1.346 0.675 0.485 0.629 1.436 0.820 0.466 0.546 0.452 0.614 0.635 0.589 0.568 0.496

SR 1/4 [°] 1.971 2.233 2.573 2.576 1.640 1.749 2.202 2.436 1.781 2.347 2.265 2.435 2.344 1.435 1.070 1.934 2.045 2.078 1.919 1.823 1.812

SR 1/8 [°] 1.537 2.012 2.066 2.818 1.209 1.330 1.836 1.947 1.493 1.912 1.980 2.367 1.874 1.145 0.834 1.481 1.587 1.615 1.502 1.463 1.438

SR 1/16 [°] 0.966 1.191 1.513 1.752 0.676 0.803 1.180 1.238 0.850 1.177 1.453 1.421 1.294 0.610 0.377 0.912 0.963 0.997 0.954 0.924 0.882

Table 3 and Fig. 19 should, in theory, present the smallest step-size range for the prototypes with the 262

V01 V02 V03 V04 V05 V06 V01 V07 V08 V09 V01 V10 V11 V12 V13 V01 V14 V15 V16 V01 V17 V18 V01 V19 V20 V21

V

R 1/16 [°]

Table 3. Step angle statistical range

2.5 2 1.5 1 1.65 1.25 0.85 0.45

Fig. 19. Graphical presentation of the simulated stepping step-size range

Fig. 20. Simulation of the rotational displacement when stepping a V04 in the full-stepping mode.

For a better understanding of the data in Fig. 19, the simulated rotor displacement vs. time plots for the most accurate (V15) and most inaccurate (V04) are presented in Figs. 20 to 22. Fig. 20 shows a simulation

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StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)4, 255-264

of the stepping in the full-stepping mode for the V04 motor. Full stepping is normally not problematic for low-cost steppers. The accuracy is, in most cases and applications, satisfactory. The problems occur when driving the motor in micro-stepping modes. The steps can deviate considerably from the theoretical ones. The V04 motor is the most inaccurate motor in 1/16 stepping (unequal steps). The rotational displacement is shown in Fig. 21.

also shows that some compromises have to be made when designing a low-cost PM stepper, because the strongest motor might not be the most accurate in a particular operation. The strongest motor was found to be V21 with the highest permanent magnet, and the most accurate the V15 with increased vertical chamfer. Note that the parameters investigated in this work are not the only means of manipulating the detent and driving torque. Making some de-symmetrization [20] on the stator teeth or altered magnetization [21], [22] of the PM are patented and sometimes effective methods. More information about the optimization of the PM machines can be found in [2]. 6 ACKNOWLEDGMENT

Fig. 21. Simulation of rotational displacement when stepping a V04 in the 1/16 stepping mode.

The difference between the stepping accuracy of the least accurate (V04) and the most accurate (V15) can be seen by comparing Figs. 21 and 22. The steps of V15 motor are in comparison with V04 much more consistent (equal displacements), which is important for precise and repeatable positioning.

Fig. 22. Simulation of the rotational displacement when stepping a V15 in the 1/16 stepping mode

5 CONCLUSIONS In this paper we present the effects of motor geometry details on the stepping-accuracy potential for over 20 different prototypes simulated with an experimentally validated numerical model. The investigated parameters were varied around a base prototype in order to give the reader exact information about the effect of the applied change. This analysis

With special thanks to prof. Janez Diaci, University of Ljubljana, Faculty of Mechanical Engineering. Operation is partly financed by the European Union, European Social Fund. 7 REFERENCES [1] Acarnley, P. (2007). Stepping Motors: A Guide to Theory and Practice, 4th ed. IET, London, DOI:10.1049/ PBCE063E. [2] Gieras, J.F. (2010). Permanent Magnet Motor Technology: Design and Applications, 3rd ed. CRC Press, Boca Raton. [3] Bianculli, A.J. (1970). Stepper motors: Application and selection. IEEE Spectrum, vol. 7, no. 12, p. 25-29, DOI:10.1109/MSPEC.1970.5213082. [4] Jung, D. S., Lim, S.B., Kim, K.C., Ahn, J.S., Go, S.C., Son, Y.G., Lee, J. (2007). Optimization for improving static torque characteristic in permanent magnet stepping motor with claw poles. IEEE Transanctions on Magnetics, vol. 43, no. 4, p. 1577-1580, DOI:10.1109/ TMAG.2006.892102. [5] Liu, C.P., Li, Y.C., Liu, K.H., Wu, K.T., Yao, Y.D. (2006). Analysis of the performance of permanent magnetic stepping motor with trapezoid stator tooth. Journal of Applied Physics, vol. 99, no. 8, p. 08R31608R316-3, DOI:10.1063/1.2167637. [6] Liu, C.P., Jeng, G.R., Chen, W.C., Tsai, M.C., Wu, K.T., Yao, Y.D. (2007). Performance of claw-poled pmstepping motor. Journal of Magnetism and Magnetic Materials, vol. 310, no. 2, p. e910-e912, DOI:10.1016/j. jmmm.2006.10.945. [7] Coenen, I., van der Giet, M., Hameyer, K. (2012). Manufacturing tolerances: Estimation and prediction of cogging torque influenced by magnetization faults. IEEE Transactions on Magnetics, vol. 48, no. 5, p. 1932-1936, DOI:10.1109/TMAG.2011.2178252. [8] Ahn, J.H., Park, S.C., Rhyu, S.H., Jung, I.S. (2005). Claw-pole shape design of permanent magnet stepping

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Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 265-273 © 2014 Journal of Mechanical Engineering. All rights reserved. DOI:10.5545/sv-jme.2013.1515

Original Scientific Paper

Received for review: 2013-10-23 Received revised form: 2014-01-06 Accepted for publication: 2014-02-28

Experimental Chatter Characterization in Metal Band Sawing Thaler, T. – Potočnik, P. – Kopač, J. – Govekar, E. Tilen Thaler1,* – Primož Potočnik2 – Janez Kopač2 – Edvard Govekar2 2 University

1 PETRA stroji d.o.o., Slovenia of Ljubljana, Faculty of Mechanical Engineering, Slovenia

One of the most detrimental instability phenomena in metal band sawing is chatter, i.e. high amplitude vibrations of the tool and/or workpiece. In this paper, the influence of the cutting speed and of the distance between the blade supports on chatter phenomena is investigated. For this purpose a series of experiments with triangular cutting speed variation at several pre-selected distances between the blade supports was conducted on structural steel (St37, DIN 17100) workpieces. A feature for chatter detection was extracted from the power spectra of the machine vibration signal, and a set of characteristics was introduced for experimental chatter characterization. The results showed the presence of a chatter hysteresis which depended on the cutting speed. Additionally, apart from the blade support distance, the cutting speed was shown to be a strongly influencing parameter, and as such also promising for chatter control in band sawing processes. Keywords: chatter, metal band sawing, empirical characterization

0 INTRODUCTION In cutting processes, chatter can be characterized as the self-excited, high amplitude vibrations of the cutting blade and/or workpiece [1]. It is caused by the instability which occurs in a nonlinear cutting process, and can result in harmful effects on process performance from the point of view of quality, economy, and ecology [1] and [2]. The chatter phenomenon has been intensively researched in finalizing machining operations such as turning [3] and milling (removal rate maximization and chatter suppression) [4] and [5]. Grinding is also highly susceptible to the occurrence of chatter, so several researchers have contributed to the understanding of the chatter phenomenon by non-linear coarse-grained entropy rate analysis [6], non-linear modelling of the grinding process and chatter prediction [7], as well as chatter detection [8] and stability lobe prediction [9]. As opposed to these machining operations, band sawing is most often used at the beginning of the machining chain, and as such does not influence the final product properties as significantly as the finalizing processes. However, with the development of expensive and sometimes difficult-to-cut metal alloys and other materials, such as mono-crystalline silicon, minimization of waste material and of the time for final machining operations, as well as maximisation of the corresponding surface quality, have become increasingly important. In band sawing, the negative impact of chatter on the quality of products was first noticed in lumber cutting [10] and [11], and was reflected in the so-called wash-boarding [10] of the cut surface. To be able to understand, predict and avoid chatter vibrations in band sawing, several mathematical models were derived.

Their aim was to predict the natural frequencies of the moving continuum, representing the band saw blade. In order to predict the natural frequencies, which are related to the chatter phenomenon, models based on bending [10] and torsional deformations [11] were developed, as well as several models that were based on a moving plate with respect to tangential loading [12], parametric excitation [13], and non-conservative force excitation [14]. Experimental studies of the band saw blade vibrations showed that tensioning of the blade increases the fundamental torsional frequency, but does not affect the lateral bending frequency [15]. Current experimental and model-based solutions for chatter avoidance in band sawing suggest control of the cutting speed [10] and [16], avoidance of excitation at the natural frequencies of the cutting blade, and maximizing the tension of the cutting blade [15]. However, the relevance of complex, theoretical models based on band sawing process characterization is determined, to a high degree, by the approximations of the material characteristics, friction and temperature effects in the cutting zone, all of which are non-linear. For this reason in this research the experimental characterization was chosen for chatter characterization. In our previous research, several features based on spectral analysis were defined and successfully applied for chatter detection in band sawing [16]. The aim of this paper is to apply the most relevant chatter detection feature [16] for the characterization of chatter and related chatter hysteresis. Information on chatter and related chatter hysteresis characteristics is important for the implementation of chatter control and process optimization in the near-chatter regions. In the following section, the experimental setup and experiments are described. Band sawing

*Corr. Author’s Address: PETRA Stroji d.o.o., Cesta Andreja Bitenca 68, Ljubljana, Slovenia, tilen.thaler@pe-tra.com

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experiments with different distances between the blade supports Lb and varying cutting speeds vc were performed in order to collect experimental data about chatter. In the section: “Analysis and characterization of chatter”, the most informative feature for chatter detection is defined as the basis for further analysis. It is applied to the acquired experimental data, and the results are presented, indicating the strong influence of the cutting speed and the distance between the blade supports on onset of chatter and the characteristics of the observed chatter hysteresis. Finally, the results are summarized in the conclusions, where the most significant contribution of this work is presented and possible implementation issues in band sawing are discussed.

In order to measure and control the cutting speed vc, the band saw machine is equipped with a frequency inverter, a controller, and software for controlled variation of the cutting speed. A three-component Kistler 9257B dynamometer, a three-component PCB 356A16 piezo accelerometer, a Brüel & Kjaer 4190 microphone, a preamplifier 2669, and an amplifier 2691-OS1 were used to detect cutting forces, machine vibrations and sound emitted by the cutting process. The frequency responses of the dynamometer and accelerometer were, in terms of their first natural frequency and according to the manufacturers, 2 and 10 kHz respectively for the applied type of mounting. The microphone frequency response was ±2 dB in the frequency range 3.15 Hz to 20 kHz. Placement of the sensors is shown in Fig. 1. The workpiece was mounted on the threecomponent dynamometer, in order to measure the cutting force components F = (Ff, Fc, Fl). The subscripts c, f, and l denote the components in the cutting, feed, and lateral directions. In order to measure the machine vibrations a = (af,  ac,  al) the three-component accelerometer was mounted on the left cutting blade support. The blade supports are in direct contact with the cutting blade and represent one of the most dynamically exposed parts of the machine structure. The emitted sound pressure p was measured by the microphone, which was positioned 32 cm above the workpiece and directed toward the cutting zone. All the sensory data obtained during the band sawing process were acquired by a 16bit resolution A/D data acquisition system, and were transferred to a computer for off-line analysis and chatter characterisation. The sampling frequencies for the cutting force components (Ff, Fc, Fl), the machine

1 EXPERIMENTAL SETUP The cutting experiments were conducted on a double column PE-TRA Toolmaster 300DC band saw with a 300 mm maximum cutting width. The maximum cutting width is defined by the distance between the vertical cutting blade supports Lb, as shown in Fig.1 right. This distance Lb can be pre-set within the range from 230 to 420 mm. A bimetal cutting blade of length 4150 mm was tensioned by 2.0 kN. The characteristic parameters of the cutting blade are given in Table 1. Table 1. Cutting blade parameters Parameter Material Loop length Width/Thickness Teeth pitch Rake/Clearance angle

Value M42 4150 [mm] 34/1.1[mm] 2 to 3 [teeth per inch] 10/32 [°]

Sound Left blade support

Lb

Microphone Feed setting Right blade support

Machine vibration Accelerometer

blade supports

Cutting blade

Forces Vice

Workpiece Blade Dynamometer protection

Fc

Fl Ff

Dynamometer Blade support

Fig. 1. Double-column horizontal band saw experimental set-up

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vibration (af, ac, al), and the sound pressure p signals, were 20, 25.6 and 20 kHz, respectively. 1.1 Experiments The experimental set-up on the multi-sensory double column horizontal band saw provided the basis for the chatter characterization. In the band sawing experiments, rectangular solid profile workpieces with a width of 100 mm and a height of 60 mm, made of structural steel type St37 (according to DIN 17100), were used. The main control parameters of the experiments were the cutting speed vc and the distance between the blade supports Lb. In order to investigate the influence of the cutting speed vc(t) on the band sawing process dynamics and the related chatter, the cutting speed vc was varied in a triangular manner as shown in Fig. 2a. The experiments started with the cutting speed vc set to 34 m/min. During the cutting the vc was linearly increased so that in 30 seconds it reached maximal value 133 m/min. After a second the cutting speed was decreased at a constant rate so that the minimal value of 34 m/min was again reached in 30 seconds, and the cutting experiment, at the preset distance between the blade supports Lb, was stopped. To investigate the influence of the distance between the blade supports Lb the signals of the cutting forces F, accelerations a, and sound pressure p were acquired during a set of 35 experiments at seven preset distances between the blade supports Lb = [250, …, 400] mm with increments of 25 mm. At each preset distance between the blade supports, 5 experiments were performed, which resulted in the total number of 35 experiments. The number of 5 repeated experiments was chosen as a compromise between achieving sufficient statistical significance of the acquired data, and minimizing the wear of the cutting edges of the blade during the complete set of experiments. The process parameters depend on the selected workpiece material and are recommended by the cutting blade manufacturer. Based on this, the experiments were performed at feed rate of vf = 45 mm/min. At this feed rate, the feed component of the cutting force Ff was within the recommended limits, i.e. less than 70 N per cutting tooth of the band saw blade. The corresponding cutting rate was 100 mm2/s. Examples of the acquired signals of feed force Ff, the acceleration in the cutting direction ac, and the sound pressure p measured during the band sawing of the workpieces at a preset distance between the

blade supports of Lb = 350 mm, are shown in Fig. 2 b, c and d. In the presented signals the regions of low amplitude oscillations related to the regular chatterfree cutting are marked in blue, and the regions of large amplitude vibrations related to the cutting by chatter are marked in red. More in detail, during the linear increase of the cutting speed vc an abrupt increase in the amplitudes of the acquired signal ac can be observed at time t ≈ 9.7 s and a cutting speed of vc ≈ 67.1 m/min respectively. The large amplitude oscillations terminate at t ≈ 12.5 s and are re-excited at t ≈ 19.5 s and vc ≈ 97.3 m/min. The excited high amplitude oscillations remain present until t ≈ 41.3 s, when the cutting speed vc is decreased to vc ≈ 100.2 m/min. Similarly as during the increase of the cutting speed vc large-amplitude oscillations of short duration are excited again at time t ≈ 48.6 s and vc ≈ 75.1 m/min. The observed high amplitude vibrations at vc  ≈  67.1, 97.3 and 75.1 m/min are caused by the instability of the regular chatter free cutting, where an abrupt transition to chatter takes place. Additionally, a small cutting velocity hysteresis can be observed with respect to the first chatter onset at vco  ≈  67.1  m/min and the last chatter termination velocity vct ≈ 66.8 m/min. This observation, together with the observed abrupt changes in the oscillation amplitude, indicates that band sawing process instability and related chatter phenomena can be described by sub-critical Hopf bifurcation [1]. 2 ANALYSIS AND CHARACTERIZATION OF CHATTER In general it has been observed that all of the acquired signals appear to be informative with respect to chatter [16]. However, a time-dependent mean value of the feed force component can be observed due to the change in the cutting speed and the effects of the workpiece geometry [17] and [18]. The mean value of the feed force is proportional to the number of the teeth in the cut and the depth of the cut [18]. Thus, the geometry of the workpiece, as well as the cutting speed variation, causes the non-stationarity of the feed force component observed in Fig. 2. In the acceleration and sound measurements there are no evidences of the workpiece geometry influence. Accelerometers also have several on-site advantages compared to the dynamometers and microphones, so the acceleration signal in the cutting direction ac was used in the following analysis for the characterization of chatter in band sawing.

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Fig. 2. a) the cutting speed vc(t) triangular profile; examples of the corresponding acquired signals at the blade support distance of Lb = 350 mm; b) feed force Ff , c) the acceleration in the cutting direction ac and, d) the sound pressure p; low amplitude regular cutting is marked in blue and large amplitude chatter cutting is marked in red

Fig. 3. Spectrogram of the acceleration signal ac versus the cutting speed vc at three different distances between the blade supports Lb

2.1 Data Analysis Chatter in cutting is reflected in high amplitude oscillation of the feed force Ff, the acceleration in the cutting direction ac, and the sound pressure p were presented with indicated chatter regions (Figs. 2b, c and d). Since chatter is a dynamical phenomenon, apart from the observed chatter-related high amplitude vibrations, chatter can be more precisely characterized by changes and amplification of the power spectra components at the specific harmonic frequencies [16], [19] and [20] in the power spectra of the acquired signals. Fig. 3 shows examples of spectrograms of the acceleration signal in the cutting direction ac acquired during the band sawing process, with variation of the 268

cutting speed vc performed at three different distances between the blade supports, i.e. Lb = 250, 350 and 400 mm. The pronounced horizontal bands of amplified frequencies in the spectrograms were found to be characteristic for chatter cutting, whereas a relatively uniform, low power frequency pattern is characteristic for the regular chatter-free cutting. Based on the frequency pattern observed in the spectrograms we can see that chatter is the most present at the distance Lb = 350 mm what indicates a nonlinear influence of distance Lb on chatter in band sawing. Typical ac signal power spectra of the chatter and chatter free cutting regimes are shown in Fig. 4. As already evident from the spectrograms in the chatter

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Fig. 4. Characteristic power spectra and defined basic power spectra features x1…x5 of the acceleration signal ac in; a) chatter and, b) chatter-free cutting

power spectrum (Fig. 4a) pronounced peaks at certain frequencies are characteristic. For more detailed analysis and chatter characterisation a quantitative power spectra chatter feature y was defined as a linear combination of the basic normalized power spectra features y [16]: y = x1 + x2 + x3 – x4 – x5 .

(1)

The basic power spectra features x1, …, x5 denoted in Fig. 4 and are defined as: x1: amplitude of the 1st maximal peak in frequency range [0, 2] kHz, x2: amplitude of the 2nd harmonic peak, x3: amplitude of the 3rd harmonic peak, x4: amplitude in the middle between the 1st and 2nd harmonic peak, x5: amplitude in the middle between the 2nd and 3rd harmonic peak.

are the

the the

The defined feature y characterizes the properties of the power spectra as they are related to chatter and chatter-free cutting. The feature’s value is high for spectra with a high peak-to-valley difference, which is generally the case in the chatter regime with more deterministic signals. On the other hand the value of the feature y is low for regular cutting regimes, whose signals resemble a stochastic process. 2.2 Results Fig. 5 shows the values of the feature y of the acceleration signal ac in dependence of the cutting speed vc during the increasing of the cutting speed (blue curves), and during the decreasing of the cutting speed (red curves), for five repetitions of the cutting experiment at a distance between the blade supports of Lb = 350 mm.

Fig. 5. Feature y plotted against the increasing (blue) and decreasing (red) cutting speed vc , and corresponding chatter threshold yt = 1.47

The feature y maintains relatively low values in chatter free cutting regime until the cutting speed reaches a value of approximately vc  ≈  61  m/min. At this breakpoint cutting speed vc, y starts to increase rapidly. This rapid increase in y is caused by the transition from chatter free to chatter cutting. The threshold value yt of the feature y which indicates the onset of chatter cutting, was defined as an average value of the feature values y at the breakpoint cutting speed vc In Fig. 5 the calculated chatter threshold value yt  = 1.47 is denoted by a dashed horizontal line. Thus values of the feature y which lie below and above the threshold line yt indicate chatter-free and chatter cutting respectively. The high values of y in the interval around vc ≈ 70 m/min and vc  ≈ 104 m/min are indicators of strong chatter. A closer inspection of the feature y in the region around the breakpoint is provided in Fig. 6a, where the values of the feature y during increasing (solid) and decreasing (dashed) cutting speed vc can be seen. It is

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Fig. 6. a) Detailed view of the feature y vs. increasing (solid) and decreasing (dashed) cutting speed vc and corresponding threshold value yt. b) Mean values of the feature y and defined chatter hysteresis characteristics with respect to the threshold value yt

evident from the Fig. 5 that in the case of increasing cutting speed, the feature y on average crosses the threshold value yt at higher cutting speeds vc, than in the case of decreasing cutting speed. This indicates the presence of a hysteresis in the band sawing process with respect to the cutting speed vc, which is more clearly observed in the plot of the feature y mean value during the increasing (solid) and decreasing (dashed) cutting speed vc, shown in Fig. 6b. The presence of the hysteresis further indicates that, in the case of the band sawing process, the instability, which causes the onset of chatter, is the same as that observed in turning processes, and can be attributed to sub-critical Hopf bifurcation [1]. For further analysis and characterization of the observed chatter hysteresis, several characteristics as shown in Fig. 6b were defined. The characteristic of chatter onset vco was defined by the cutting speed vc at which the feature y (solid) at increase of vc crosses the threshold value yt. Similarly the characteristic of chatter termination cutting speed vct was defined by the cutting speed vc at which the feature y (dashed) at decrease of vc crosses the threshold value yt.

which lasted 60 seconds. The proposed set of chatter characteristics is presented in Table 2. In the following the dependence of the defined characteristics versus the distances Lb of the cutting blade supports are shown by mean of the box plots. The box plots were obtained based on 5 experiments of cutting speed variation at each considered distance Lb. The central horizontal line in the box indicates the median, whereas the upper and lower box boundaries are located at the 1st and 3rd quartiles respectively, with whiskers placed at 3-times the standard deviation of the sample, and single points (+) outside the whiskers representing outliers. In Fig. 7, the box plot of chatter onset vco, (solid) and chatter termination cutting speed vct, (dashed) against the distance between the blade supports Lb are shown.

Table 2. Chatter characteristics Characteristic Chatter onset cutting speed [m/min] Chatter termination cutting speed [m/min] Chatter hysteresis width [m/min] Relative chatter duration [%]

Symbol vco vct ∆vc t

By the difference of the chatter onset vco and chatter die-out vct cutting speeds, the width of the chatter hysteresis Δvc was defined. Further, as an additional chatter characteristic, the relative chatter duration τ, was defined as the ratio between chatter duration and the duration of the entire experiment 270

Fig. 7. Box plot of the chatter onset vco (solid) and chatter termination vct cutting speeds (dashed) vs. the distance between the blade supports Lb

From the box plot (Fig. 7) it can be seen that both the chatter onset and chatter termination median cutting speeds are around 100 m/min for the first two shorter distances Lb. At a distance of Lb  = 300 mm,

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Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 265-273

the median chatter termination cutting speed drops to vct  = 62.4 m/min, whereas the median chatter onset cutting speed remains at vco = 100 m/min. With a further increase in the distance to Lb = 325 mm, the median cutting speed of chatter onset drops to approximately vco = 68.1 m/min, and further decreases to a value of vco  = 53.4 m/min at a distance of Lb = 400 mm. The median chatter onset cutting speeds vco are always somewhat higher than the corresponding chatter termination cutting speeds vct, represented by the solid and dashed lines respectively. The connecting lines of the median chatter onset and termination speeds reveal the chatter hysteresis at all distances between the blade supports where Lb > 275 mm. The influence of the distance between the blade supports Lb on the width of the chatter hysteresis ∆vc is shown in Fig. 8.

is not affected by the distance Lb. In case of distances Lb  >  275 mm, the relative duration of the chatter τ increases as the blade support distance Lb increases.

Fig. 9. Relative duration of chatter τ vs. the blade support distance Lb

The longest relative duration of the chatter τ  =  85% took place in the case of the largest blade support distance Lb = 400 mm. 3 CONCLUSIONS

Fig. 8. Width of the cutting speed hysteresis ∆vc vs. the blade support distance Lb

In the case of distances Lb ≤ 275 mm, the median values of ∆vc are close to zero and hysteresis cannot be observed. The largest median hysteresis width ∆vc = 34.8 m/min occurs at a blade support distance of Lb = 300 mm, denoting the largest bi-stable region of the band sawing process. With further increases in the blade support distance Lb, the length of the cutting speed hysteresis ∆vc decreases, and scatters around ∆vc = 8 to 10 m/min. The observed cutting speed differences ∆vc between the chatter onset vco and chatter termination vct cutting speeds, as they depend on the blade support distance Lb, additionally confirm the presence of the hysteresis, which is characteristic for the onset of a non-linear chatter phenomenon in cutting [1] and [21]. The box plot of the relative chatter duration τ against the distance between the blade supports Lb is presented in Fig. 9. Considering the median values of τ, it can be seen from Fig. 9 that the relative chatter duration τ is short in the case of short distances Lb, and

In the paper the characterization of chatter phenomena in metal band sawing process is considered. This characterization is based on an analysis of the acquired acceleration signals in the cutting direction ac during the band sawing process. In particular, the influence of the cutting speed vc and of the distance between the blade supports Lb on chatter occurrence were investigated. For chatter detection a feature is extracted from the power spectra of the machine vibration signal, and a set of characteristics is introduced for chatter characterization. The set of characteristics beside the relative chatter duration t includes: chatter onset cutting speed vco, chatter termination cutting speed vct, and the width of the cutting speed hysteresis ∆vc. Based on the analysis of the influence of the cutting speed vc, the following conclusion can be drawn: • In general, chatter onset takes place at higher cutting speeds vco, and terminates at lower cutting speeds vct. The observed cutting speed differences ∆vc between the chatter onset and chatter termination cutting speeds indicate the presence of hysteresis, which is characteristic for subcritical Hopf bifurcation based instability of the cutting process and the related onset of the chatter phenomenon in cutting.

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The results of the analysis of the influence of the distance between the cutting blade supports Lb show that: • With an increase in the distance between the cutting blade supports Lb there is a decrease in both the chatter onset cutting speed vco and in the chatter termination cutting speed vct. • The relative chatter duration τ increases with the increase of the distance between the cutting blade supports Lb. The observed influence of the distance between the blade supports on the chatter onset speed vco and the chatter termination speed vct as well as on the relative chatter duration τ, suggest that the blade supports should be as close to the workpiece as possible. Furthermore, apart from the blade support distance Lb, which is defined by the geometry of the workpiece, the cutting speed vc is shown to be a promising parameter for chatter control in the band sawing processes, with the potential for optimal cutting in near-chatter regions. Although not considered in the paper, it can be expected that the type of material under test, as well as the tool material and wear, will have a significant effect on the chatter onset and hysteresis characteristics. 4 ACKNOWLEDGEMENT The operation was partly financed by the European Union and European Social Fund. Operation, which was implemented within the framework of the Operational Program for Human Resources Development for the period 2007 to 2013, Priority axis 1: Promoting entrepreneurship and adaptability, Main type of activity 1.1.: Experts and researchers for competitive enterprises. 5 REFERENCES [1] Gradišek, J., Govekar, E., Grabec, I. (2001). Chatter onset in non-regenerative cutting: A numerical study. Journal of Sound and Vibration, vol. 242, no. 5, p. 829838, DOI:10.1006/jsvi.2000.3388. [2] Quintana, G., Ciurana, J. (2011). Chatter in machining processes: A review. International Journal of Machine Tools and Manufacture, vol. 51, no. 5, p. 363-376, DOI:10.1016/j.ijmachtools.2011.01.001. [3] Pušavec, F., Govekar, E., Kopač, J., Jawahir, I.S. (2011). The influence of cryogenic cooling on process stability in turning operations. CIRP Annals - Manufacturing Technology, vol. 60, no. 1, p. 101-104, DOI:10.1016/ j.cirp.2011.03.096.

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Original Scientific Paper

Received for review: 2013-05-10 Received revised form: 2014-01-15 Accepted for publication: 2014-01-29

Improved Integration of Renewable Energy Sources with the Participation of Active Customers Corn, M. – Černe, G. – Papič, I. – Atanasijević-Kunc, M. Marko Corn1,2* – Gregor Černe1 – Igor Papič2 – Maja Atanasijević-Kunc2 1 INEA, Slovenia 2 University of Ljubljana, Faculty of Electrical Engineering, Slovenia

European countries are promoting the use of renewable energy sources in order to reduce their dependence on imported energy. Renewable sources of energy are a promising solution, but there are downsides too. The integration of renewable energy sources causes problems with the electric power systems, as their energy production is meteorologically dependent. In this paper a system is introduced that represents a solution to this problem, with the activation of users of the distribution part of the electricity network. The system connects the users of the distribution network that are willing to participate in the system environment based on market rules. The participation allows users to offer an adaptation of their consumption of electric energy or production in return for financial incentives. The system accepts or rejects the user’s energy adaptation offers on the basis of market principles that are driven by the optimization method (genetic algorithms in the presented case). The system optimizes the total cost amount that is reflected in the integration of a larger portion of the renewable energy sources into the electric energy system. Testing the results of the system operation demonstrates that the developed solution has improved the integration of renewable energy sources and has enabled users to profit from their adaptation to energy consumption or production. Keywords: renewable energy sources, smart grid, active customers, control, scheduling, optimization, genetic algorithms

0 INTRODUCTION The European Union (EU) is encouraging the usage of renewable energy sources (RESss) (especially the usage of sun and wind energy) for two main reasons: a reduction in carbon-emission pollution and a reduction of the import dependence of the EU on foreign energy sources. However, the large proportion of RESs causes problems in electric power systems due to the stochastic nature of their primary power sources (e.g., sun and wind). In this way the power production of RES is dependent on the current meteorological situation, which causes problems in ensuring the energy balance between the production and consumption for electric power systems [1]. The integration of a larger proportion of RES into electric power systems therefore demands additional power plants to be built, just to ensure the energy balance in cases of sudden changes in the meteorological conditions. Taking into account these RES integration problems, their usage is not so efficient and offers an opportunity for the development of new solutions. One technology that is promising more efficient RES integration is called “smart grid” and represents an upgrade of the current electric power systems. A smart grid is an electricity network that can intelligently integrate the actions of all the connected users in order to efficiently deliver a sustainable, economic and secure electricity supply [2]. Smart grids promise a more efficient RES integration with the participation of willing users that can control 274

their electric energy consumption or production. It is to be expected that, in this way, it would be easier to maintain an energy balance for electric power systems. The candidates for the corresponding users are mainly located in the distribution part of the electric network. They are the household consumers, industrial sector consumers, commercial buildings, distributed energy sources (small hydro and biomass power plants, etc.) and this also includes the development of new technologies, like energy-storage devices (electric cars and hydrogen fuel cells, etc.). This participation of users involves three major changes to the network: the monitoring of all the users, the adaptation of the distribution part of electric network and the enhancement in the stability control of the electric power system. Firstly, the monitoring of all the users, especially the RES electric energy production, enables the better stability control of the electric power system as it ensures the calculation of RES energy-production predictions [3]. Secondly, the production units, installed in the distribution part of the network, changes the network’s energy flows and, as a consequence, an adaptation of the transformers, conductors and other elements has to be made in order to ensure that the newly installed capacities can deliver their energy to the network. And thirdly, the stability control with a voltage regulation and an energy balance has to be enhanced. Voltage control enables the appropriate voltage levels for all the users to ensure the adequate operation of their appliances [4] and [5].

*Corr. Author’s Address: University of Ljubljana, Faculty of electrical engineering, Tržaška cesta 25, Ljubljana, Slovenia, marko.corn@inea.si


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The energy balance between energy production and energy consumption ensures the stable and secure operation of the electric power system. Currently, the energy balance of the electric power system is only provided by a controlled energy production. With the introduction of smart grids, providing the energy balance may be expanded to all the users that are willing to participate in the balance process. Users that are willing to participate can help to ensure the energy balance, either by the control of their electric energy consumption, by the control of their electric energy production or by a corresponding combination of both. Firstly, the control of electric energy consumption means the control of the loads. One type of load is the household appliance, like washing machines, dish washers, etc. that can postpone their energy consumption. The second type are those that can act as a storage facility for storing the electricity in thermal or any other form, like electric heaters, cooling devices, etc. [6]. These technologies are known as demand-response systems and can shift in time their energy consumption to help achieve the energy balance [7]. Secondly, control of the energy production refers to small power-plant units, like small hydro power plants, biomass power plants and others that can control their energy production. Systems that merge several distributed energy producers into a larger one are also called virtual power plants [8]. Thirdly, energy-storage devices like electric cars, hydrogen fuel cells [9] and others can be included as users, whose consumption and/or production of electric energy can be controlled. They are often called producers/consumers or simply â&#x20AC;&#x153;pro-sumersâ&#x20AC;?. This paper focuses on solving the energy balance problem that emerges with the installation of RESs into electric power systems. The proposed solution describes all the important elements. In Section 1 the system that connects all the users participating in an energy-balance process is introduced. In Section 2 the developed simulation environment is presented, comprising a group of consumers, RES producers and users that participate in the energybalance process. This environment was designed for testing and optimization purposes. In Section 3 the simulation results are presented, proving that the presented system successfully reduces the effects of RES integration and ensures the energy balance of the simulated group. Concluding remarks and plans for future investigations are explained in Section 4. The paper ends with a section that summarizes the used nomenclature.

1 SYSTEM FOR ENERGY BALANCING In the electric energy market consumers are divided into balance groups that are connected through the power network with the producers of electrical energy. Balance groups are a collection of metering points (representing consumers and producers) used to calculate the balance between the consumption and production of electrical energy. With a larger amount of RES integrated into the balance group, achieving an energy balance is more difficult and more expensive. The presented system for energy balancing (SEB) fuses demand-response systems, virtual powerplant systems and the new pro-sumers [10] in order to minimize the cost of energy used to achieve the energy balance of a balance group. Fusion is realized with the implementation of an internal energy market that consists of market participants, market products and a market control algorithm. 1.1 Participants in the Internal Energy Market The participants in the internal energy market represent all the users of the electric power system that have the capability to control their electric energy production or consumption and are also willing to participate. They all compete with each other to provide the energy that ensures the energy balance of the balance group. There are two types of participants: active customers and external participants. Active customers are the users of the balance group that are willing to participate in the energybalance process. The users of the electric power system that were used in demand-response systems, virtual power-plant systems and pro-sumers have become active customers in the internal energy market. The external participants represent electric power producers that are not a part of the balance group but can provide the energy to achieve the energy balance. The first type of external participant is an organized electric market that is a central place where the supply and demand for electricity is faced. The second type of external participants consists of different electric energy producers that can provide energy to ensure the energy balance based on individual contracts. Both represent different options for ensuring the energy balance of the balance group with the main difference in the energy prices. In order to participate on the internal energy market all the participants must trade their energy in the form of energy products.

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1.2 Products of the Internal Energy Market The basic trading products in the internal energy market are offers of electric energy. There are two types of offers: flex offers and external offers. Flex offers are generated by the active customers and external offers are provided by the external participants. 1.2.1 Active Customers’ Offers The SEB connects all the active customers into an internal energy market using flex offers. A flex offer is an energy offer generated by the active customer that sends a message to the SEB with the information – how much, when, and for which price – about how the active customer is willing to consume or produce its electric energy [11]. The flex offer (Oflex) is defined by eight parameters, as indicated in Eq. (1). The parameters of the flex offer (they are illustrated in Fig. 1) describe the availability window, the duration window, the power window and the prices. The availability time window is defined by the interval (tmin a, tmax a), the duration interval (tmin d, tmax d) and the power window (Pmax, Pmin). O flex = {tmin a , tmax a , tmin d , tmax d ,

Pmin , Pmax , p pro , pcon }. (1)

They represent the flex offer’s flexibility as they can be shifted in terms of availability time (indicated by index a), time of duration (indicated by index d) or electric power. The price parameters specify the financial incentive for the active customer’s service.

same time and demands a different cost refund for the offered services. If an active customer service is needed, the SEB assigns a flex offer, which means that the flex offer is set to an exact start time (tstart), end time (tend) and power value (P) with the respect to the flex-offer limitations. An assigned flex offer can be seen in Fig. 1 as a grey square. The active customer will start to execute its service only if its assigned flex-offer start time is equal to the current time. With the start of the execution of the assigned flex offer the active customer gets its financial incentive. 1.2.2 External Participants Offers An external offer (Oext) represents the energy that can be purchased from the external participants and is defined with three parameters: start time (text start), end time (text end), and the price for the unit of electric energy (pext), as is indicated in Eq. (2):

Oext = {text start , text end , pext }. (2)

Active customers compete with each other and with the external participants to provide the energy for achieving the energy balance of the balance group. This process takes place on the internal energy market that is driven by the internal market control algorithm. 1.3 Internal Market Control Algorithm The market control algorithm, illustrated in Fig. 2, controls the active customers by optimizing the costs of the electric energy used to achieve the energy balance of a balance group, which is reflected in a more efficient RES integration and so enables the installation of a higher portion of RESs. In the first step the market control algorithm has to calculate the predicted energy imbalances (Eprd imb) of balance group: E prd imb (l ) = E prd pro (l ) + E prd con (l ) +

Fig. 1. Example of a flex offer and assigned flex offer

There are two types of prices in the flex offer, one for production (ppro) and the other for consumption (pcon), because an active customer can offer both energy production and energy consumption at the 276

+ E prd ext (l ) + E prd AC (l ),

(3)

where l = 1, ..., m and m is prediction time. The predicted energy imbalances are imbalances between the predicted energy production (Eprd pro), the predicted energy consumption (Eprd con), the energy purchased from external participants (Eprd ext) and the energy of the active customers (Eprd AC). An example of the predicted energy imbalances can be seen in Fig. 3. The energy from the external participants represents the energy that was purchased in previous iterations

Corn, M. – Černe, G. – Papič, I. – Atanasijević-Kunc, M.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 274-282

of the control algorithm’s execution. The energy from the active customers represents the energy of the assigned flex offer that started to execute. In the next step the market control algorithm collects all the flex and external offers and it arranges them into a pool of offers.

also very successful in the cases of complex problems with a large number of parameters to be optimized. For solving the presented problem a geneticalgorithm-optimization method has been chosen, with its implementation in the Matlab Global optimization toolbox [13]. The optimization fitness function (fitsch) represents the SEB costs of balancing the energy, as indicated in Eq. (4):

fitsch =

h

∑p

i AC

i =1

Fig. 2. Steps of the internal market control algorithm

+

m

∑p

i ext (i ). (4)

i =1

The fitness function is a sum of the energy costs of active customers and the energy costs of the external offers. The last step is the purchasing of the scheduled external offer energies that the scheduler has calculated. The algorithm progresses with time by repeating the steps of prediction, scheduling and purchasing the energies. The internal market control algorithm uses the internal energy market to face all the offers of energy and searches the cheapest way to minimize the predicted imbalances. 2 SIMULATION ENVIRONMENT

Fig. 3. Example of predicted energy imbalances

The process of choosing the optimal set of offers for energy balancing is carried out by the program part called the scheduler. This scheduler calculates which offers from the pool are the most appropriate for reducing the predicted energy imbalances. The scheduler assigns or rejects the flex offers and calculates the amount of energy that has to be purchased from external offers in such a way that the predicted imbalances are eliminated and the costs are minimized. The accuracy of the scheduler determines the cost reduction of the balance group and the benefits of the active customers. Optimization search space is discrete and its complexity rises with the amount of flex offers. For this type of problem, evolutionary-based algorithms have proven to be the right tool [12]. Their main advantage is that they are

A simulation environment was developed for testing all the elements and activities that are important from the point of view of the balance group. It was implemented in Matlab [14], as this program offers efficient algorithm development and is supported by a range of toolboxes for different aspects of modeling and optimization. The balance group model consists of four elements: passive producers, passive consumers, external participants and active customers. They are illustrated in Fig. 4.

Fig. 4. Balance group model

The passive producer’s element represents the electric energy sources that cannot adapt their energy production. The passive consumer’s element represents all the consumers that do not adapt their

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energy consumption (households, industry, etc.). The external participant’s element represents the external participants of the internal energy market and the active customer’s element is the energy production or consumption derived from the active customers’ services. All four elements are connected with two discrete equations that describe the electric energy flow (Eimb) and the financial flow (CSEB): Eimb (k ) = E pr (k ) + Eco (k ) + Eext (k ) + E AC (k ), (5)

CSEB (k ) = C pr (k ) + Cco (k ) + Cext (k ) + (6)

The energies of one interval (k) from all four elements are summarized and the difference between the production and consumption is equal to the imbalanced energy (Eimb). The earnings of the SEB (CSEB) are equal to the SEB costs, which have to be paid to passive producers (Cpr), external participants (Cext) and active customers (CAC), with added SEB incomes charged to the passive consumers (Cco). The energy and the costs of the imbalances (Cimb) are always eliminated by the market control algorithm. The presented model does not include a model of the electric power network, which can be carried out under the assumption that the elements included in the balance group are relatively close to each other and the transport losses of the energy can be omitted from the description. Each element in the balance group model generates the energy and financial flow that reflects its main operation. 2.1 Passive Producers’ Model The passive producers’ model consists of the solar and wind power plants model. The combined energy production (Epr) of the passive producer model is the sum of all the energy from the solar power plants (Esolar) and the wind power plants (Ewind): E pr (k ) =

n

Esolar i (k ) +

i =1

m

∑E

wind i ( k ). (7)

i =1

The cost for the produced energy (Cpr) is the product of the produced energy and the corresponding price (psolar for solar and pwind for wind energy): C pr (k ) =

n

∑ i =1

278

Esolar i (k ) ⋅ psolar +

m

∑E i =1

2.2 Passive Consumers’ Model The passive consumers’ model consists of several consumer types, like residential, industrial and business buildings, resulting in a total passive consumer’s energy consumption (Eco):

+C AC (k ) + Cimb (k ).

The energy production of the solar and wind power plants is obtained from measurements of the corresponding power plants, which are then scaled down to meet the required energy for the simulated balance group [15].

wind i ( k ) ⋅

pwind . (8)

Eco (k ) =

n

∑E

consumer type ( k ). (9)

i =1

The SEB’s income for the consumed energy (Cco) is charged regarding the contracted price of the energy (pcontract) between the SEB and the consumer:

Cco (k ) =

n

∑E

consumer i ( k ) ⋅

pcontract i . (10)

i =1

For each of these consumers the energyconsumption profile was obtained from the measurements [15]. 2.3 External Participants’ Model The external participants’ model represents the electrical energy that is purchased from, or is sold to, the external participants. The SEB purchases electric energy from external participants if the predicted imbalances show a deficit in its energy balance and sell its energy otherwise. An example of the purchase and selling energy prices’ time profile for the external participants is presented in Fig. 5. The dynamic of the prices has its background in the energy production from different types of power plants. The most expensive energy is produced by gas power plants, as they have the shortest response time. The consequence is that they are listed closer to the current time and have the highest prices. By moving away into the future the response times of other types of power plants are getting longer and their costs are falling until they reach the prices of the cheapest power plants. The representatives of such power plants are nuclear, coal and other power plants with longer response times. They operate at nearly constant operating powers, where they have the highest efficiency and therefore the cheapest energy [16].

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distortion limits rise and at the maximum prediction time (m) the distortion (emax) can be achieved.

Fig. 5. Energy prices from external participants Fig. 6. Example of providing energy predictions from measurements

2.4 Active Customers’ Model The active customers’ model incorporates all the active customers of the balance group and describes their energy production or consumption (EAC) and the corresponding costs for the SEB (CCA):

E AC (k ) =

n

∑E

AC i ( k ), (11)

i =1

C AC (k ) =

n

∑( p

proi

or pcon i )(k ). (12)

i =1

Each active customer behaves as a separate object that messages its flex offers to the SEB and executes the results of the scheduler. After the execution of its service the active customer sends a new flex offer. 3 SIMULATION RESULTS All the parameters that define the simulation environment are presented in Table 1. The simulation of the balance group was realized by using a group of 1000 household consumers with a total energy consumption of 25 MWh per day, which generates around 1950 € per day of income with the contracted price for consumer energy at 78 €/MWh [16]. The error in the predictions that are used to calculate the predicted energy imbalance (Eq. (3)) is dependent on the applied prediction method. As a calculation of the predictions is not within the scope of this paper, another approach was used. Because the simulation is based on the production of electric energy and consumption measurements, the prediction can be calculated as data distortion in a way that the desired error of the prediction is achieved. Fig. 6 shows an example of energy predictions. The calculation of predictions (data distortions) from measurements is made by randomly alternating measurements in such a way that at the present time the measurements are not distorted and with an increasing prediction time the

This means that the measurement data was randomly altered up to 10% of the original values. The value of the maximum prediction error was chosen on the basis of the error predictions of other methods that range from 5 to 20% accuracy [16]. Table 1. Simulation parameters Prop. n m ts emax ppr max ppr min pco max pco min pres

Value 24 h 12 h 15 min 10% 150 €/MWh 50 €/MWh 100 €/MWh -10 €/MWh 41 €/MWh

Description Simulation time Prediction time Sampling time Maximum prediction error Maximum price to purchase energy Minimum price to purchase energy Maximum price to sell energy Minimum price to sell energy Price of energy from RES

The prices of the external energies reflect the prices of the energy produced by different types of power plants, from 150 €/MWh (ppr max) for most expensive gas-powered plants to 50 €/MWh (ppr min) for the cheapest nuclear-power plants [17]. The testing was divided into two parts. The first test shows the effects of RES integration on the SEB earnings, and the second test demonstrates how the usage of active customers reduces the effects of RES integration. The used parameters of the simulation are presented in Table 1. 3.1 Effects of RES Integration The results of this test are illustrated in Fig. 7. They show how the RES integration affects the earnings of the SEB without the usage of active customers. The results also indicate a 1.7% increase of the costs for the SEB with every 10% of new RES installations, which corresponds to a 7% reduction of the SEB earnings for every 10% of the new RES installations.

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The costs of the SEB presented in Fig. 7 consist of the cost of the purchased energy on the internal energy market and the cost of the RES energy.

Table 2. Parameters of AC energy offers Parameter [tmin a, tmax a] [tmin d, tmax d] [Pmin, Pmax] ppro, pcon

Value [0, 5] h [0, 2] h [-10, 10] kW 0 €, 0 €

Fig. 7. Effects of RES integration

With the higher ratio of the RES the costs for the SEB are rising despite the fact that the RES energy costs less than the cheapest energy from the external participants. This is a consequence of the higher amount of RES energy production that influences the accuracy of the predicted imbalances. The reduction of the accuracy leads to a larger amount of energy purchased from the external participants. The results are used as a reference to compare them with the SEB capabilities of the RES integration. 3.2 RES Integration with Active Customers The presented results reveal the the capabilities of SEB when it comes to reducing the effects of RES integration. It is divided into two tests. The results of the first test show the potential rise of SEB earnings without incentives to the active customers. The results of the second test show how the different prices of the active customer’s services influence the SEB earnings. 3.2.1 SEB with Free Active Customers Services The purpose of this test is to show how much the SEB can earn without incentives to the active customers. The maximum rise of the SEB’s earnings demands the minimum price for the active customer’s services. All the active customers have the same parameters for generating flex offers and are listed in Table 2. An active customer flex offer consists of 20 kWh of energy, which in 24 hours of simulation time is offered 12 times. These settings bring a maximum of 240 kWh of flexible energy per active customer, which represents approximately 0.5% of the total energy consumption for one simulation run. The results of this test are presented in Fig. 8. 280

Fig. 8. Dependence of the SEB costs from the RES ratio and active customer energy share

The results show that the usage of active customers successfully reduced the costs for the SEB, which is reflected in higher earnings for the SEB. With the use of five active customers representing 2.5% of all the consumed energy it is possible to integrate 10% more RES, while the costs or the SEB’s earnings stay the same. With an increase of the active customers’ capacity, the SEB earnings per active customer are reduced because they are approaching the limit when the active customers are covering all the energy that has to be purchased to cover the predicted energy imbalances. The SEB’s earnings created by using the active customer’s services represent a profit that is shared with the active customers. The amount of the SEB’s earnings is driven by the competition of the active customers’ prices, assuming that the external participants do not change their prices. 3.2.2 SEB with Active Customers that Charge for Their Services The presented results demonstrate how the different prices of the active customers’ services influence the amount of SEB earnings. A test was performed at 20% of the installed RES (as this is also the goal of the EU to be fulfilled by 2020 [18]) and with 10 active customers, representing 5% of the flexible energy. The results are presented in Fig. 9 for the producer type of active customers and in Fig. 10 for the consumer type of active customers.

Corn, M. – Černe, G. – Papič, I. – Atanasijević-Kunc, M.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, 274-282

Fig. 9. SEB earnings and active customer earnings for the producer type of active customers

Fig. 10. SEB earnings and active customer earnings for the consumer type of active customers

The results in Fig. 9 shows that the initial earnings of the SEB group are dropping with an increase of the active customer’s prices. The earnings run dry when the active customer’s prices are higher than the external energy prices. The results in Fig. 10 show the earnings of the production active customers are rising with the rising prices and reach a peak at 28 €/MWh. Then they start to fall because the offered prices are getting closer to the external offer’s costs. The consumer types of active customers offer their energy consumption and are willing to pay for it as long as the difference between the contracted power price (78 €/MWh) and the offered price suits their economic eligibility. 4 CONCLUSION This paper presents an effective solution for solving the energy-balance problem that emerges with the installation of renewable energy sources in electric power systems. The results revealed two ways that the proposed SEB can increase its profit. The first possibility is to use the active customers to balance the predicted

imbalances due to the changing predictions. In this case the active customers represent the stored energy that is always available to the balance group. This energy costs nothing as it waits to be used in contrast to the balance group with no active customers, where in such situations it is necessary to purchase energy from external participants for the compensation of the error in the predictions. The second way to increase the profit is to purchase cheaper energy for the active customer’s services that is available on the internal energy market for more distant times in the future. The SEB (besides the improvement in the RES integration) also opens a whole new market for active customer’s services that can profit from it. The earnings of the active customers depend on the market competition and the results show that it is possible for the active customer to sell energy production for 60 €/ MWh and purchase energy for 10 €/MWh. The active customer’s prices are always limited at the top end by the short-term energy prices from external participants and limited at the bottom end by the economic eligibility of its service. From the perspective of energy management in the electric power system the SEB can offer a solution for more efficient RES integration of the balance group. It is particularly suitable for small communities or for helping to reduce the costs of infrastructure investments for the electric power system’s operators by distributing the energy peaks of a part of an electric network. The presented solution of integration does not explicitly take into account the energy losses because of the energy transportation and the problems regarding the stability control that can arise from the changes in the energy flows for the distribution network. Our future investigations will also be directed to studying the possibilities to include a more advanced power system model that will include the grid constraints like limited capacities and transport losses, which would bring the SEB to more widespread use. 5 NOMENCLATURE Oflex tmin a, tmax a tmin d, tmax d Pmin, Pmax ppro, pcon tstart, tend, P Oext

Flex offer Availability window of the flex offer Duration window of the flex offer Power window of the flex offer Production and consumption prices of the flex offer Parameters of the assigned flex offer External offer

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text start, text end pext Eprd imb Eprd pro Eprd con Eprd ext Eprd AC fitsch Epr Eco Eext EAC Eimb CSEB Cpr Cco Cext CAC Cimb Esolar, Ewind psolar, pwind pcontract

Time window of external offer Price of the external offer Predicted energy imbalance Predicted energy production Predicted energy consumption Energy purchased from external participants Energy purchased from AC Fitness function of scheduling algorithm Energy from producers Energy to consumers Energy from external participants Energy from AC Energy imbalance Earnings of system for energy balancing Costs of energy for producers Income for energy for consumers Cost of energy for external participants Cost of energy for AC Cost of energy imbalances Energy from RES Price of RES Contracted price of energy for consumers 6 REFERENCES 

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Corn, M. – Černe, G. – Papič, I. – Atanasijević-Kunc, M.


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4 Vsebina

Vsebina Strojniški vestnik - Journal of Mechanical Engineering letnik 60, (2014), številka 4 Ljubljana, aptil 2014 ISSN 0039-2480 Izhaja mesečno

Razširjeni povzetki člankov Sergey N. Grigoriev, Victor K. Starkov, Nikolay A. Gorin, Peter Krajnik, Janez Kopac: Globoko brušenje: pregled kinematike, parametrov in vplivov na procesno učinkovitost Tine Seljak, Matjaž Kunaver, Tomaž Katrašnik: Emisijsko vrednotenje različnih vrst utekočinjenega lesa Tatiana Minav, Henri Hänninen, Antti Sinkkonen, Lasse Laurila, Juha Pyrhönen: Primerjava električnih in hidravličnih sistemov za shranjevanje energije pri regalnih viličarjih Tomaž Bešter, Matija Fajdiga, Marko Nagode: Uporaba dinamičnih testov s konstantno amplitudo za napovedovanje dobe trajanja pri različnih kontrolnih parametrih Marija Blažić, Stevan Maksimović, Zlatko Petrović, Ivana Vasović, Dragana Turnić: Določitev trajektorije rasti utrujenostne razpoke in preostale dobe uporabnosti v mešanih načinih Jan Škofic, David Koblar, Miha Boltežar: Parametrična analiza natančnosti korakanja koračnega motorja s krempljastimi poli in trajnim magnetom Tilen Thaler, Primož Potočnik, Janez Kopač, Edvard Govekar: Eksperimentalna karakterizacija drdranja pri tračnem žaganju kovin Marko Corn, Gregor Černe, Igor Papič, Maja Atanasijević-Kunc: Izboljšana integracija obnovljivih virov energije z vključitvijo aktivnih ponudnikov Osebne vesti Doktorski disertaciji, diplomske naloge

SI 43 SI 44 SI 45 SI 46 SI 47 SI 48 SI 49 SI 50

SI 51


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 43 © 2014 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2013-11-14 Prejeto popravljeno: 2014-02-04 Odobreno za objavo: 2014-02-11

Globoko brušenje: pregled kinematike, parametrov in vplivov na procesno učinkovitost

Grigoriev, S.N. – Starkov, V.K. – Gorin, N.A. – Krajnik, P. – Kopac, J. Sergey N. Grigoriev1 – Victor K. Starkov1* – Nikolay A. Gorin1 – Peter Krajnik2 – Janez Kopac2 1 Moskovska državna tehniška univerza “Stankin”, Ruska federacija 2 Univerza v Ljubljani, Fakulteta za strojništvo, Slovenija Globoko brušenje predstavlja postopek obdelave, ki omogoča velike specifične odvzeme materiala (produktivnost) zaradi velike globine rezanja. Uporablja se v raznovrstnih industrijah za npr. obdelavo lopatic plinskih turbin, obdelavo zobnikov. Za ta postopek se običajno uporabljajo superabrazivna orodja iz kubičnega borovega nitrida, vendar so v članku prikazane uspešne uporabe postopka s konvencionalnimi abrazivi iz aluminijevega oksida. Globoko brušenje se od klasičnih postopkov brušenja z majhnimi globinami rezanja razlikuje predvsem v kinematiki procesa. Slednja je od zgodnjih temeljnih raziskav, katere je opravil prof. Peklenik tekom svojega doktorskega študija, doživela veliko interpretacij iz katerih izhajajo splošno znani parametri. Kakorkoli, na tem področju je bilo v zadnjih 30 letih veliko raziskav narejenih tudi v Rusiji o čemer pa je malo znanega. Predstavitev teh raziskav je tudi eden izmed glavnih razlogov za objavo članka v angleškem jeziku v mednarodni reviji kot je Strojniški vestnik. Interpretacija učinkov procesa je odvisna od uporabljenih parametrov za njihovo analizo. Tradicionalno ti parametri vključujejo debelino odrezka, topografijo brusa in podobno. Za razliko od osnovnih parametrov so v članku na osnovi osnovnega modela kinematike procesa vpeljane alternativni parametri za analizo: velikost navideznega območja odvzema materiala, kot vprijema rezalne sile, razmerje med normalno in tangencialno silo, ter razmerje med globino rezanja in premerom brusa. Metodologija temelji na vpeljanih parametrih, ki izhajajo iz analitičnih modelov kinematike brušenja. Poudarek ni na modelih samih, zato niso podane podrobnosti eksperimentalnega dela za verifikacijo modelov. Za slednje so navedene ustrezne reference. Vključena je le interpretacija uporabljenih parametrov z vidika praktične uporabe. Iz istega razloga so podane trije primeri študije, ki ilustrirajo uporabo globokega brušenja v treh različnih industrijskih aplikacijah. Prikazani so doseženi razponi parametrov in smernice za njihovo izbiro. Smernice ki zagotavljajo da proces obratuje v področju globokega brušenja, kjer je manjša verjetnost toplotnih poškodb obdelovanca ter kjer proces dosega visoko produktivnost. Eden izmed ugotovitev je tudi ta da je smotrna uporaba brusov manjših premerov, ker so v tem primeru koristni učinki procesa doseženi pri manjših globinah rezanja. Objavljeni so tudi podatki o specifičnih stopnjah odvzema materiala, ki predstavljajo ključno merilo produktivnosti. V članku so predstavljeni do zdaj neznani parametri za analizo kinematike brušenja. Ti parametri so se vrsto desetletij uporabljali v raziskavah globokega brušenja v Rusiji o čemer je malo znanega mednarodni znanstveni in strokovni javnosti. Doprinos članka je predvsem v možnosti praktične uporabe procesa z brusi izdelanimi iz konvencionalnih abrazivov. Ti so opredeljeni z veliko poroznostjo (odprto strukturo), ki zagotavlja ustrezno hlajenje procesa in izogib toplotnim poškodbam ki vedno omejujejo zgornjo mejo produktivnosti procesa. Key words: brušenje, globoko brušenje, konvencionalno, kinematika, parametri

*Naslov avtorja za dopisobanje: Moskovska državna tehniška univerza “Stankin”, Vadkovskiy pereulok, d.1, GSP-4, Moskva, 127055, Ruska federacija, v.starkov@stankin.ru

SI 43


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 44 © 2014 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2013-05-26 Prejeto popravljeno: 2013-11-30 Odobreno za objavo: 2013-12-20

Emisijsko vrednotenje različnih vrst utekočinjenega lesa Seljak, T. – Kunaver, M. – Katrašnik, T. Tine Seljak1,2,* – Matjaž Kunaver1 – Tomaž Katrašnik2 1 Center odličnosti PoliMat, Slovenija 2 Univerza v Ljubljani, Fakulteta za strojništvo, Slovenija

V želji po nadomeščanju fosilnih goriv je v zadnjem času vedno več pozornosti usmerjene v biogoriva in goriva, proizvedena iz različnih vrst odpadkov. Zaradi svojih fizikalnih in kemičnih lastnosti sta na tem področju izpostavljena predvsem biodizel in etanol, ki ju je mogoče brez obsežnih adaptacij uporabiti v motorjih z notranjim zgorevanjem. Kot tipična predstavnika biogoriv prve generacije ju pesti predvsem njuna konkurenčnost s surovinami, namenjenimi pridelavi hrane. Ena izmed možnosti za izogibanje prehrambnim surovinam je izkoriščanje odpadne lignocelulozne biomase kot surovine za goriva motorjev z notranjim zgorevanjem. Za potrebe njene predelave so uporabni predvsem termokemični procesi za konverzijo v plinasto in tekoče stanje. Novost na tem področju je proces, ki izkorišča utekočinjanje lignocelulozne biomase v večfunkcionalnih alkoholih in zahteva bistveno nižje energijske vložke kot obstoječi procesi, kakor tudi tehnično nezahtevno procesno opremo za konverzijo biomase. V predstavljeni študiji je uporabljen prav utekočinjen les, ki nastane kot produkt tega procesa.Članek se osredotoča na vrednotenje izpustov onesnažil ob uporabi različnih tipov utekočinjenega lesa med zgorevanjem v laboratorijski plinski turbini, v intervalu temperatur pred turbino med 750 °C in 850 °C. Tipi utekočinjenega lesa obsegajo osnovno formulacijo (pH 2,5 in 25% vsebnost biomase), formulacijo z zvišano vrednostjo pH (pH 5,5 ter 25% vsebnost biomase) ter formulacijo s povečano vsebnostjo lignocelulozne biomase (pH 2,5 ter 33% vsebnost biomase). Odmiki od osnovne formulacije tako sledijo privlačnejši tehnoekonomski vrednosti samega goriva. Za izvedbo študije je bila uporabljena laboratorijska plinska turbina, opremljena s tlačnimi zaznavali, termočleni tipa K, laminarnim merilnikom masnega toka zraka ter Coriolisovim merilnikom masnega toka goriva. Sestava produktov zgorevanja je bila določena z analizatorjem izpušnih plinov z ločenimi merilnimi celicami za posamezne komponente onesnažil (CO, NOx in THC). Za uspešno razprševanje in s tem zagotovitev ustreznih koncentracijskih polj v primarnem delu zgorevalne komore so bila vsa tri preizkušana goriva predgreta na temperaturo 80 °C. Dizelsko gorivo s temperaturo 20 °C, ki je zadoščalo standardu EN 590:2009, je bilo uporabljeno za meritve referenčnih vrednosti izpustov onesnažil. Rezultati so razkrili močno odvisnost izpustov CO in THC od vsebnosti lignocelulozne biomase v posameznih tipih utekočinjenega lesa. Večja vsebnost lignocelulozne biomase vodi v povišane izpuste CO in THC v celotnem območju delovanja laboratorijske turbine. Razlog za to izhaja iz specifične molekularne strukture razpadnih produktov lignina in celuloze ter iz višje povprečne molekulske mase utekočinjenega lesa z višjo vsebnostjo biomase. Glavni vpliv lahko pripišemo visoki viskoznosti goriva, ki onemogoča uspešno formiranje curka in s tem povečuje interval nastanka gorljive zmesi, kar vodi v povečane izpuste CO in THC. Nasprotno vpliv delne nevtralizacije na izpuste CO in THC ni izrazit in potrjuje hipotezo, da koncentracija CO in THC v izpušnih plinih ni pogojena z 0,71 % dodatkom 25 % amonijevega hidroksida. Vpliv delne nevtralizacije utekočinjenega lesa je razviden le v izpustih NOx. Pri delno nevtraliziranem tipu utekočinjenega lesa so izpusti NOx povečani za približno 30 ppm po celotnem delovnem območju eksperimentalne turbine. Razlog za to je 0,71% dodatek amonijevega hidroksida, s katerim v delno nevtraliziranem tipu utekočinjenega lesa povečamo vsebnost dušika v gorivu za 0,1 mol/L. Povečani izpusti NOx so tako v veliki meri posledica nastanka NOx po t.i. FBN-mehanizmu. Nasprotno se izpusti NOx ob uporabi utekočinjenega lesa s povečano vsebnostjo biomase nekoliko znižajo, kar je mogoče pripisati nižjim temperaturam v primarni coni zgorevalne komore. V študiji je bilo prvič doseženo uspešno in stabilno zgorevanje utekočinjenega lesa s povečano vsebnostjo lignocelulozne biomase, ki je razkrilo znatno povečane izpuste CO in THC v primeru 33 % vsebnosti lignocelulozne biomase v utekočinjenem lesu. Raziskani so bili vplivi spremenjene sestave na izpuste onesnažil ter predlagani krovni mehanizmi nastanka posameznih komponent izpustov. Raziskan je bil tudi vpliv delne nevtralizacije utekočinjenega lesa na izpuste onesnažil, kjer je bila potrjena hipoteza o nezaznavnem vplivu dodatka dušika na izkoristek zgorevanja utekočinjenega lesa v smislu izpustov CO in THC. Obe ugotovitvi imata neposredno uporabno vrednost pri nadaljnjem razvoju goriva in njegovi uporabi v komercialnih mikroturbinah s podobno zasnovo. Ključne besede: utekočinjen les, gorivo, plinska turbina, emisije, odpadki, solvoliza SI 44

*Naslov avtorja za dopisovanje: Center odličnosti PoliMat, Tehnološki park 24, SI-1000 Ljubljana, Slovenija, tine.seljak@polimat.si


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 45 © 2014 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2013-11-28 Prejeto popravljeno: 2014-01-09 Odobreno za objavo: 2014-01-17

Primerjava električnih in hidravličnih sistemov za shranjevanje energije pri regalnih viličarjih Minav, T. – Hänninen, H. – Sinkkonen, A. – Laurila, L. – Pyrhönen, J. Tatiana Minav1,* – Henri Hänninen1 – Antti Sinkkonen1 – Lasse Laurila2 – Juha Pyrhönen2 1 Aaltova univerza, Tehnična šola, Oddelek za konstrukcijo in proizvodnjo, Finska 2 Tehniška univerza v Lappeenranti, LUT Energy, Finska

Cilj predstavljenega dela je analiza možnosti za regeneracijo energije pri elektrohidravličnih viličarjih. Namen študije je primerjava energijskega izkoristka električnih in neposrednih hidravličnih sistemov za shranjevanje energije. Članek preučuje metode električne in hidravlične regeneracije shranjene potencialne energije pri elektrohidravličnem viličarju. Podana je primerjava dveh podobnih viličarjev, opremljenih z električnim oz. neposrednim hidravličnim sistemom za shranjevanje energije. Prvi viličar ima električni servomotorni pogon za dviganje. Servomotor poganja hidravlično črpalko, ki lahko med spuščanjem deluje tudi kot hidromotor. Drugi viličar ima hidravlični sistem za shranjevanje energije s tlačnimi akumulatorji in blok digitalnih hidravličnih ventilov z natančnim krmiljenjem pretoka brez uhajanja. Z meritvami je bilo dokazano, da je možna regeneracija potencialne energije pri sistemih s hidravličnim in električnim shranjevanjem energije. Preučen je sistem za shranjevanje energije in podana je teoretična analiza izkoriščanja energije pri elektrohidravličnem viličarju. Članek opisuje pripravo eksperimenta in primerjavo rezultatov dveh sistemov z vidika energijske učinkovitosti. Izračunan je tudi delež prihranjene energije za električni in hidravlični testni sistem pri različnih hitrostih vilic in obremenitvah. Rezultati kažejo, da je največji 45-odstotni prihranek energije v območju prostega dviganja dosežen pri optimiziranih tlačnih akumulatorjih z neposrednim hidravličnim shranjevanjem energije. V praksi pa se pri neposrednem sistemu pojavlja težava zaradi potrebe po nadzorovanju tlaka predobremenitve hidravličnega akumulatorja ali zahteve po izbiri fiksne vrednosti tlaka predobremenitve, zato so realne povprečne vrednosti lahko tudi precej manjše. Največji delež prihranjene energije pri električnem viličarju je bil 36%. Prihranek razočara v primerjavi z rezultati za isti viličar v območju delovanja drugega cilindra. Avtorji so skušali zagotoviti podobne delovne pogoje za dva nekoliko različna regalna viličarja. Najenostavnejša rešitev ob upoštevanju omejitev in podobnosti obeh sistemov je bila z izvedbo preskusov v območju prostega dviga in z dodatno omejitvijo hoda na 1,6 m. Razmeroma kratek gib prinaša nekaj odstotkov prednosti pri prihranku za hidravlični sistem, saj je bila kapaciteta hidravličnega akumulatorja pri preskusu omejena na 16 litrov. Z izbiro območja prostega dviga namesto območja drugega cilindra se je zmanjšal dosegljivi prihranek energije za oba sistema. Za manjšo rabo energije stroja je treba bodisi izboljšati učinkovitost komponent ali pa z regeneracijo izkoristiti energijo, ki bi se sicer izgubila v procesu. V zadnjem primeru je pogosto možno ponovno uporabiti kinetično oz. potencialno energijo stroja ali njegovega podsistema. Odvisno od sistema in procesa lahko shranjevanje energije prinese pomembno skupno zmanjšanje rabe energije in daljši čas delovanja mobilnih strojev. V članku je zato podana primerjava neposrednega hidravličnega in električnega sistema za shranjevanje energije. Neposredni hidravlični sistem za shranjevanje energije odpravlja potrebo po pretvorbi energije iz hidravlične v električno obliko v fazi shranjevanja in obratno v fazi regeneracije. Raziskava je sicer izpostavila hidravlični sistem shranjevanja pri testnih viličarjih, kljub temu pa ima tudi električni pogonski sistem številne prednosti. Ne zahteva namreč predobremenjevanja in uravnavanja shranjevanja energije za določeno breme ali dvižno višino. Pristop z neposrednim hidravličnim shranjevanjem energije zato ni praktičen tam, kjer so bremena različnih tež in se premikajo na različne višine. V tem primeru ima prednost električni ali posredni sistem za hidravlično shranjevanje energije. Ključne besede: Električna regeneracija energije, shranjevanje energije, viličar, hidravlična regeneracija energije, hidravlika, svinčevo-kislinski akumulator, indukcijski motor (IM), hidravlični akumulator, sinhronski motor s trajnimi magneti (PMSM), regalni viličar

*Naslov avtorja za dopisovanje: Aaltova univerza, Tehnična šola, Oddelek za konstrukcijo in proizvodnjo, Espoo, Finska, Tatiana.minav@aalto.fi

SI 45


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 46 © 2014 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2013-07-31 Prejeto popravljeno: 2013-11-15 Odobreno za objavo: 2013-12-11

Uporaba dinamičnih testov s konstantno amplitudo za napovedovanje dobe trajanja pri različnih kontrolnih parametrih Bešter, T. – Fajdiga, M. – Nagode, M. Tomaž Bešter* – Matija Fajdiga – Marko Nagode

Univerza v Ljubljani, Fakulteta za strojništvo, Slovenija

Proizvajalci zračnih vzmeti uporabljajo dinamične preizkuse za določanje kakovosti zračnih vzmeti. Na teh testih morajo vzmeti prestati zahtevano število obremenitvenih ciklov s konstantno amplitudo. Preizkus se ustavi, ko je doseženo zahtevano število obremenitvenih ciklov ali ko pride do prevelike poškodbe na mehu vzmeti. Natančnejšo napoved dobe trajanja lahko dobimo s preizkušanjem vzmeti s standardnim obremenitvenim kolektivom. Standardni obremenitveni kolektivi so bili določeni na podlagi meritev sil na kolesu vozila pri dejanski rabi vozila, zato preizkusi s standardnim obremenitvenim kolektivom dajejo zelo dobre napovedi dobe trajanja zračne vzmeti. Težava takšnih preizkusov je, da trajajo zelo dolgo, saj mora tovorno vozilo v svoji dobi trajanja prestati 1.5×108 obremenitvenih ciklov. Prvi cilj raziskave, predstavljene v tem članku, je bilo poiskati možnost uporabe obstoječih dinamičnih testov s konstantno amplitudo za napoved dobe trajanja. Obremenitve standardnega obremenitvenega kolektiva imajo različne dinamične faktorje R, zato je potrebno za določitev dobe trajanja določiti vsaj dve Wöhlerjevi krivulji in Goodmanov diagram, da lahko izračunamo ekvivalentne obremenitve z dinamičnim faktorjem R, pri katerem poznamo Wöhlerjevo krivuljo. Obstoječe dinamične teste lahko uporabimo za konstruiranje Wöhlerjevih krivulj, če jih ne prekinemo, ko vzmet prestane zahtevano število obremenitvenih ciklov. V tem primeru tak preizkus predstavlja eno točko na Wöhlerjevi krivulji. Za določitev Wöhlerjeve krivulje je potrebno določiti vsaj še eno točko. To točko lahko določimo z dodatnim dinamičnim preizkusom; ker pa se Wöhlerjeve krivulje sekajo v točki natezne trdnosti, je dopustno to točko uporabiti za določitev Wöhlerjeve krivulje. Ker je pri zračnih vzmeteh maksimalen pomik omejen, ni mogoče vzmeti obremeniti tako, da bi povzročili pok meha. Natezne trdnosti vzmeti zato ni mogoče direktno določiti, jo je pa mogoče oceniti. Če tlak v vzmeti povečujemo do poka vzmeti, lahko silo pri tem tlaku vzamemo za natezno trdnost vzmeti. Na podlagi tako določene natezne trdnosti in rezultatov dveh dinamičnih preizkusov s konstantno amplitudo sta bili določeni dve Wöhlerjevi krivulji in Goodmanov diagram, na podlagi katerega je bil narejen izračun ekvivalentnih obremenitev. Za te obremenitve so bile izračunane dobe trajanja na podlagi elementarne in osnovne Palmgren-minerjeve hipoteze, ter Haibachove hipoteze. Drugi cilj je raziskava možnosti uporabe eksperimentalnih rezultatov, dobljenih pri enem kontrolnem parametru, n.pr. sili, za napoved dobe trajanja z drugim parametrom, n.pr. napetostjo. Če kontrolna parametra nista linearno odvisna, potem ima par dinamičnih preizkusov z enakim dinamičnim faktorjem R pri enem kontrolnem parametru različne dinamične faktorje R pri drugem kontrolnem parametru. V tem primeru določitev Wöhlerjevih krivulj in Goodmanovega diagrama ni mogoča neposredno iz rezultatov preizkusov, vendar je v članku predstavljen sistem enačb, ki omogoča določitev Goodmanovega diagrama. Primerjava napovedi dobe trajanja z uporabo različnih kontrolnih parametrov bi bila zelo draga in časovno potratna, če bi morali za vsak kontrolni parameter eksperimentalno določiti Goodmanov diagram. Z enačbami smo dokazali, da je mogoče pretvoriti Goodmanov diagram za en kontrolni parameter, npr. silo, v Godmanov diagram za drug kontrolni parameter, npr. napetost. Transformacija Goodmanovega diagrama na poljuben kontrolni parameter bo olajšala nadaljnje raziskave o vplivu kontrolnega parametra na napoved dobe trajanja. Ključne besede: podvozje vozila, zračna vzmet, obremenitveni kolektiv, dinamični preizkusi, časovna trdnost, utrujanje

SI 46

*Naslov avtorja za dopisovanje: Univerza v Ljubljani, Fakulteta za strojništvo, Aškerčeva 6, 1000 Ljubljana, Slovenija, tomaz.bester@fs.uni-lj.si


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 47 © 2014 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2013-08-01 Prejeto popravljeno: 2013-12-06 Odobreno za objavo: 2013-12-11

Določitev trajektorije rasti utrujenostne razpoke in preostale dobe uporabnosti v mešanih načinih

Blažić, M. – Maksimović, S. – Petrović, Z. – Vasović, I. – Turnić, D. Marija Blažić1,* – Stevan Maksimović1 – Zlatko Petrović3 – Ivana Vasović2 – Dragana Turnić4 1 Vojno-tehnični

institut, Srbija Goša, Srbija 3 Univerza v Beogradu, Fakulteta za strojništvo, Srbija 4 Univerza v Nišu, Fakulteta za gradbeništvo in arhitekturo, Srbija 2 Institut

Snovanje letalskih konstrukcij po pristopu dopuščanja poškodb pogosto zahteva napovedovanje rasti utrujenostnih razpok v mešanih načinih. Pot napredovanja razpoke v komponenti je po tem pristopu bistveni del simulacije utrujenostne trajnostne dobe z metodologijo lomne mehanike. Raziskava obravnava ugotavljanje trajektorije rasti razpok in preostale dobe uporabnosti za dvodimenzionalne konstrukcijske elemente v mešanih načinih. Za preučitev rasti razpok je bil obravnavan preskušanec z dvema luknjama, med katerima je bila razpoka, oblikovana za trajektorijo rasti razpoke v mešanem načinu I/II. Za ustvarjanje trajektorije rasti začetne razpoke med dvema luknjama pri cikličnih obremenitvah je bil uporabljen servohidravlični sistem MTS. Trajektorija rasti razpoke je opredeljena z numeričnimi simulacijami po metodi končnih elementov. Ko je bilo določeno polje napetosti in deformacij okrog vrha razpoke, so bili izračunani lomni parametri za mešani način, ki napovedujejo pot napredovanja razpoke v plošči. V ta namen so bili uporabljeni lomni parametri kot sta KI, KII. Ko so določeni lomni parametri, je potreben še kriterij za napovedovanje smeri rasti razpoke v problemu mešanega načina. Do sedaj je bilo predlaganih že več takšnih kriterijev. Preskusi za napovedovanje trajektorije rasti utrujenostne razpoke v mešanem načinu niso le dragi, temveč tudi ne pojasnjujejo tega, kako bi bilo mogoče optimizirati vsako komponento kompleksne sestavljene konstrukcije za napovedovanje utrujenostne trajnostne dobe celotne konstrukcije v sprejemljivih mejah, s čimer bi omogočili predpisovanje pogostosti pregledov. V tem delu je bil uporabljen računski postopek za napovedovanje smeri in hitrosti rasti utrujenostne razpoke v mešanem načinu z uporabo podatkov načinov I in II iz analize po MKE. Izračuni trajektorije rasti razpoke so bili primerjani z rezultati eksperimentov in ugotovljeno je bilo dobro ujemanje računskih in eksperimentalnih trajektorij. Preostala utrujenostna doba po trajektoriji rasti razpoke v mešanem načinu je bila določena analitično in eksperimentalno. Ugotovljeno je bilo dobro ujemanje ocene preostale dobe uporabnosti z rezultati eksperimentov. Rezultati numerične simulacije trajektorije rasti razpoke v mešanih načinih po MKE so bili primerjani z rezultati eksperimentov. Preostala utrujenostna doba vzdolž ukrivljene trajektorije rasti razpoke v mešanem načinu je bila določena analitično in eksperimentalno. Iz rezultatov je mogoče povzeti, da se rezultati predlagane numerične simulacijske metode ugotavljanja trajektorije rasti razpoke in preostale dobe uporabnosti dobro ujemajo z rezultati eksperimentov. Ključne besede: rast utrujenostne razpoke, mešani načini, ocenjevanje preostale dobe uporabnosti, simulacija po metodi končnih elementov

*Naslov avtorja za dopisovanje: Vojno-tehnični institut, Ratka Resanovića 1, 11000 Beograd, Srbija, vti@vti.vs.rs

SI 47


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Prejeto v recenzijo: 2013-10-29 Prejeto popravljeno: 2014-01-20 Odobreno za objavo: 2014-02-11

Parametrična analiza natančnosti korakanja koračnega motorja s krempljastimi poli in trajnim magnetom Škofic, J. – Koblar, D. – Boltežar, M. Jan Škofic1 – David Koblar2 – Miha Boltežar3,* 1 Iskra-Mehanizmi,

Slovenija Slovenija 3 Univerza v Ljubljani, Fakulteta za strojništvo, Slovenija 2 Domel,

Pri koračnih motorjih s krempljastimi poli in trajnim magnetom, katerih glavna lastnost je natančno kotno pozicioniranje rotorja, v procesu razvoja hitro naletimo na težave pri zagotavljanju natančne pozicije tako pri polno-koračnem režimu, kot tudi pri mikro-koračnem režimu krmiljenja. Čeprav napaka pozicije ni akumulativna, je takšno delovanje v večini primerov nesprejemljivo in hkrati izkazuje slabšo kvaliteto izdelka. V tem prispevku je obravnavan pristop modeliranja gibanja rotorja ter parametrična analiza vplivnosti obsežnega števila parametrov geometrije motorja za pridobitev širokega zornega kota razumevanja obnašanja obravnavanega motorja. Modificiran pristop modeliranja gibanja rotorja, tako v rotacijski kot tudi aksialni smeri, je sestavljen iz treh temeljnih segmentov. Za izračun poteka sil in navorov na rotor motorja v odvisnosti od kota zasuka se uporabi celoten 3D MKE model motorja brez predpostavljenih simetrij. Pri tem se trajen magnet modelira s pristopom spremenjene geometrije, ki omogoča realno sinusno porazdelitev gostote magnetnega polja in direkten vnos magnetnih lastnosti uporabljenega magnetnega prahu. Preračunavanje obsežnega modela za vse kombinacije tokov v tuljavah je časovno potratno, zato metoda v drugem segmentu omogoča enostavno rekonstrukcijo željenih navornih krivulj iz osnovnih primarnih štirih krivulj. V zadnjem segmentu s pomočjo novega sistema diferencialnih enačb za popis rotacijskega in aksialnega gibanja izračunamo korakanje motorja. Simulacija gibanja poteka ločeno od MKE modela (uporablja le matriko motornih navorov in sil, pridobljenih v drugem segmentu metode) in je zato zelo hitra, kar omogoča praktično neomejene možnosti simuliranja premikov rotorja. Aksialno gibanje rotorja ima zelo majhen vpliv na rotacijsko gibanje in je zato praviloma zanemarjeno. Pomembnejšo vlogo aksialni premiki igrajo v vibro-akustičnih aspektih motorja in kasneje mehatronskega sklopa, saj lahko močno vzbudijo strukturo in posledično generirajo nesprejemljiv hrup. Validacija modela je bila izvedena z eksperimentom na prototipnem motorju z enakimi geometrijskimi in materialnimi lastnostmi kot pri modelu. Rotacijski premiki so bili merjeni z miniaturnim enkoderjem, pritrjenim na gred motorja. S tem dodana vztrajnost in trenje enkoderja sta bili upoštevani v modelu za povečanje realnosti in pridobitev ustreznega ujemanja. Validiran model je bil uporabljen za analizo vplivnosti parametrov geometrije motorja na natančnost korakanja motorja. Poleg parametrov krempljastih polov sta obravnavani še variaciji razmaka med posameznima fazama in višina trajnega magneta rotorja. Več kot 20 virtualnih prototipov je bilo nato analizirano na področju zadrževalnega in aktivnega navora ter natančnosti korakanja v 1, 1/2, 1/4, 1/8, 1/16 – koračnem režimu. Rezultati maksimalnih vrednosti navorov, simetričnosti naklonov krivulj v stabilnih legah in napake dolžin korakov v posameznih režimih so za enostavnejšo berljivost predstavljene tako v tabelarični kot grafični obliki. Rezultati analize ponujajo konkretno rešitev za izboljšanje natančnosti korakanja koračnega motorja. Hkrati je podan obsežen pregled nad posameznimi vplivi na natančnost korakanja, kar omogoča učinkovitejše konstruiranje (optimiranje) tega tipa naprav in s tem zmanjšuje število potrebnih simulacij za dosego cilja. Ključne besede: koračni motor, trajni magnet, metoda končnih elementov, magnetostatika, navor, simulacija gibanja

SI 48

*Naslov avtorja za dopisovanje: Univerza v Ljubljani, Fakulteta za strojništvo, Aškerčeva 6, 1000 Ljubljana, Slovenija, miha.boltezar@fs.uni-lj.si


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 49 © 2014 Strojniški vestnik. Vse pravice pridržane.

Prejeto v recenzijo: 2013-10-23 Prejeto popravljeno: 2014-01-06 Odobreno za objavo: 2014-02-28

Eksperimentalna karakterizacija drdranja pri tračnem žaganju kovin Thaler, T. – Potočnik, P. – Kopač, J. – Govekar, E. Tilen Thaler1,* – Primož Potočnik2 – Janez Kopač2 – Edvard Govekar2 2 Univerza

1 PETRA stroji d.o.o., Slovenia v Ljubljani, Fakulteta za strojništvo, Slovenia

Pri visoko hitrostnem odrezovanju kovin s tračno žago se med procesom odrezovanja zaradi nestabilnosti procesa pogosto pojavijo visoko-amplitudna nihanja orodja in/ali obdelovanca. Nastala visoko amplitudna nihanja imenujemo drdranje. Drdranje izrazito škodljivo vpliva na kakovost površine obdelovancev, obrabo orodja in posredno na storilnost procesa. Eden glavnih ciljev razvoja modernih tračnih žag je izdelava inteligentne tračne žage, ki bi bila zmožna samodejno prilagoditi procesne parametre pri pogoju maksimalne storilnosti odrezovalnega procesa. Procesna parametra, ki najbolj vplivata na pojav drdranja, sta rezalna hitrost in razdalja med podporama orodja, zato smo z raziskovalnim delom, ki ga objavljamo v tem članku, okarakterizirali pojav drdranja v odvisnosti od omenjenih dveh parametrov. Karakterizacija drdranja v odvisnosti od parametrov procesa je ključnega pomena pri razvoju algoritma za samodejno prilagajanje procesnih parametrov pri maksimalni storilnosti procesa. Namen raziskav je bil določitev vplivnih parametrov ter karakterizacija njihovega vpliva na pojav drdranja. Procesna parametra, ki najbolj vplivata na pojav drdranja, sta rezalna hitrost in razdalja med podporama rezalnega orodja. Za karakterizacijo drdranja smo pri različnih razdaljah med podporama orodja izvedli niz rezalnih poizkusov z linearno naraščajočo in padajočo rezalno hitrostjo. Tak način spreminjanja rezalne hitrosti je omogočil analizo vpliva celotnega območja rezalne hitrosti, tako za primer naraščanja kot tudi zniževanja rezalne hitrosti v enem poskusu. Za posredni opis procesa odrezovanja smo uporabili signale pospeškov nihanj podpore rezalnega orodja, signale rezalnih sil in signale zvočnega tlaka, ki se generira med procesom odrezovanja. Pri nadaljnji analizi in karakterizaciji pojva drdranja smo kot najbolj informativne uporabili signale pospeškov nihanj podpore rezalnega orodja. Zajete signale smo analizirali z empirično metodo, ki smo jo za namen karakterizacije drdranja razvili na osnovi kratkočasovne Fourierjeve transformacije. Izkazalo se je, da se na osnovi analize izpeljane značilke različno odzivajo na dinamski pojav drdranja in normalnega odrezovanja in tako omogočajo karakterizacijo drdranja v odvisnosti od procesnih parametrov. Opredeljena je bila informativna značilka, ki omogoča zanesljivo zaznavanje nastanka in prisotnosti drdranja. Zaznana je bila tudi histereza drdranja v odvisnosti od rezalne hitrosti, ki je značilna za nestabilnost procesa odrezovanja oziroma pojav drdranja pri odrezovalnih procesih. V nadaljevanju smo histerezo drdranja okarakterizirali s širino histereze in z relativnim časom trajanja drdranja, ki predstavljata osnovo za razvoj algoritma za samodejni nadzor drdranja. Karakterizacija drdranja, je bila izvedena na preizkušancih različnih profilov iz ogljikovega konstrukcijskega jekla St37 po standardu DIN 17100. Razvita metodologija eksperimentov in analize podatkov je splošno uporabna tudi pri raziskavi pojava drdranja pri preizkušancih iz drugih kovinskih in nekovinskih materialov. Rezultati in ugotovitve so izjemnega pomena za nadaljnji razvoj inteligentnih tračnih žag s sposobnostjo samodejnega zaznavanja in preprečevanja pojava drdranja. Ključne besede: tračno žaganje, nestabilnost procesa, drdranje, histereza, empirična karakterizacija, spektralna analiza

*Naslov avtorja za dopisovanje: PETRA Stroji d.o.o., Cesta Andreja Bitenca 68, Ljubljana, Slovenija, tilen.thaler@pe-tra.com

SI 49


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Prejeto v recenzijo: 2013-05-10 Prejeto popravljeno: 2014-01-15 Odobreno za objavo: 2014-01-29

Izboljšana integracija obnovljivih virov energije z vključitvijo aktivnih ponudnikov Corn, M. – Černe, G. – Papič, I. – Atanasijević-Kunc, M. Marko Corn1,2* – Gregor Černe1 – Igor Papič2 – Maja Atanasijević-Kunc2 1 INEA, Slovenija 2 Univerza v Ljubljani, Fakultete za elektrotehniko, Slovenija

V tem članku je predstavljen sistem, ki vključuje vse tipične uporabnike elektroenergetskega sistema na osnovi tržnih pravil. Obstoječe rešitve v aktivnih distribucijskih omrežjih so virtualne elektrarne in sistemi prilagajanja odjema na zahtevo. Virtualne elektrarne predstavljajo sisteme, ki povezujejo množico malih proizvajalcev električne energije na distribucijskem delu elektroenergetskega omrežja v večji sistem, ki nato lažje nastopa na trgu z električno energijo. Podobno pa tudi sistemi prilagajanja odjema združujejo več enot, ki lahko prilagajajo svoj odjem na zahtevo v večji sistem. Predstavljeni pristop združuje sistem virtualnih elektrarn in sistem prilagajanja porabe na zahtevo v enoten sistem. V ta sistem smo nato vključili še povezavo z organiziranim trgom električne energije ter tako vzpostavili konkurenco med vsemi tremi dobavitelji energije. Osnova predstavljenega sistema je tako model bilančne skupine. Model vključuje vse tipične uporabnike bilančne skupine in je sestavljen iz štirih sklopov: pasivni proizvajalci, pasivni porabniki, zunanji ponudniki in aktivni ponudniki. Pasivni proizvajalci so vsi uporabniki, ki ne zmorejo prilagajati proizvodnjo električne energije. Pasivni porabniki električne energije predstavljajo porabnike, ki porabljajo električno energijo po svojih potrebah. Zunanji ponudniki so vsi proizvajalci električne energije, ki sodelujejo na organiziranem trgu z električno energijo. Aktivni ponudniki pa predstavljajo porabnike in manjše proizvajalce električne energije, ki lahko prilagodijo svojo porabo oz. proizvodnjo električne energije na zahtevo upravljalca bilančne skupine. Energijska ponudba obsega 7 parametrov: najzgodnejši in najpoznejši čas veljavnosti, najkrajši in najdaljši čas trajanja ponudbe, najmanjša in največja moč ponudbe ter cena za energijo, ki jo proda (če je aktivni ponudnik proizvajalec) in cena za energijo, ki je ne porabi (če je aktivni ponudnik porabnik in prestavlja svoj odjem). Proces izravnave odstopanj med porabo in proizvodnjo električne energije znotraj bilančne skupine zagotavlja energijsko bilanco in je sestavljen iz štirih korakov: izračun napovedanih odstopanj, zbiranje energijskih ponudb, izračun razporeditve energijskih ponudb in nakupa manjkajoče energije na organiziranem trgu električne energije. Testiranje je potekalo v treh fazah: testiranje vpliva integracije obnovljivih virov v bilančno skupino, testiranje vključitve aktivnih ponudnikov, ki ponujajo svoje storitve brezplačno in testiranje vključitve aktivnih ponudnikov, ki zahtevajo plačilo za svoje storitve. Rezultati prvega testiranja so pokazali, da se stroški upravljalca bilančne skupine povečujejo za približno 2% na vsakih 10% novih obnovljivih virov energije. Rezultati drugega testa kažejo, koliko lahko aktivni ponudniki znižajo stroške upravljalcu, ki so posledica povečanega deleža obnovljivih virov energije, če za svoje storitve ne zahtevajo plačila. Z večanjem števila aktivnih ponudnikov se stroški za upravljalca bilančne skupine nižajo. Tako je mogoče z aktivnimi ponudniki, ki predstavljajo 2,5 % celotne energije, integrirati do 10% novih obnovljivih virov energije in pri tem stroški upravljalca bilančne skupine ostajajo enaki. Tretji test pa prikazuje gibanje zaslužkov upravljalca bilančne skupine in zaslužkov aktivnih uporabnikov. Test je razdeljen v dva dela: uporaba aktivnih ponudnikov, ki ponujajo proizvodnjo električne energije in uporaba aktivnih ponudnikov, ki ponujajo porabo električne energije. Prvi del testa je pokazal, da pri nizki ceni večino zaslužka pobere upravljalec bilančne skupine, ko ponudniki višajo cene, se zaslužek upravljalca zmanjšuje in povečuje zaslužek aktivnih ponudnikov. Drugi del testa pa vključuje aktivne ponudnike, ki ponujajo porabo električne energije. Ko aktivni ponudniki ponudijo plačilo za porabo električne energije po ceni navadnega porabnika (80€/MWh), ima upravljalec bilančne skupine največji zaslužek, ko pa ponujajo nižje plačilo, se zaslužek upravljalca bilančne skupine niža in povečuje zaslužek aktivnega ponudnika. Rezultati testiranj kažejo, da tak sistem sicer lahko uspešno deluje, vendar so zaslužki vsaj na strani aktivnih ponudnikov, ki ponujajo proizvodnjo električne energije, prenizki, za ponudnike, ki ponujajo porabo električne energije pa je tak sistem ekonomsko bolj smiseln, saj obljublja zaslužek, ki bi jih znal stimulirati za vključitev v tak sistem. Testiranje je bilo omejeno samo na bilančno skupino, brez upoštevanja omrežja s svojimi izgubami in omejitvami pri dostavi električne energije. Ključne besede: aktivna distribucijska omrežja, obnovljivi viri energije, aktivni ponudniki, razporejanje, optimizacija, genetski algoritmi 50

*Naslov avtorja za dopisovanje: Univerza v Ljubljani, Fakultete za elektrotehniko, Tržaška cesta 25, Ljubljana, Slovenija, marko.corn@inea.si


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 51-53 Osebne objave

Doktorski disertaciji, diplomske naloge

DOKTORSKI DISERTACIJI Na Fakulteti za strojništvo Univerze v Ljubljani sta obranila svojo doktorsko disertacijo: ●    dne 28. marca 2014 BOJAN STARMAN z naslovom: »Modeliranje mehanskega odziva anizotropne pločevine pri velikih plastičnih deformacijah« (mentor: prof. dr. Boris Štok); Tekom procesa preoblikovanja je pločevina podvržena obremenitvam, pri katerih je napetostno deformacijsko stanje v pločevini velikokrat na meji porušitve. Za pravilno in verodostojno napoved mejnega stanja preoblikovalnosti je izbiran ustrezni fizikalno objektivni konstitutivni model, ki temelji na popisu plastične anizotropije, utrjevanja in fenomenov mehanike poškodbe. Model je implementiran v programsko kodo končnih elementov, za določitev parametrov modela so izvedeni mehanski preizkusi standardnih nateznih, strižnih in zarezanih preizkušancev. Na primerih numeričnih simulacij je pokazano, da ima pri napovedi končnega stanja ključno vlogo pravilen popis omenjenih fenomenov na podlagi izvedenih mehanskih preizkusov, prikazana metodologija pa rezultira v izboljšani napovedi mejnega stanja preoblikovalnosti; ●    dne 31. marca 2014 David VEGELJ z naslovom: »Adaptivno lasersko spajanje rotorskih in statorskih lamel« (mentor: prof. dr. Janez Možina); Doktorska disertacija obravnava razvoj eksperimentalnega sistema in optimizacijo postopka laserskega spajanja statorskih in rotorskih paketov. Na osnovi hipoteze, da je možno spajanje lamel optimirati z usmerjanjem laserskega snopa na stičišča med lamelami, je bil zasnovan in postavljen računalniško upravljan adaptivni laserski sistem z dvema laserskima viroma, s katerima je možno izvajati bodisi modulirano kontinuirno bodisi adaptivno bliskovno spajanje lamel. Za natančno določitev lege posameznih lamel glede na gorišče laserskega snopa je bil uporabljen laserski odbojnostni senzor. Rezultati širokega nabora eksperimentov adaptivnega laserskega spajanja dokazujejo, da lahko na ta način izdelamo mehansko ustrezne statorske in rotorske pakete z bistveno nižjo porabo laserske energije na dolžino zvara. Ker se ob tem znatno zmanjša tudi nepotrebno pretaljevanje materiala po celotni dolžini zvarov, je predvsem pri adaptivnem bliskovnem spajanju opazno tudi izboljšanje elektromagnetnih lastnosti statorskih in rotorskih paketov. Tako izdelani manjši zvari bodo pripomogli k večjemu izkoristku

elektromotorjev, katerih jedra bodo izdelana z adaptivnim laserskim bliskovnim postopkom. DIPLOMSKE NALOGE Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv univerzitetni diplomirani inženir strojništva: dne 26. marca 2014: Miha AVGUŠTIN z naslovom: »Optimizacija procesa brizganja zobate letve« (mentor: doc. dr. Davorin Kramar, somentor: prof. dr. Mirko Soković); Jan JUVAN z naslovom: »Vpliv premaza aktivnih delcev pri varjenju nerjavnega jekla« (mentor: doc. dr. Damjan Klobčar, somentor: prof. dr. Janez Tušek); Anton KORAČIN z naslovom: »Upravičenost kurilne naprave na lesno biomaso za stanovanjsko stavbo« (mentor: prof. dr. Vincenc Butala); Anže TAVČAR z naslovom: »Vgrajena energija v povprečni slovenski stavbi« (mentor: prof. dr. Vincenc Butala); Miha OGRIS z naslovom: »Testiranje sesalnikov za razvrstitev v energetski razred« (mentor: izr. prof. dr. Jože Tavčar, somentor: prof. dr. Jožef Duhovnik); Jure ŠKORJA z naslovom: »Numerična analiza zgorevanja homogene zmesi s kompresijskim vžigom« (mentor: izr. prof. dr. Tomaž Katrašnik); Matija TRAMPUŠ z naslovom: »Določitev zmogljivosti in analiza energijskih tokov prototipnega tekmovalnega električnega vozila« (mentor: izr. prof. dr. Tomaž Katrašnik); dne 27. marca 2014: Tomaž GUZEJ z naslovom: »Prenova industrijske kotlovnice in prigradnja kogeneracijskega postrojenja« (mentor: izr. prof. dr. Andrej Senegačnik); Špela KOSEC z naslovom: »Fizikalna priprava procesne vode za hladilne stolpe« (mentor: prof. dr. Iztok Golobič); Denis MUHIĆ z naslovom: »Razvoj družine kompaktnih odpraševalnih naprav« (mentor: prof. dr. Marko Nagode, somentor: prof. dr. Iztok Golobič); Tadej Stepišnik PERDIH z naslovom: »Numerična analiza tokovnih razmer v kotlu na prašno kurjavo« (mentor: izr. prof. dr. Mihael Sekavčnik); dne 28. marca 2014: Jure SMILJANIĆ z naslovom: »Numerična analiza mehanskega odziva fiziološke arterije pri vgradnji žilne opornice« (mentor: prof. dr. Boris Štok); SI 51


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 51-53

Saša ŠAVER z naslovom: »Snovanje in konstruiranje pomožnih pekarskih naprav pri peki kasetnega kruha« (mentor: doc. dr. Samo Zupan, somentor: prof. dr. Ivan Prebil); dne 31. marca 2014: Janez BOZJA z naslovom: »Vodenje proizvodnje na osnovi standardnih kazalnikov« (mentor: prof. dr. Alojz Sluga); Primož MARAŽ z naslovom: »Zasnova, analiza in optimiranje pogonske membrane kot ojačevalnika poti aktuatorja za pnevmatični preklopni ventil« (mentor: izr. prof. dr. Niko Herakovič); Marko SKRT z naslovom: »Vpliv parametrov sušenja na lastnosti silikagela« (mentor: izr. prof. dr. Roman Šturm); Tomaž STARMAN z naslovom: »Raziskava vpliva izbranih hidravličnih tesnil na trenje« (mentor: doc. dr. Franc Majdič, somentor: prof. dr. Mitjan Kalin); Črt TRČEK z naslovom: »Vpliv radialnih utorov na pretoke skozi kolobarjaste reže v hidravliki« (mentor: doc. dr. Franc Majdič); Staš KOS z naslovom: »Modeliranje shranjevanja toplote v zemeljskih hranilnikih solarnih ogrevalnih sistemov« (mentor: prof. dr. Sašo Medved); Dejan KOZJEK z naslovom: »Kavitacijska erozija vodnega kamna« (mentor: prof. dr. Branko Širok, somentor: izr. prof. dr. Marko Hočevar); Luka NAGODE z naslovom: »Vpliv energetske sanacije stavb na učinkovitost soproizvodnje električne energije in toplote« (mentor: prof. dr. Sašo Medved). * Na Fakulteti za strojništvo Univerze v Ljubljani je zagovarjal svoje diplomsko delo (UNI - Erasmus): dne 26. marca 2014: Carlos LOZANO HONTECILLAS z naslovom: »Učinkovitost numeričnih metod za modeliranje tokovnega polja v polnilnem taktu motorjev s prisilnim vžigom / Evaluation of numerical methods for modeling flow field during the intake stroke of spark-ignited engines« (mentor: izr. prof. dr. Tomaž Katrašnik, somentor: izr. prof. dr. Matevž Dular). * Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv magister inženir strojništva: dne 26. marca 2014: Marko TEMENT z naslovom: »Vpliv hlajenja in prezračevanja na kakovost notranjega okolja in SI 52

vpeljava skupnega merila« (mentor: prof. dr. Vincenc Butala, somentor: doc. dr. Matjaž Prek); dne 27. marca 2014: Ana MARUŠIČ z naslovom: »Multiparametrična enačba stanja za zmesi« (mentor: prof. dr. Iztok Golobič). dne 28. marca 2014: Jurij ŠVEGELJ z naslovom: »Razvoj sistema za samodejno približanje vzglavnika glavi potnika pri naletu vozila od zadaj« (mentor: izr. prof. dr. Jernej Klemenc); dne 31. marca 2014: Matej JARM z naslovom: »Regresijska analiza pri kontroli elektronskih korektorjev prostornin zemeljskega plina« (mentor: prof. dr. Branko Širok). * Na Fakulteti za strojništvo Univerze v Mariboru sta pridobila naziv magister inženir strojništva: dne 26. marca 2014: Leon MARKL z naslovom: »Projekt strojnih instalacij za Galerijo« (mentor: doc. dr. Matjaž Ramšak, somentor: izr. prof. dr. Jure Marn); Žan ŠINKOVEC z naslovom: »Konstruiranje premnikov električnega avtomobila« (mentor: prof. dr. Srečko Glodež). * Na Fakulteti za strojništvo Univerze v Mariboru sta pridobila naziv magister inženir mehatronike: dne 7. marca 2014: Aleš KAPUN z naslovom: »Virtualna proizvodna celica z robotom ACMA XR701« (mentor: izr. prof. dr. Karl Gotlih, somentorja: asist. dr. Simon Brezovnik, doc. dr. Miran Rodič); Ivan KELEMINA z naslovom: »Integrirani pretvornik za napajanje pogona in baterij električnega vozila v režimu polnjenja baterij« (mentor: izr. prof. dr. Karl Gotlih, somentorja: doc. dr. Miran Rodič, prof. dr. Miro Milanovič). * Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv diplomirani inženir strojništva: dne 6. marca 2014: Marko BRADEŠKO z naslovom: »Preizkus vzdržljivosti hidravličnih valjev« (mentor: doc. dr. Franc Majdič);


Strojniški vestnik - Journal of Mechanical Engineering 60(2014)4, SI 51-53

Matevž GERŽELJ z naslovom: »Konstruiranje heksakopterja za namen snemanja iz zraka« (mentor: izr. prof. dr. Tadej Kosel); Alen MURTIČ z naslovom: »Pretok hidravlične kapljevine skozi zaslonke in dušilke pri nizkih temperaturah« (mentor: doc. dr. Franc Majdič); Gregor VODOPIVEC z naslovom: »Orodje za injekcijsko brizganje nosilca navijalnika cevi iz termoplasta« (mentor: doc. dr. Joško Valentinčič, somentor: izr. prof. dr. Tomaž Pepelnjak). * Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv diplomirani gospodarki inženir (UN): dne 27. marca 2014: Jernej BREČKO z naslovom: »Organizacija in razvrščanje delovnih mest v proizvodno usmerjenih podjetjih« (mentor: doc. dr. Iztok Palčič, somentor: prof. dr. Vojko Potočan); Matjaž STEMELAK z naslovom: »Nadzor kakovosti procesa izdelave pokrova diferenciala« (mentor: prof. dr. Bojan Ačko, somentor: prof. dr. Duško Uršič);

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SI 53


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

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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/. Please note that file size limit at the journal’s website is 8Mb. Announcement: The authors are kindly invited to submitt the paper through our web site: http://ojs.sv-jme.eu. Please note that file size limit at the journal’s website is 8Mb. The Author is also able to accompany the paper with Supplementary Files in the form of Cover Letter, data sets, research instruments, source texts, etc. The Author is able to track the submission through the editorial process - as well as participate in the copyediting and proofreading of submissions accepted for publication - by logging in, and using the username and password provided. Please provide a cover letter stating the following information about the submitted paper: 1. Paper title, list of authors and affiliations. 2. The type of your paper: original scientific paper (1.01), review scientific paper (1.02) or short scientific paper (1.03). 3. A declaration that your paper is unpublished work, not considered elsewhere for publication. 4. State the value of the paper or its practical, theoretical and scientific implications. What is new in the paper with respect to the state-of-the-art in the published papers? 5. We kindly ask you to suggest at least two reviewers for your paper and give us their names and contact information (email). Every manuscript submitted to the SV-JME undergoes the course of the peer-review process. THE FORMAT OF THE MANUSCRIPT The manuscript should be written in the following format: - A Title, which adequately describes the content of the manuscript. - An Abstract should not exceed 250 words. The Abstract should state the principal objectives and the scope of the investigation, as well as the methodology employed. It should summarize the results and state the principal conclusions. - 6 significant key words should follow the abstract to aid indexing. - An Introduction, which should provide a review of recent literature and sufficient background information to allow the results of the article to be understood and evaluated. - A Theory or experimental methods used. - An Experimental section, which should provide details of the experimental set-up and the methods used for obtaining the results. - A Results section, which should clearly and concisely present the data using figures and tables where appropriate. - A Discussion section, which should describe the relationships and generalizations shown by the results and discuss the significance of the results making comparisons with previously published work. (It may be appropriate to combine the Results and Discussion sections into a single section to improve the clarity). - Conclusions, which should present one or more conclusions that have been drawn from the results and subsequent discussion and do not duplicate the Abstract. - References, which must be cited consecutively in the text using square brackets [1] and collected together in a reference list at the end of the manuscript. Units - standard SI symbols and abbreviations should be used. Symbols for physical quantities in the text should be written in italics (e.g. v, T, n, etc.). Symbols for units that consist of letters should be in plain text (e.g. ms-1, K, min, mm, etc.) Abbreviations should be spelt out in full on first appearance, e.g., variable time geometry (VTG). 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. Figures must be cited in a consecutive numerical order in the text and referred to in both the text and the caption as Fig. 1, Fig. 2, etc. Figures should be prepared without borders and on white grounding and should be sent separately in their original formats. Pictures may be saved in resolution good enough for printing in any common format, e.g. BMP, GIF or JPG. However, graphs and line drawings should be prepared as vector images, e.g. CDR, AI. When labeling axes, physical quantities, e.g. t, v, m, etc. should be used whenever possible to minimize the need to label the axes in two languages. Multi-curve graphs should have individual curves marked with a symbol. The meaning of the symbol should be explained in the figure caption. 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. Books: Surname 1, Initials, Surname 2, Initials (year). Title. Publisher, place of publication. [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. The proof will be sent to the author showing the final layout of the article. Proof correction must be minimal and fast. Thus it is essential that manuscripts are accurate when submitted. 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

60 (2014) 4

Strojniški vestnik Journal of Mechanical Engineering

Since 1955

Papers

213

Sergey N. Grigoriev, Victor K. Starkov, Nikolay A. Gorin, Peter Krajnik, Janez Kopac: Creep-Feed Grinding: An Overview of Kinematics, Parameters and Effects on Process Efficiency

221

Tine Seljak, Matjaž Kunaver, Tomaž Katrašnik: Emission Evaluation of Different Types of Liquefied Wood

232

Tatiana Minav, Henri Hänninen, Antti Sinkkonen, Lasse Laurila, Juha Pyrhönen: Electric or Hydraulic Energy Recovery Systems in a Reach Truck – A Comparison

241

Tomaž Bešter, Matija Fajdiga, Marko Nagode: Application of Constant Amplitude Dynamic Tests for Life Prediction of Air Springs at Various Control Parameters

250

Marija Blažić, Stevan Maksimović, Zlatko Petrović, Ivana Vasović, Dragana Turnić: Determination of Fatigue Crack Growth Trajectory and Residual Life under Mixed Modes

255

Jan Škofic, David Koblar, Miha Boltežar: Parametric Study of a Permanent-Magnet Stepper Motor’s Stepping Accuracy Potential

265

Tilen Thaler, Primož Potočnik, Janez Kopač, Edvard Govekar: Experimental Chatter Characterization in Metal Band Sawing

274

Marko Corn, Gregor Černe, Igor Papič, Maja Atanasijević-Kunc: Improved Integration of Renewable Energy Sources with the Participation of Active Customers

Journal of Mechanical Engineering - Strojniški vestnik

Contents

4 year 2014 volume 60 no.


Journal of Mechanical Engineering 2014 4