Journal of Mechanical Engineering 2014 11 by Darko Svetak

http://www.sv-jme.eu

60 (2014) 10

Since 1955

Papers

619

Uroš Stritih, Andrej Bombač: Description and Analysis of Adsorption Heat Storage Device

629

Rita Ambu, Andrea Manuello Bertetto, Costantino Falchi: Design of a Prototype System Operant in Lunar Environment

638

Xin Jin, Hua Liu, Wenbin Ju: Wind Turbine Seismic Load Analysis Based on Numerical Calculation

649

Srečko Glodež, Marko Šori, Tomaž Verlak: A Computational Model for Bending Fatigue Analyses of Sintered Gears

656

Periyakgounder Suresh, Rajamanickam Venkatesan, Tamilperruvalathan Sekar, Natarajan Elango, Varatharajan Sathiyamoorthy: Optimization of Intervening Variables in MicroEDM of SS 316L using a Genetic Algorithm and Response-Surface Methodology

665

Liu Lei, Zhang Desheng, Zhao Jiyun: Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve

675

Oğuz Çolak: Optimization of Machining Performance in High-Pressure Assisted Turning of Ti6Al4V Alloy

Journal of Mechanical Engineering - Strojniški vestnik

Contents

10 year 2014 volume 60 no.

Strojniški vestnik Journal of Mechanical Engineering

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 photos have been taken in the framework of the European project RES-e Regions (Renewable Energy Sources – Electricity) in which the Faculty of Mechanical Engineering participated. The task was a secondary school competition for the best photography on the subject of renewable energy sources in Slovenia. The co-ordinator of the project was the Upper-Austrian Energy Agency, from Linz.

Courtesy: University of Ljubljana, Faculty of Mechanical Engineering, Slovenia.

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. We would like to thank the reviewers who have taken part in the peerreview process.

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.

The journal is subsidized by Slovenian Research Agency. Strojniški vestnik - Journal of Mechanical Engineering is also available on http://www.sv-jme.eu, where you access also to papers’ supplements, such as simulations, etc.

Instructions for Authors All manuscripts must be in English. Pages should be numbered sequentially. The maximum length of contributions is 10 pages. Longer contributions will only be accepted if authors provide justification in a cover letter. Short manuscripts should be less than 4 pages. For full instructions see the Authors Guideline section on the journal’s website: http://en.sv-jme.eu/. 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.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10 Contents

Contents Strojniški vestnik - Journal of Mechanical Engineering volume 60, (2014), number 10 Ljubljana, October 2014 ISSN 0039-2480 Published monthly

Papers Uroš Stritih, Andrej Bombač: Description and Analysis of Adsorption Heat Storage Device Rita Ambu, Andrea Manuello Bertetto, Costantino Falchi: Design of a Prototype System Operant in Lunar Environment Xin Jin, Hua Liu, Wenbin Ju: Wind Turbine Seismic Load Analysis Based on Numerical Calculation Srečko Glodež, Marko Šori, Tomaž Verlak: A Computational Model for Bending Fatigue Analyses of Sintered Gears Periyakgounder Suresh, Rajamanickam Venkatesan, Tamilperruvalathan Sekar, Natarajan Elango, Varatharajan Sathiyamoorthy: Optimization of Intervening Variables in MicroEDM of SS 316L using a Genetic Algorithm and Response-Surface Methodology Liu Lei, Zhang Desheng, Zhao Jiyun: Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve Oğuz Çolak: Optimization of Machining Performance in High-Pressure Assisted Turning of Ti6Al4V Alloy

619 629 638 649 656 665 675

Received for review: 2014-03-26 Received revised form: 2014-05-26 Accepted for publication: 2014-06-24

Description and Analysis of Adsorption Heat Storage Device Stritih, U. – Bombač, A. Uroš Stritih* – Andrej Bombač

University of Ljubljana, Faculty of Mechanical Engineering, Slovenia In this study, the main goal was to develop an adsorption heat storage system for domestic heating system gained by solar collectors and to indicate a new way of maintaining the energy. Main characteristic of the storage system is to retain the energy for a longer period, as long as the adsorbent and adsorbat are separated. This research analysis the influence of several parameters on the adsorption heat storage system such as: the quantity of the stored heat, the in- and out-flowing water temperatures, the water mass flow and the saturation of the adsorbent. Adsorbent HX-13 sodium aluminosilicate Na2O Al2O3 * 2SiO2 in the form of granule was used. Regarding the results we can be conclude that the low adsorption heat storage is low hydrophilic characteristic of the used adsorbent. Keywords: heat storage device, thermo-chemical heat storage, adsorbat, adsorbents, experiment

0 INTRODUCTION Increased use of energy, diminishing supplies of fossil fuels and concern for a better environment tends to exploit energy from renewable energy sources, among others e.g., gas accumulator in water hydraulic systems [1]. As a result, energy storage devices were developed. They store energy from natural sources (solar [2], geothermal) as well as industrial waste heat [3] that makes heat storage especially useful in industry, where lots of heating or chilling is needed [4]. Increasing emphasis is also given to heat storage systems for domestic heating and cooling [5] to [7]. Sorption technologies gained a lot of interests for solar heat storage of solar energy in recent years due to their high energy densities and long-term preservation ability for thermal energy [8]. Sorption heat storage device is based on the principle of thermo-chemical storage. Ongoing research and development studies show that the challenges of the technology focus on the aspects of different types of sorption materials [9] and [10], the configurations of absorption cycles and advanced adsorption reactors. Booming progress illustrates that sorption thermal storage is a realistic and sustainable option for storing solar heat energy, especially for long-term applications. In the thermo-chemical energy storage methodology, sorption material and gas are chosen as the sorption working couple for energy storage, whereby thermal energy is stored in the form of chemical bonds resulting from the sorption process between the material and the gas [11]. The current state in the field of adsorptive heat transformation (AHT) cannot be considered as wholly satisfactory, due to inappropriate thermodynamic properties of adsorbent materials [12], so studies tend to improve materials. Developed composite sorbents made from active salts and porous materials showed improved

sorption storage capacity when compared with waterbased energy storage due to its high energy density [13] to [15]. Exergy analysis performed by Abedin and Rosen [16] showed that thermo-chemical sorption energy storage system may be as efficient as and even more stable than other types of thermal energy storage system. Restuccia et al. [17] presented the experimental results of a lab-scale chilling module working with the composite sorbent SWS-1L, which is a promising alternative to the common zeolite or silica gel, for application in solid sorption and compared experimental results with a theoretical simulation. Their approach was also used to study the influence of the main operating parameters on the system performance and to make recommendations on how to improve the chillers design and process parameters. Glaznev et al. [18] investigated how a gradual tuning of pore size affects the water sorption properties; composites with smaller pores are able to generate cold rejecting the adsorption heat to hotter environment. Stronger water bounding by the salt confined to smaller pores results in the appropriate enhancement of the desorption temperature. It was shown that the small, micro porous aluminophosphates are among the most suitable materials for low temperature (solar) heat storage. In the last two decades, several studies propose different developed adsorption solar cooling systems which allow a good compromise between high reliability and good performance. From those studies, simple tubular module is considered as one of the most efficient configurations [19] and [20]. The adsorption tubular module is a tube in which an adsorber, a condenser and an evaporator are all completely housed to construct a small scale adsorption cooling unit. Heat pipes can also be used in adsorption refrigeration systems to improve the performance of

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

619

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

the systems. Thermally driven adsorption chillers and heat pumps have received increased attention in recent years. Heat pumping devices which are operating on sorption effects, current trends and forthcoming applications were studied and discussed by Ziegler [21] and [22]. Henninger et al. [23] analysed current available active sorption materials (like silica gels and zeolites), recently developed materials (aluminophosphates) and silica-aluminophosphates (SAPO) as well as novel materials (like metal organic frameworks under continuous thermal cycling in a water vapor atmosphere) in order to evaluate their suitability for the use in a periodically working heat pump based on water as working fluid. Besides solar cooling application for thermal cooling using waste heat at low temperature level (<80 °C for adsorption systems) the combination of co-generation units with sorption chiller is promising as described in [24]. While adsorption thermal storage systems are useful for long-term (seasonal) storage, systems with phase change materials (PCM’s) are used for short-term (daily) storage [25] and [26]. The aim of this work is to present sorption thermal storage tank with HX13 which can be used for long term thermal energy storage. Parts of thermal storage, experimental results of adsorption and desorption, i.e. temperatures, heat fluxes and heat are presented and explained.

desorption. This break is the time when the storage heat is not needed.

Fig. 1. Working principle of a closed-cycle desorption/adsorption heat storage

Fig. 2. p-T-x diagram of thermochemical storage cycle

1 THEORETICAL ANALYSIS Closed sorption system utilysing water used as the adsorbate [adsorptive], is shown in Fig. 1. During the process of desorption, adsorbent is connected and heated through the heat exchanger. Surface bounded adsorbate is evaporated from adsorbent and heat is transferred from the condenser. Desorption mode stops when the adsorbent is dried out (depends on the input heat amount in the adsorber), i.e. adsorbate is condensed in the condenser, or when the adsorbent and adsorbate are separated. During the process of adsorption, adsorption heat has to be discharged from the adsorber and evaporation heat supplied into the evaporator. If this is not possible, sorption process reaches thermodynamic equilibrium and the flow of water vapour is stopped. When the adsorbent is saturated with water vapour, the process ends and the tank is discharged. It needs to be recharged again (desorption). Circular process of adsorption storage shown in Fig. 2 is the same as a circular process of a heat pump, except here is a ‘time break’ during the process of adsorption and 620

The adsorbed mass of the fluid (an adsorbate) varies between a minimum (line C-D) and the maximum value (line A-B). Adsorption takes place (between points D-A) at the evaporation pressure pe and desorption (between points B-C) at a pressure of condensation pc [5]. This four-step cycle consists of: 1. Heat storage in the tank begins with isosteric heating of the adsorbent (humidity is constant) through the heat exchanger. The valve that separates the tank from the evaporator/condenser is closed. The temperature and pressure in the adsorber are increasing along the line between points A and B as long as the temperature Tdes1 is reached. At this temperature, pressure is equal to the pressure of condensation of the adsorbed adsorbate (point B) [25]. 2. At point B the valve between the condenser and the adsorber is opened. Adsorbent is heated (the line between the points B-C) until the adsorbate is separated and the adsorbent is dry. Maximum available temperature Tdes2 is reached. Meanwhile condenser is cooled to maintain the pressure

Stritih, U. – Bombač, A.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

of condensation pc. Heat generated by the condensation Qc is entrained into the environment as a waste heat. This influence on the temperature of condensation Tc, which must be kept as low as possible in order to make optimal desorption. At maximum attained temperature Tdes2 the level of saturation of the adsorbent with adsorbate is the lowest (xmin), so cycle is stopped with the closure of the valve between condenser and adsorber. When the valve is closed - and the adsorbate and adsorbent are separated - the tank is filled. Thus, the stored energy can be used without loss, even after a longer period [27]. 3. Depending on energy storage duration and ambient temperature, the temperature of the heat storage falls. Temperature drop and, consequently, the pressure drop are shown with isostera, line between points C and D. The temperature may drop to Tads1 or even lower [27]. 4. Discharging begins with the opening of the valve that separates the evaporator from the adsorber. Before that a higher pressure level in the evaporator than in the adsorber is established, because of the added heat. Heat can be provided from a variety of sources (solar, earth, etc.). Liquid is evaporated at a pressure pe and temperature Te. The valve opens; adsorbate is re-adsorbed in the adsorbent. Adsorption heat is released (line D-A). Heat generated is called useful heat and can be transferred or used in the heating system through a heat exchanger in the adsorber [22]. 2 MASS AND ENERGY BALANCE OF ADSORPTION HEAT STORAGE Schematic mass balance of sorption heat storage system is presented in Fig. 3, with adsorber container to the left and the evaporator/condenser container to the right. In the desorption mode, the mass flow of the water vapor ( m vapor ) is transferred from the adsorber to the condenser. In the adsorption mode, water vapor is transferred in the opposite direction from the evaporator to the adsorber. The condensate is evaporated and bounded with adsorber [28]. Due to the mass transfer certain properties in the control volume (dashed line) change as follows: • the mass of water in the condenser/evaporator (Δm), • saturation of the adsorber (Δx). Fig. 4 schematically shows the energy balance, where some of the processes occur only during the adsorption and some of them only during desorption mode. The adsorber is heated with the heat source

(QDes). This heat is transferred through the heat exchanger to adsorbent where it is used to separate molecules of water from adsorbent and for evaporation. The water vapour entering the evaporator/condenser contains internal energy (Uv) and heat of evaporation (QE) that was needed to desorb the water from the aluminosilicate. Water vapour is condensed in the condenser. Its internal energy is discharged (i) through the heat exchanger as condensation heat (QC) which could be used to preheat the sanitary water and (ii) partially through the tank walls into the surroundings as heat loss (QLoss).

Fig. 3. Mass balance of adsorption heat storage

In adsorption mode, the evaporator is heated with the heat source (QAds). Heat input is on a lower temperature level than QDes. The heat vaporizes the condensate; chemical reaction forms molecular bonds between water vapor and adsorbent. Water vapour molecules convert its kinetic energy into thermal energy to form the adsorption heat. Most of the adsorption heat (QAds) was drained out through a heat exchanger, the rest represents heat loss into surroundings through the tank walls (QLoss).

Fig. 4. Energy balance of adsorption heat storage

The energy balance based on Fig. 4 can be written as:

QDes + QE – QAds – QC – QLoss = 0 .

Description and Analysis of Adsorption Heat Storage Device

(1) 621

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

3 PRINCIPLE OF THE OPERATION The heat produced by solar collector is stored in adsorption heat storage and later used for house heating. The system is designed so that it can be connected to a heating or a solar cycle. Sorption system is cyclical, which means that in the summer time tank is being charged and discharged in the winter. Scheme of closed system for residential house heating with an adsorption storage tank and the main parts of this system is shown in the Fig. 5. The scheme on Fig. 5 shows the solar collector, the adsorption heat reservoir, the heating system and the low temperature source of energy. Some elements operate only in the summer time regime and others in the winter. When it is sunny and warm and the level of solar radiation is highest (up to 900 W/m2), the water passing through the solar collector heats up. Solar collector used for heating buildings and sanitary hot water is medium temperature system converting solar radiation into heat. This system works in the range from 45 to 120 °C [29]. Heated water comes through the valve for winter/summer arrangement into sorption tank and is led from here to the adsorber, wherein the adsorbent is moisturized.

heating is required. Then the valve for winter/summer regime is moved to the winter circuit. Again, we have a primary and a secondary circuit. In the primary circuit water in the evaporator/condenser is evaporated with the heat from solar collector. During the cold weather, as often in winter, the collector does not get sufficient energy and additional energy source, which acts as a reserve, is needed. These can be electric heaters or heat pumps. Heated water is led into an evaporator/condenser and condensate is evaporated through the heat exchanger. Cooled water from the evaporator/condenser returns to the solar energy collector or to the additional source. In the secondary circuit, the valve between the adsorber and the evaporator/condenser is open and water vapour is bound to the adsorbent, chemical reaction occurs and begins to release heat. Thus formed adsorption heat warms up the water in heat exchanger. From here, the heated water is led to the water heating system (i.g. radiators, pipes for floor heating). Adsorption heat can be used until the adsorbent is moistened to a certain degree. When moistened, the tank is empty and we switch to the next container (depending on model). 4 EXPERIMENTAL STUDY Inspiration for the tank design came from the sketch of the tank that has been published in the report of the European project Modular high energy density sorption heat storage (MODESTORE) [5]. This is the second generation of the storage tank, which differs from the first mainly in the fact that the adsorber and the evaporator/condenser are in a single container. It is shown in Fig. 6.

Fig. 5. System concept of the sorption reservoir for solar heating

The water releases the heat through the heat exchanger and the dry adsorbent. Water returning from the heat exchanger is cooled and is led into the collector. Described circle is called the primary circuit. In the secondary circuit, due to the heating of adsorbent with hot water, adsorbate separates and is led to the evaporator /compensator where it condensates. If the released heat during condensation is at a sufficiently high temperature level it can be used for example to preheat domestic hot water. Condensate, which is collected in the evaporator/ condenser, is separated from the adsorbent with valve (depending on the design). Once the adsorbent is dried, the phase of charging the storage ends. Usually, the discharge phase begins when the temperature of the surroundings is sufficiently low and building 622

Stritih, U. – Bombač, A.

Fig. 6. Sorption thermal storage tank

StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

Adsorbent as material was meant as polydispersed system, formed of granules in diameter from 1 to 5 mm, Fig. 7. Chemical name is â&#x20AC;&#x2DC;sodium aluminosilicate adsorbentâ&#x20AC;&#x2122; with chemical formulae Na2O Al2O3 * 2SiO2. The colour of the adsorbent is a greyish white, the bulk density of 323 kg/m3 and the molecular weight of 365 kg/kmol. The heat exchanger has two vertical copper tubes (10 mm) and eight horizontal tubes (8 mm). The vertical tubes are sealed at the bottom with two copper pins. Height of the vertical pipe is 600 mm. The length of the horizontal tube is 1430 mm. The pipes are at an interval of 50 mm. For the manufacture of the basket for the adsorbent two pieces of expanded (pitting) copper plate dimensions with holes 0.5 mm was used.

(no. 1, 2, 3, 4) and connected to the data acquisition device.

a) b) Fig. 9. a) Heat exchanger with basket; and b) top view of the heat exchanger in the tank

Fig. 7. Photo of aluminosilicate Hx-13

In Fig. 8a the front view of evaporator/condenser and interior of the tank with electric heater (Fig. 8b) is shown. Further, heat exchanger with adsorbent basket (Fig. 9a) and whole for water supply (Fig. 9b) are presented in Fig. 9.

b) Fig. 8. a) Evaporator/condenser; and b) top view flange and electric heater

In our experiment the behaviour of the adsorption tank (Fig. 10), during the charging and discharging was observed. Fig. 11 presents the measuring line during the adsorption/desorption process where the inlet and outlet water temperature, the adsorbent temperature and water temperature were measured. Temperature sensors were placed on certain locations

Fig. 10. Heat storage tank

Fig. 11. Heat exchanger and the measuring instruments

Agilent 34970 (9), measured temperatures were further processed with personal computer (10). Working pressure of the adsorption was found to be between 20 and 60 mbar absolute, which means that it was necessary to vacuumize the heat exchanger first. In order to establish the required pressure condition

Description and Analysis of Adsorption Heat Storage Device

623

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

in the reservoir, vacuum pump (6) and manometer (5) were used. Water was pumped from the reservoir. The water coming from the heat exchanger (11) was conveyed into drains (13). The water container (7) was placed on the balance (8) which enabled measuring of precise (exact) quantity of injected water in the evaporator. For the process of desorption, instantaneous water heater (12) was installed in order to enable hot water. At the time of charging and discharging the tank, Agilent 34970 was collecting the following parameters: • time – used PC internal clock, • inlet temperature of the water (1) in the heat exchanger (11), • outlet temperature of the water (2) from the heat exchanger (11), • the temperature (3) in the evaporator/condenser (14) and • the temperature in the adsorber or the temperature of the adsorbent (4). The temperature was measured with NiCr-Ni (Type K) temperature sensors, which were calibrated by standard procedure. Uncertainty of temperature measurement was found 0.4%. Pressure was measured with manometer with the uncertainty 1.6%. Mass flow rate was determined by weighting the mass in a given time interval and was found to be 5.1 kg/h. The uncertainty of the mass flow was 1.2%. By the heat flux calculus a temperature dependant specific heat c(T) was considered. The combined relative uncertainty of the heat flux of adsorption was found to be 1.8%. 5 RESULTS AND DISCUSSION 5.1 Adsorption In pursuit of the measurement of the adsorption phase (discharging tank), it was necessary to dry out the tank interior first. Then 5700 g of dry aluminosilicateHX13 was put into the adsorbent basket. After the tank was properly assembled and sealed, vacuum pump sucked air from the tank to 20 mbar of absolute pressure. Water mass flow through the heat exchanger was set to 5124 g/h at temperature of 22 °C (Fig. 12). Agilent 34970 started with measurements of temperatures in duration of 24 h. As seen in Fig. 13, when the valve for dispensing the water opened, water in tank evaporated and was adsorbed on the surface of the adsorbent. As consequence, the temperature of the adsorbent has rapidly (after 3 min and 15 s) risen up to 50 °C and 624

soon reached maximum at 54.9 °C. As seen in Fig. 13, when the valve for dispensing the water opened, water in tank evaporated and was adsorbed on the surface of the adsorbent.

Fig. 12. Heat storage tank

As consequence, the temperature of the adsorbent has rapidly (after 3 min and 15 s) risen up to 50 °C and soon reached maximum at 54.9 °C. Later, the temperature of the adsorbent was gradually decreasing, probably due to the degree of saturation of the adsorbent. The temperature of the adsorbent reached a common point with the temperature of the water flowing from the heat exchanger after about three hours. The situation after this point is maintained until the end of measurement. The temperature of the adsorbent is lowered to 23.1 °C, which means that the water vapor remained adsorbed, but in very small quantities. The difference between the entry of water into the exchanger, and the adsorbent after 24 h was 1.3 °C. The temperature of the water which was entering the heat exchanger was practically constant throughout the measurement. The temperature of the water leaving the heat exchanger was, as expected, gradually increased, primarily as a result of adsorption heat, which was transferred through a pipe of heat exchanger. Maximum temperature 43.1 °C was achieved after 22 min and 45 s. Due to a temperature drop of the adsorbent, the amount of heat transferred was less than at the beginning. Therefore the water temperature at the exit from the exchanger is lowered. The downward trend curve is maintained until the end of the measurement.

Stritih, U. – Bombač, A.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

heat exchanger. After this point, the heat flux started to fall all the way to the end of the measurement. At the end of the measurement heat flux 0 W should be obtained, but because of temperature settings which shuted-off after 24 h, the temperature difference between the inlet and outlet of the water in the heat exchanger was still showing some minimal ΔT. The average heat flow was found to be 21.8 W.

Fig. 14. Heat flux from heat exchanger

Heat flow multiplied by the time gives the heat obtained. Instantaneous heat flow was multiplied with duration of Δt = 15 s according to Eq. (3): Fig. 13. Temperature change during adsorption mode

From the curve, which shows the temperature of the water in the evaporator, it can be seen that the temperature of the water very quickly fell to 8.3 °C which is the minimum. Soon after this bending point occurs, the temperature quickly rises to 14.5 °C and then to the end of measurements varies between 13 and 15 °C. The rapid temperature drop in the beginning was probably due to intensive evaporation. When the water is evaporated at such a low temperature, its inner energy is used and that lowers the water temperature in the evaporator. Due to establishing a steady state in the container after intensive evaporation of the water, bending occurs. At a temperature of 14.5 °C this variation was stabilized and varied with the difference of 2 °C until the end of measurement. Using the obtained data and Eq. (2):

Q i = m i ⋅ c p ,i ⋅ ∆Ti , (2)

the instantaneous heat flow is shown in Fig. 14 shows that heat flux increased proportionally with the increase temperature of the water leaving the heat exchanger. The maximum heat flux 123 W was reached at the maximum temperature difference (ΔT) between the water inlet (Tex,i) and outlet (Tex,o) of the

Qi = Q i ⋅ ∆t , (3)

The total heat from the tank represents the sum of Qi according to Eq. (4) for the observed time steps, where final time step is n = 5760.

Qtotal = ∑ i Qi . (4) n

The resulting total heat as a function of time is depicted in Fig. 15. As can be seen from the gradient dQ / dt the most of the heat was obtained at the beginning. After about 1.5 hour the gradient started to decrease gradually. In a 24 hours period the tank managed to transfer the total heat of 1.83 MJ or 0.508 kWh. The specific heat/ volume is according to volume of the heat exchanger (0.0104 m3) equal to 48.9 kWh/m3. According to work [6] testing the ability of energy storage in latent and sensible heat exchanger (at an initial temperature difference of 40 °C between the inlet and the outlet of the storage tank) showed the following results: • latent heat storage, 63 kWh/m3 and • water sensible heat storage, 52 kWh/m3. As can be seen the adsorption heat storage fall into the class of water sensible heat storage, but it is a bit better, also according to data found in Yu et al. [8] where similar results with aluminosilicate can be

Description and Analysis of Adsorption Heat Storage Device

625

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

found (i) type SG-127B = 24.9 kWh/m3 and (ii) type SG-LE32 = 14.9 kWh/m3.

Fig. 15. Heat from the heat exchanger

There are several reasons that we must take into consideration. One of the most influential reasons is ability of adsorber to adsorb water vapour. Adsorption heat decrease with the degree of saturation of the adsorbent and therefore is lower. Adsorbent which was used was probably not developed for heat storage purpose, so to achieve better results of adsorption mode another adsorber with higher adsorption properties should be used, developed in last two or three years. So, due to its low energy density using the aluminosilicate HX-13 in closed sorption thermal storage systems is not competitive for short-term applications; but when it used for longer period storage, aluminosilicate still possesses the heat storage ability in same extent and its energy density could be higher than sensible storage of water. 5.2 Desorption In desorption mode, it was necessary to dry out the adsorbent. For this purpose the incoming water to the heat exchanger was heated to a higher temperature level. Water mass flow was 41.7 kg/h at temperature 92.7 °C, the main parameters were measured 72 hours. Fig. 16 shows the changes in temperature as a function of time. Strong temperature fluctuations can be seen at the beginning of the measurement due to unsteady flow and unsteady heating. Later, temperature stayed still at 95 °C. As expected, the temperature got closer to the inlet temperature, and was then stabilized at a certain ΔT related to the inlet temperature, until the heating was turned off. The temperature of the heat sink should be equated with the inlet, but did not because the temperature sensors have not been calibrated to such a high temperature and therefore exhibit a certain error. Rapid fall of inlet temperature was caused with late valve closing. 626

Later, temperatures got closer to temperature of surroundings. The temperature of the adsorber and adsorbent temperature was increasing quite rapidly and reached maximum at the 85 °C. This temperature was maintained until the heating was stopped (after 4 hours) and then slowly falling to the ambient temperature. Adsorbent cooled down much more slowly than the water in the heat exchanger, probably because the adsorbent was in a vacuum, which served as an insulator. After the desorption mode, the adsorption mode should start again, but the pressure in the tank increased from the initial 20 to 100 mbar of absolute pressure.

Fig. 16. Temperature change during desorption mode

The results of desorption was evaluated by weighting of removed moisture from the adsorbent. The tank drained 4758 g of water. The mass of dry adsorbent was known, moistened mass also, so the difference was 1150 g of water in the adsorbent. Summing these two masses gives 5908 g of water in the tank after both (two) modes. This result should be matched with the mass of water which has been dosed at the start of measurement (6032 g), but it does not. The mass difference of 124 g probably remained on the tank and bottom walls. The temperature during both weighing was the same (21.5 °C). The mass difference was 2%, which is acceptable. From the measured data it is now possible to determine the level of moisture in the adsorbent. Assuming that the adsorbent was completely dry at the beginning, results show that the adsorbent contained about 20.2% moisture after the desorption, so minimal or almost no adsorbed water vapour was desorbed in the desorption. This probably happened because during the process of desorption, walls of the tank should be cooled in order to carry out intensive condensation. The results would be even better if the water in condenser was cooled. Another very important feature is the fact that a vacuum and adsorbent inside the tank are both

Stritih, U. – Bombač, A.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

insulators. Heat transfer from the heat exchanger to the adsorbent was most likely due to attenuation only at the junction and elsewhere already negligible, therefore the water was evaporated from only a small part of the adsorbent. We can achieve very good results by a slightly different design of the heat exchanger, where the emphasis would be right at the interfaces of the heat exchanger and the adsorbent.

pe QAds QDes QC QE Qi QLoss Qtotal

evaporation pressure [Pa] adsorption heat [J] desorption heat [J] condensation heat [J] evaporation heat [J] heat of individual point [J] heat loss [J] total heat from heat exchanger [J]

6 CONCLUSIONS

Q i

heat flux [W]

For efficient utilization of solar heat energy, compact and cost effective seasonal thermal storage systems are essential. One of the options is to use sorption heat storage principle. Heat storage tank with aluminosilicate HX-13 as adsorbent has been designed. This material is a product of Slovenian company. As for the construction, there were a lot of challenges with the under-pressure in the tank which were successfully exceed, but it should be noted that in the implementation of further attempts is to avoid sealing tape, which has proved to be inefficient when it is necessary to keep a vacuum in a container for several days. However, it should be noted that the results of the adsorption mode would be better if the basket would be totally spilled with an adsorbent. Because of the absence of the adsorbent, the other two tubes of heat exchanger were revealed, which means that during the adsorption mode the water vapor raised above the basket with revealed tubes, and cooled water in them. As a result, the temperature of the water at the exit from the heat exchanger was lower, with maximum 43.1 °C and was than lowering until the end of the measurement to approximately 25 °C. Consequently, the resulting heat generated from the heat storage was lower. It was demonstrated that the adsorption heat storage tank actually works in practice. In 24 h the heat from the heat exchanger was 1.8 MJ. In addition, it has one outstanding characteristic - heat can remain stored in the tank as long as adsorbent and adsorbate are separated and this is a huge advantage over sensible and latent heat storage. 7 NOMENCLATURE cp Δm m vapor m vapor pc

specific heat [J/kgK] mass change [kg] mass flow [kg/s] vapor mass flow [kg/s] condensation pressure [Pa]

heat flux of individual point [W] Q i TAds1 temperature at the beginning of adsorption [K] TAds2 temperature at the end of adsorption [K] TDes1 temperature at the beginning of desorption [K] TDes2 temperature at the end of desorption [K] inlet temperature of heat exchanger [K] Tex,i outlet temperature of heat exchanger [K] Tex,o Δt time interval [s] ΔT change in temperature [K] internal energy of vapor [J] Uv minimal saturation (kgAdsobat /kgAdsorbent) xmin Δx saturation change (kgAdsobat /kgAdsorbent) water heater WH 8 REFERENCES [1] Majdič, F., Bombač, A. (2014). Work efficiency of the new water hydraulic piston-type gas accumulator. Ventil, vol. 20, no. 2, p. 118-124. (in Slovene) [2] Bombač, A., Šelih, Z. (2011). Thermodynamic analysis by sulphuric acid production in sorption tower. Ventil, vol. 17, no. 3, p. 178-184. (in Slovene) [3] Jaehnig, D., Hausner, R., Wagner, W., Isakkson, C. (2006). Thermochemical Storage for Solar Space Heating in a Single-Family House. AEE – Institute for Sustainable Technologies, Gleisdorf. [4] Stritih, U., Osterman, E., Evliya, H., Butala, V., Paksoy, H. (2013). Exploiting solar energy potential through thermal energy storage in Slovenia and Turkey. Renewable & Sustainable Energy Reviews, vol. 25 p. 442-461, DOI:10.1016/j.rser.2013.04.020. [5] Wang, D.C., Li, Y.H., Xia, Y.Z., Zhang, J.P. (2010). A review on adsorption and refrigeration technology and adsorption deterioration in physical adsorption systems. Renewable & Sustainable Energy Reviews, vol. 14, no. 1, p. 344-353, DOI:10.1016/j.rser.2009.08.001. [6] Stritih, U., Butala, V. (2011). Energy savings in building with a PCM free cooling system. Strojniški vestnik - Journal of Mechanical Engineering, vol. 57, no. 2, p. 125-134, DOI:10.5545/sv-jme.2010.066. [7] Osterman, E., Tyagi, V.V., Butala, V., Rahim, N.A., Stritih, U. (2012). Review of PCM based cooling

Description and Analysis of Adsorption Heat Storage Device

627

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 619-628

technologies for buildings. Energy and Buildings, vol. 49, p. 37-49, DOI:10.1016/j.enbuild.2012.03.022. [8] Yu, N., Wang, R.Z., Wang, L.W. (2013). Sorption thermal storage for solar energy. Progress in Energy and Combustion Science, vol. 39, no.5 p. 489-514, DOI:10.1016/j.pecs.2013.05.004. [9] Ristić, A., Maučec, D., Henninger, S.K., Kaučič, V. (2012). New two-component water sorbent CaCl2FeKIL2 for solar thermal energy storage. Microporous and Mesoporous Materials, vol. 164, p. 266-272, DOI:10.1016/j.micromeso.2012.06.054. [10] Ristić, A., Zabukovec Logar, N., Henninger, S.K., Kaučič, V. (2012). The performance of small-pore microporous aluminophosphates in low-temperature solar energy storage. Advanced Functional Materials, vol. 22, p. 1952-1957, DOI:10.1002/adfm.201102734. [11] Hauer, A. (2007). Sorption theory for thermal energy storage. Paksoy, H.Ö. (ed.) Thermal Energy Storage for Sustainable Energy Consumption, Fundamentals, Case Studies and Design, NATO Science Series, vol. 234, p. 393-408. [12] Wang, D.C., Li, Y.H., Li, D., Xia, Y.Z., Zhang, J.P. (2010). A review on adsorption refrigeration technology and adsorption deterioration in physical adsorption systems. Renewable and Sustainable Energy Reviews, vol. 14, no. 1, p. 344-353, DOI:10.1016/j. rser.2009.08.001. [13] Jänchen, J., Ackermann, D., Stach, H., Broesicke, W. (2004). Studies of the water adsorption on zeolites and modified mesoporous materials for seasonal storage of solar heat. Solar Energy, vol. 76, no. 1-3, p. 339-344, DOI:10.1016/j.solener.2003.07.036. [14] Aristov,Y.I. (2013). Challenging offers of material science for adsorption heat transformation: A review. Applied Thermal Engineering, vol. 50, no. 2, p. 16101618, DOI:10.1016/j.applthermaleng.2011.09.003. [15] Cot-Gores, J., Castell, A., Cabeza, L.F. (2012). Thermochemical energy storage and conversion: a-state-of-the-art review of the experimental research under practical conditions. Renewable and Sustainable Energy Reviews, vol. 16, no. 7,p. 5207-5224, DOI:10.1016/j.rser.2012.04.007. [16] Abedin, A.H., Rosen, M.A. (2012). Close and open thermochemical energy storage: energy- and exergybased comparisons. Energy, vol. 41, p. 83-92, DOI:10.1016/j.energy.2011.06.034. [17] Restuccia, G., Freni, A., Vasta, S., Aristov, Y. (2004). Selective water sorbent for solid sorption chiller: experimental results and modelling. International Journal of Refrigeration, vol. 27, no. 3, p. 284-293, DOI:10.1016/j.ijrefrig.2003.09.003. [18] Glaznev, I., Ponomarenko, I., Kirik, S., Aristov, Y. (2011). Composites CaCl2/SBA-15 for adsorptive

628

transformation of low temperature heat: Pore size effect. International Journal of Refrigeration, vol. 34, no. 5 p. 1244-1250, DOI:10.1016/j.ijrefrig.2011.02.007. [19] Critoph, R.E. (2001). Simulation of a continuous multiple-bed regenerative adsorption cycle. International Journal of Refrigeration, vol. 24, no. 5, p. 428-437, DOI:10.1016/S0140-7007(00)00026-8. [20] Khattab, N.M. (2004). A novel solar-powered adsorption refrigeration module. Applied Thermal Engineering, vol. 24, no. 17-18, p. 2747-2760, DOI:10.1016/j.applthermaleng.2004.04.001. [21] Ziegler, F. (2009). Sorption heat pumping technologies: Comparisons and challenges. International Journal of Refrigeration, vol. 32, no. 4, p. 566-576, DOI:10.1016/j. ijrefrig.2009.03.007. [22] Ziegler, F. (2002). State of the art in sorption heat pumping and cooling Technologies. International Journal of Refrigeration, vol. 25, no. 4, p. 450-459, DOI:10.1016/S0140-7007(01)00036-6. [23] Henninger, S.K., Munz, G., Ratzsch, K.-F., Schossig, P. (2011). Cycle stability of sorption materials and composites for the use in heat pumps and cooling machines. Renewable Energy, vol. 36, no. 11, p. 30433049, DOI:10.1016/j.renene.2011.03.032. [24] Daßler, I., Mittelbach, W. (2012). Solar cooling with adsorption chillers. Energy Procedia, vol. 30, p. 921929, DOI:10.1016/j.egypro.2012.11.104. [25] Reddy Meenakshi, R., Nallusamy, N., Prasad Anjaneya, B., Reddy Hemachandra, K. (2012). Thermal energy storage system using phase change materials: constant heat source. Thermal Science, vol. 16, no. 4, p. 10971104, DOI:10.2298/TSCI100520078R. [26] Kashani S., Lakzian E. (2014). Numerical analysis of melting of nano-enhanced phase change material in latent heat thermal energy storage system. Thermal Science, vol. 18, suppl. 2, p. S335-S345 , DOI:10.2298/ TSCI111212163K. [27] Schmidt, F.P. (2004). Optimizing Adsorbents for Heat Storage Applications: Estimation of Thermodynamic Limits and Monte Carlo Simulations of Water Adsorption in Nanopores. Inaugural-Dissertation, Albert-Ludwigs-University, Faculty of Mathematics und Physics, Freiburg. [28] Jähnig, D., Wagner, W. (2004). Monitoring and Evaluation of the 1st Generation System Installed in Austria. Institute for Sustainable Technologies, Gleisdorf. [29] Hug, R. (1999). Summer Sun all Winter: Seasonal Heat Reservoir covers heating needs up to 100%. Solar Magazine, ISE Fraunhofer Institute for Solar Energy Systems, Freiburg im Breisgau.

Stritih, U. – Bombač, A.

Received for review: 2014-01-30 Received revised form: 2014-05-14 Accepted for publication: 2014-06-16

Design of a Prototype System Operant in Lunar Environment Ambu, R. – Manuello Bertetto A. – Falchi, C. Rita Ambu* – Andrea Manuello Bertetto – Costantino Falchi University of Cagliari, Department of Mechanical, Chemical and Materials Engineering, Italy In this paper, the design of a prototype system developed for a rover intended for the removal and transport of rocks on lunar soil is reported. The part of the rover dedicated to some of the main tasks, i.e. the lifting of objects and moving on rugged terrain, while controlling of the balance of the vehicle, is considered. These tasks are accomplished through the mechanical components assembled in a column connected to wheels. The study has been conducted with the aim of obtaining a simple and lightweight structure satisfying the requirements necessary to operate on the lunar soil. After a description of the architecture of the rover, the layout of the components of the column is detailed. A compliant, spiral-spring wheel is proposed to complete the mobility system. The primary components of the column are then structurally assessed by means of FEM numerical simulations. A numerical model of the wheel has also been implemented, in order to define in detail the wheel geometry and performance. The proposed layout could be promising for lunar applications since it has a configuration suitable for the specific characteristics of the environment it has to operate. Keywords: design, working operation, lunar environment, column, wheel, FEM simulation

0 INTRODUCTION There has long been great interest in the exploration of the moon. Different missions have been carried out over the years, including the Apollo missions [1], which enabled the collection of a large amount of data, including the physical and chemical characteristics of the environment and the configuration of the soil [2]. The information collected during the lunar missions has also enabled the possibility of establishing installations with human presence on the planet for the potential advantages that this may offer, including space tourism, which has been demonstrated to be an economically viable activity [3]. Several types of structures have been proposed for lunar outposts, ranging from simple inflatable structures to more advanced projects with complex buildings and supporting infrastructure [4] to [6]. However, the installation of a facility on the moon will require extensive operations for proper site preparation, including terrain excavation and transportation of rocks; furthermore, complementary mining operations can be advantageous since construction materials would be prepared from local materials and life support materials such as oxygen and water may be provided by indigenous resources [7]. The essential tasks of the devices required for working operations are those typical of construction machines, but different design solutions are necessary to create mechanical devices that can efficiently operate on the lunar soil.

Generally, the design of the components of a lunar vehicle has to take into account the peculiar characteristics of the fine and dry soil called lunar regolith, as well as those of the environment in which it will have to operate, e.g. temperature and radiation effects [8]. Previous lunar rovers used in exploration missions were assigned to transport crew, exploring the terrain and collecting the material samples. These vehicles, the Lunokhod and the Apollo Lunar Roving Vehicle (LRV) [9], were primarily designed to move in dusty environments. In the design of these vehicles, primary attention was given to the mechanical components that directly interact with the soil; in the development of rovers assigned to working missions, in addition to mobility, characteristics such as the load-bearing and carrying capability in a hostile environment also have to be accurately considered, while simultaneously keeping the structure simple and versatile. In recent years, several robot prototypes have been developed specifically for lunar excavation [10] to [12] and in situ resource utilization (ISRU) [13] and [14], which have varied widely in weight and tooling configuration. Bucket-wheel excavators have been shown to produce low-resistance forces when taking only small portions of regolith at a time; however, by repeatedly taking a large number of these small portions, high production is maintained. This paper deals with the research relative to the vehicles assigned to working missions on the lunar soil. In particular, it reports the design of a basic part of a rover, under development, conceived

*Corr. Author’s Address: University of Cagliari, Department of Mechanical, Chemical and Materials Engineering, Via Marengo 2, 09123 Cagliari, Italy, ambu@unica.it

629

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

to clear rocks from areas allocated to buildings or infrastructure. The device discussed in this paper is conceived as an auxiliary device to work on the lunar soil together with other machines; therefore, a direct comparison with other proposed design solutions of working devices, such those in references from 10 to 14, is difficult or inappropriate. The primary job of this rover is to grab rocks and move them, and comprises a gripping mechanism and four columns on which the wheels are assembled. In this paper, the components of the column have been considered, with particular attention to the lifting mechanism. The architecture of the wheel has also been developed. This study has been conducted via an analysis of a suitable arrangement and the choice of the materials of the components to obtain a structure to optimally operate in the lunar environment, while simultaneously minimizing size and mass. Parametric computer-aided design (CAD) modelling was used to determine the geometry of the components while the structural performance of the most significant parts under extreme loading conditions was assessed by means of finite element modelling (FEM) numerical simulations. Analogous methods were employed to define the characteristics and the performance of the wheel. A prototype of the column and the wheel was finally manufactured to verify the proposed configuration. The main aim of this paper is to propose and discuss a feasible configuration of a subsystem of a lunar rover, establishing suitable geometry for the components and quantitatively verifying with numerical simulations the stress level and loadbearing capability. Since the vehicle project is an extensive study, the results reported here are focused on a particular part of the overall design.

The vehicle specifications are reported in Table 1, which summarizes the main geometric and performance characteristics of the vehicle. Table 1. Vehicle specifications Length [mm] Width [mm] High [mm] Mass [kg] Avg. speed (in plane) [m/s] Max. speed (in plane) [m/s] Maximum terrain slope [°] Number of wheels [-]

1855 918 1627 200 0.5 1 30 8

The rover is designed to grasp and lift objects with weighs up to 800 N, corresponding to a mass of about 500 kg in the lunar gravity. A schematic diagram of the vehicle with the main components is reported in Fig. 1.

1 OVERVIEW OF THE ROVER The design of the rover was aimed to obtain a functional structure able to accomplish the working tasks described in the introduction, while simultaneously satisfying general requirements. A primary requirement is relative to the vehicle mobility since the design should satisfy specific geographical conditions, such as the ability to cross obstacles. Furthermore, it has to have sufficient endurance for working operations, including reliability and robustness in the environment. Thirdly, overall system efficiency requirements, such as low weight and compact size, must be satisfied. 630

Fig. 1. The architecture of the rover

It comprises a rigid frame to which the gripping mechanism of the load and four columns supported on the wheels are connected. The columns transmit the load to the frame by means of slides assembled to the frame. The gripping system [15], whose detail in a grasping position is also reported in the lower part of Fig. 1, is a plane-articulated mechanism with three degrees of freedom. The layout this mechanism allows it to grasp solid objects of different shapes, such as the numerous rocks on the lunar surface.

Ambu, R. – Manuello Bertetto A. – Falchi, C.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

This architecture makes the rover easily transportable and, possibly, adapted to reuse in the lunar base for the transportation of various objects. 2 DESCRIPTION OF THE LAYOUT OF THE PROTOTYPE SYSTEM The columns of the rover have been designed to accomplish different essential tasks for correct operations. The requirements that the proposed configuration will have to satisfy are to propel the vehicle, lift the load, and control the balance in relation to the contour of the soil. In this paragraph, the geometric configuration of the components is described, while the structural analysis is reported later. The paper is focused on the geometrical configuration and structural performance, while a future active control to steer the vehicle trim is foreseen. However, the fundamental issue of control strategies is behind the aim of this paper and is not discussed here. The prototype column of the rover is shown in Fig. 2; at the right side, a detail of the lower part is also reported, showing the assemblage of the wheels and the motors.

Fig. 2. The column of the rover

The lifting mechanism, which controls the shift of the load and enables the vehicle to remain on course, comprises a slide joined to two linear guides

with a low coefficient of friction, a precision coupling between a lead screw and nut driven by an electric motor and a particular device, which will be later described in detail, designed for the correct operating of the screw and nut mechanism. The electric motor used for the lifting mechanism is a high efficiency brushless motor with an electrical power of about one kilowatt and a mass less than one kilogram. The shaft of the electric motor is connected to a series connection of two reduction gears, an epicyclical reduction gear and a worm screw gear, with an overall speed ratio of 30. This layout allows choosing the position of the frame of the rover according to the size and geometry of the rock. Furthermore, since each column can be driven independently, the vehicle can cover rugged terrain and cross obstacles as large as the diameter of its wheels maintaining a proper balance, thus restricting the likelihood of overturning during motion on rough terrain. The correct operating of the screw and nut coupling under load is assigned to a decoupling joint that is inserted between the screw and nut mechanism and the slide.

Fig. 3. A detail of the column; a) the plate and the screw and nut mechanism, and b) the components of the decoupling device and the nut

Fig. 3a shows the portion of the column with the slide and its connection to the screw and nut

Design of a Prototype System Operant in Lunar Environment

631

StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

mechanism; the area that includes the joint has been highlighted in the figure. The components of this device and the nut are assembled in a compact apparatus, since the size and shape of each part were chosen so as to minimize the overall volume. The exploded view shown in Fig. 3b shows the components of the assembly between the joint and the nut. In particular, with regards to Fig. 3b, the joint (1) is rigidly connected to the slide, and it is simply placed on the flange of the nut (2) interposing a meniscus (3) between the two parts to reduce the friction between the surfaces. These components are finally assembled, introducing another similar meniscus (5) at the opposite end of the joint, and a proper counterpart (6) fastened to the flange of the nut by means of a screw. The joint, whose geometry has been accurately optimized [16], was designed for transmitting to the lead screw only the normal component of the load, simultaneously transferring the other components to the load-bearing structure. In fact, the geometry and the assembly conditions give this component the capability of both rotating about an axis perpendicularly to the axis of the screw and of sliding in a direction orthogonal to the same axis, also compensating the potential misalignment between the axis of the screw and the axes of the rails fixed to the column. The arrangement of the components of the lower part of the column was intended to be a system in which the transmission, suspension and motorization mechanisms are integrated into a compact apparatus. The motion of the rover and the selection of the trajectory are obtained by means of a differential traction that controls a pair of wheels driven by an electric motor. The electric motor chosen for the motion of the rover has specifications analogous to that used for the lifting mechanism, satisfying the mobility requirements, as well as those of low mass and size. The gears that drive the wheels have an overall speed ratio of 960 between the motor shaft and the wheels, for an efficiency ratio of about 0.30. This speed ratio allows it to have an irreversible system by means of a kinematic chain with a worm gear and an epicyclical gear. The design of the architecture of the wheel was conducted starting from a preliminary analysis of the different configurations used in previous lunar missions. These spread from the wheel made of a rigid structure with a wire carcass connected by spokes used in the Lunokhod vehicle to a flexible wheel with a wire mesh carcass and a stiff inner frame employed in the LRV [17]. Apart from the two cited, the other 632

wheel used on the moon soil, which operated on the mobile equipment transporter (MET), was made of rubber. Following the general requirements derived from the previous practice, the wheel architecture also was developed taking into account the specific requirements of the proposed vehicle and a more recent study on non-pneumatic wheels [18] whose design also benefits of the use of advanced composite materials instead of the traditional ones. The conceived wheel is shown in Fig. 4.

Fig. 4. The wheel

The essential components are a central hub, a series of spiral springs, geometrically shaped as bitangent semicircles, and a tread which is composed of non-rigid triangular elements with straight sides. The triangular shape was the only geometry considered and was chosen for its efficiency in transmitting the driving and braking torque. The spokes are fixed to the hub and to a rigid rim. This configuration allows the wheel to perform different tasks, such as those of elastic and dissipative suspension and the reduction of the effects of the local asperities of the soil on the vehicle. The spiral springs act as elastic elements in order to provide both load bearing and shock absorption capabilities. The triangular elements of the tread undergo local deformation when the rover crosses an asperity of the rugged terrain, thereby avoiding transferring possible stresses to the structure. 3 NUMERICAL ASSESSMENT OF THE PROPOSED LAYOUT 3.1 The Lifting Mechanism The components directly involved in the lifting of the weights and the balance of the vehicle were considered in order to evaluate the performance of the column.

Ambu, R. â&#x20AC;&#x201C; Manuello Bertetto A. â&#x20AC;&#x201C; Falchi, C.

StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

In particular, the lead screw and nut coupling and the connection of this mechanism to the slide, which was accomplished by means of the decoupling joint, have been analysed in detail. A standard Acme thread profile was selected for the screw and the nut with a nominal screw diameter of 15.5 mm. The nut was designed in two parts, both depicted in Fig. 3b, consisting of a prism-shaped part internally threaded, labelled as (4) in the figure, and a plane element, labelled as (2), which serves as the flange. The use of a nut made into two parts is also advantageous for manufacturing purposes. The lead screw and nut coupling was structurally verified by means of finite element analysis. The parametric CAD modelling used for the design of each component, among the many advantages, also allowed to perform the structural validation of the real geometries. In fact, the effects of the lead angle and the helix of the threaded profile, neglected in axisymmetric finite element models, can be taken into account. The finite element model of the two parts, implemented in a FEM commercial software (ANSYS), was obtained via 142,133 3D-brick elements, while contact elements were introduced at the interface between the threaded surfaces of the screw and the nut. Fig. 5 shows the FE model obtained.

MPa, Î˝ = 0.3) lead screw. The numerical simulations were relative to an extremely heavy limit condition. In fact, during the working operations on the lunar soil, an unexpected balance condition of the vehicle can occur when it can accidentally be sustained only on two of the four columns. If this happens while the rover is loaded at its upper limit, the two operating columns are forced to carry the maximum load. The stresses on the mechanism in this condition were evaluated, giving particular attention to the analysis of the nut, since it is the weakest component of the coupling. As for the lead screw, the maximum value of the Von Mises stress obtained was 292 MPa, about the 30% of the ultimate stress for the chosen material. Fig. 6 reports an iso-colour representation of the Von Mises stress distribution in the nut, expressed in MPa.

Fig. 6. Von Mises stress distribution in the nut

Fig. 5. FE model of the screw and nut mechanism

A composite material with a thermoplastic matrix (Nylon6/6) reinforced with 30% carbon fibres and 5% PTFE was chosen (E = 14500 MPa, Î˝ = 0.3) for the nut, and it was coupled with a titanium alloy (E = 111000

The component exhibits the presence of stress concentrations along with the edges of the threaded surfaces, especially in correspondence to the threads next to the outer side, but with a maximum stress that is under the ultimate stress value (221 MPa) of the material. To further validate the proposed geometry, the distribution of the axial load on the threads of the nut was investigated. In the design of screw and nut transmission mechanisms, it is usually assumed that the axial load is equally distributed on the threads. However, the load distribution on the threads can be affected by the nut contour, size, stiffness, and other factors [19]. For this purpose, a parameter, i.e. the ratio of flank load in percentage (RFL%) that represents the load carried by each thread, was calculated.

Design of a Prototype System Operant in Lunar Environment

633

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

Fig. 7. Axial load distribution on the threads of the nut

Fig. 7 depicts a diagram that summarizes the results obtained. The diagram depicts on the vertical axis the parameter RFL%, while on the horizontal axis the thread number starting from the nut fixed surface, corresponding to the flange surface, is reported. A trend line has also been added to the plot. The load distribution exhibits higher values at the outer engaged threads; however, the maximum difference between the higher and lower value is rather limited (2.6%), thus assessing the proposed geometry. As described in the previous paragraph, the screw and nut mechanism was completed with a decoupling device. The functional parameters of this device were first evaluated. It was accomplished by a mechanical variation analysis of the assembled device evaluated at the limit positions by means of a tolerance analysis software. In particular, since the joint was conceived to allow linear and angular adjustments, the maximum linear displacement and the maximum angle of rotation were estimated. As for the linear displacement, a value of about 2.5 mm was obtained, while a maximum angle of 4° was estimated. These values, even if limited, can be considered satisfactory, since the adjustment of the position of the slide will mainly occur via a combined variation of the two parameters. To assess the structural performance of the decoupling device, it was necessary to take into account the interaction of the joint with the meniscus interposed between its surface and the nut, thus considering in the numerical simulations the joint and the meniscus as a coupled pair. The simulations were relative to the same extreme operating condition previously described for the screw and nut coupling, where only two columns are supposed to bear the overall load. The finite element model was made with 32,372 3D-brick elements, while contact elements were introduced at the interface between the joint and the 634

meniscus. The symmetry of the geometry was taken into account by modelling only one half of each part and introducing symmetry boundary conditions. The bottom surface of the meniscus was fixed while the load was applied at the end surface of the joint. An aluminium alloy (E = 69000 MPa, ν = 0.3) was the material chosen for the joint, while a thermoplastic polyamide (E = 2010 MPa, ν = 0.49) was selected for the meniscus. Fig. 8 shows the FE model of the coupled components. In the analysis of the results, particular attention was given to the meniscus, regularly allocated to assist the sliding of the joint, which is the weakest component between the two parts and prone to wear.

Fig. 8. FE model of the joint and meniscus

Fig. 9 shows an iso-colour representation of the Von Mises stress distribution relative to this component where the surface in contact with the joint is visible.

Fig. 9. Von Mises stress distribution in the meniscus

In Fig. 9, the local effect on the stress distribution of the peculiar loading condition can be observed; however, higher values are limited within a restricted area and do not exceed the maximum admissible value (104 MPa) for the material chosen.

Ambu, R. – Manuello Bertetto A. – Falchi, C.

StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

3.2 The Wheel The aim of the analysis was to determine the wheel characteristics satisfying the requirements with the minimum cost in terms of mass. The wheel configuration depends on different parameters. The values of some of them, such as the wheel external diameter and the wheel width, are dictated by the system requirements, while the values of other parameters can be chosen and varied appropriately. These last parameters include the number of springs, their curvature and wall thickness, as well as the materials of the wheel components, i.e. the tread, the hub and the springs. Different sets of values of these parameters have been considered, combining parametric CAD modelling and FEM analysis. The approach considered consists of the analysis of different configurations obtained by changing the values of the input parameters such as the geometry and the thickness of the spokes. The outputs of each FEM analysis considered are the wheel deformation and the internal stress. This procedure considered discrete values of the geometric parameters of the components and some materials compatible with the lunar environment. However, this procedure does not exhaust all prospective configurations. The result of the activity has been to precisely define a wheel configuration compliant with the system requirements.

The FE model of the wheel, reported in Fig. 10, was made with 39,812 3D-brick elements. A rigid plate was introduced in order to simulate the ground. Contact elements were inserted between the plate and the tread to hinder penetration between the wheel and the plate. A glass fibre reinforced epoxy composite material (E = 11900 MPa, Î˝ = 0.217) was chosen for the spiral spokes while the tread was made of rubber (E = 5 MPa, Î˝ = 0.48). In addition, an aluminium hub was used. A vertical load was applied on the wheel hub in order to simulate the load transfer on the wheel, while the plate was fixed. Since the rover is a low-speed vehicle, as reported in Table 1, only a static analysis has been examined. Suspensions allow effective behaviour at the interface with the ground, in order to have a valid exchange of forces with it, also in view of foreseen control strategies of trim and traction. Furthermore, a global analysis is reported in this study, since the entire system and the wheel subsystem are non-linear. Distinguishing the contribution of the different parts and taking into account each separately is considered to be less significant for the evaluation of the system. In any case, it is certainly important to evaluate the influence of each part on the global behaviour, but it has been not been considered here. Fig. 11 shows the vertical displacement obtained as a function of the vertical applied load, which allows to evaluate the trend of the wheel stiffness, i.e. the resistance of the wheel to the deformation when a vertical load is applied.

Fig. 10. FE model of the wheel Fig. 11. The force-displacement curve of the wheel

The wheel width is 150 mm and the outer diameter is 500 mm while the inner diameter of the tread has a value of 400 mm. The spokes, which consist of 5 radial elements, are curved surfaces with a radius varying between 119 and 126 mm. The hub diameter is 100 mm.

Two different positions that repeatedly alternate during the motion of the vehicle were considered. The first, corresponding to Curve 1 (solid line) in the figure, was relative to a vertex of a triangular element of the tread positioned on the vertical symmetry axis

Design of a Prototype System Operant in Lunar Environment

635

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

of the wheel, while Curve 2 (dashed line) corresponds to the vertical axis of the wheel positioned between two vertices of a triangular element. Curve 1 shows a linear trend in the range considered, while Curve 2 has an analogous tendency except for lower values of the vertical applied load. The difference between the values of the vertical displacements relative to the two positions decreases as the applied load increases; in particular, around the design load (365 N), the difference is less than 1 mm. A change of stiffness during rotation should be troubling regarding the transmission of vibrations to the system. However, it can be advantageous to achieve a greater efficiency in the transmission of the drive or braking torque to the ground on soft terrain, due to an effect similar to that of a groove of the tread. The implemented numerical model also enables evaluating basic parameters, such as the wheel-ground contact area and the contact pressure. The contact area under the nominal load can be estimated to be 5985 mm², yielding an average pressure of 64.89 kPa on the ground. Finally, the stress relative to the maximum load acting on the wheel, corresponding to the column sustained on only one wheel of the couple, was evaluated. Fig. 12 depicts an iso-colour representation of the Von Mises stress distribution obtained.

Fig. 12. Von Mises stress distribution in the wheel

The maximum value of the equivalent stress is under the limit stress value and occurs in a small zone at the interface between the rim and a spoke, suggesting further improvement of the joints between these components of the wheel.

636

4 CONCLUSIONS In this paper, a prototype system of a column and wheel, essential components of a rover allocated to working tasks on the lunar soil, is proposed and assessed via numerical simulations. The resulting structure of the vehicle is straightforward and lightweight, which offers many advantages both for shipping and during operation on the lunar soil. It can be adjusted to function on rough terrain thanks to the ability to traverse obstacles by longitudinally changing the position of each pair of wheels. In addition, severe loading conditions can be sustained by the main parts, as shown via the numerical simulations. Future work will be concerned with the experimental verification of the prototype rover in an environment simulating the characteristics of the lunar soil. 5 REFERENCES [1] Harland, D.M. (2008). Exploring the moon: the Apollo Expeditions. Springer-Praxis, Berlin, DOI:10.1007/978-0-387-73997-7. [2] Heiken, G.H., Vaniman, D.T., French, B.M. (1991). Lunar Sourcebook, A User’s Guide to the Moon, Cambridge University Press, Cambridge. [3] Collins, P. (2006). The economic benefits of space tourism. Journal of the British Interplanetary Society, vol. 59, p. 400-410. [4] Grandl, W. (2007). Lunar base 2015 stage 1 preliminary design study. Acta Astronautica, vol. 60, no. 4-7, p. 554-560, DOI:10.1016/j.actaastro.2006.09.031. [5] Jones, T.D. (2007). Homesteading the Moon. Aerospace America, vol. 45, no. 4, p. 12-15. [6] Benaroya, H., Bernold, L. (2008). Engineering of lunar bases. Acta Astronautica, vol. 62, no. 4-5, p. 277-299, DOI:10.1016/j.actaastro.2007.05.001. [7] Chamberlain, P.G., Taylor, L.A., Podnieks, E.R., Miller, R.J. (1993). A review of possible mining applications in space. Lewis J.S., Matthews, M.S., Guerrieri, M.L. (eds.). Resources of Near-Earth Space. The University of Arizona Press, Tucson, p. 51-68. [8] Pirich, R., Weir, J., Leyble, D., Chu, S., DiGiuseppe, M. (2010). Effects of the lunar environment on space vehicle surfaces. Proceedings of Long Island Systems, Applications and Technology Conference, Farmingdale, p. 1-6. [9] Young, A. (2007). Lunar and planetary rovers: the wheels of Apollo and the quest for Mars. SpringerVerlag, Berlin. [10] Skonieczny, K., Moreland, S.J., Wettergreen, D.S., Whittaker, W.L. (2011). Advantageous bucket-wheel configuration for lightweight planetary excavators.

Ambu, R. – Manuello Bertetto A. – Falchi, C.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 629-637

Proceedings of 17th International Conference of the International Society for Terrain Vehicle Systems, Blacksburg. [11] Johnson, L.L., King, R.H. (2010). Measurement of force to excavate extraterrestrial regolith with a small bucket-wheel device. Journal of Terramechanics, vol. 47, no. 2, p. 87-95, DOI:10.1016/j.jterra.2009.08.002. [12] Nasa (2013). Engineers building hard-working mining robot, from http://www.nasa.gov/topics/technology/ features/RASSOR.html, accessed on 2013-14-05. [13] Bartlett, P.W., Wettergreen, D., Whittaker, W. (2008). Design of the scarab rover for mobility and drilling in the lunar cold traps. Proceedings of International Symposium on Artificial Intelligence, Robotics and automation in Space, Hollywood. [14] Astrobotics (2013). Polaris, from http://astrobotic.net/ rovers/polaris, accessed on 2013-18-06. [15] Carbone, G., Falchi, C., Manuello Bertetto, A., Ceccarelli, M. (2012). Simulation of a gripping device for obstacle removing on lunar soil. Proceedings of 21st

International Workshop on Robotics in Alpe-AdriaDanube Region, Naples. [16] Ambu, R., Falchi, C., Manuello Bertetto, A. (2010). A lunar rover leg: optimal design of a decoupling joint. International Journal of Mechanics and Control, vol. 11, no. 1, p. 45-50. [17] Asnani, V., Delap, D., Creager, C. (2009). The development of wheels for the lunar roving vehicle. Journal of Terramechanics, vol.46, no.3, p. 89-103, DOI:10.1016/j.jterra.2009.02.005. [18] Rhyne, T.B., Cron, S.M. (2006). Development of a non-pneumatic tire. Tire Science and Technology, vol. 34, p. 150-169, DOI:10.2346/1.2345642. [19] Sun, Y., Zhou, X., Wei, L., Wang, W. (2009). Development of a new type of transmission screw nut with high efficiency and heavy duty characteristics. Proceedings of the Institution of Mechanical Engineers, Part C: Journal of Mechanical Engineering Science, vol. 223, no. 5, p. 1181-1189, DOI:10.1243/09544062JMES1332.

Design of a Prototype System Operant in Lunar Environment

637

Received for review: 2014-01-04 Received revised form: 2014-04-01 Accepted for publication: 2014-05-07

Wind Turbine Seismic Load Analysis Based on Numerical Calculation Jin, X. – Liu, H. – Ju, W.B. Xin Jin1,* – Hua Liu2 – Wenbin Ju1 1 Chongqing

University, College of Mechanical Engineering, China 2 Dongfang Electric Corporation, China

Large-scale wind turbines have come into common use in Europe. Because violent earthquakes are relatively rare there, insufficient consideration has been given to the seismic impact on the wind turbine specifications; however, at present, there are many wind farms being constructed in earthquake-prone regions, and the seismic impact cannot be ignored in wind turbine designs. Based on the multi-body system dynamic theory and taking into consideration the soil-structure interaction, this paper proposes a blade-cabin-tower-foundation coupled model in order to study the load-bearing conditions of wind turbines under seismic impact. According to the basic theory of multi-body system dynamics, the wind turbine blade and tower system comprises a series of continuous discrete units, while soil-structure interaction in the tower system is realized through the spring and damping set on the interface between the foundation and the soil body; the cabin is simplified as a rigid model. Based on the Eurocode 8 earthquake load spectrum, the dynamic response of a wind turbine working under seismic impact is analysed, and the seismic load is compared. Results of the study can serve as references for designing key parts and control strategies of wind turbines for earthquake-prone regions. Keywords: wind turbine, earthquake, multi-body dynamics, soil-structure interaction

0 INTRODUCTION AND BACKGROUND As the price of wind power drops, wind turbines are playing an increasingly important role in the global power supply. In recent decades, wind turbine systems have been mainly established in North Europe, which is not earthquake-prone (or merely subject to less violent earthquakes), and thus seismic impact has not been focussed on in the specifications for wind turbine systems [1]. However, increasing numbers of wind turbine systems are being constructed in earthquakeprone regions; therefore, it is necessary to analyse the dynamic response characteristics of wind turbine systems to earthquakes, and to consider the impact of seismic load in further engineering designs. There are some relevant studies regarding wind turbines. For example, Teng et al. [2] used empirical mode decomposition on the pitting fault detection of a wind turbine gearbox. Potočar et al. [3] used plasma actuators to control separation flow over a wind turbine blade. However, there are limited studies on wind turbines under seismic impact. The Risø National Laboratory of Denmark [4] analysed the seismic load on wind turbines by using the first-order natural vibration frequency. In 2002, Bazeos et al. [5] improved the model established by his predecessors and created a wind turbine tower model. The tower body was modelled to be of three sections, with the size gradually increasing from the top to the bottom; each section had the same size, with a progressive transition between adjoining sections. Bazeos et al. used the time-history method to analyse the dynamic 638

seismic load, and then considered the impact of soilstructure interaction, and finally concluded that the soil-structure interaction had an obvious impact on the overall system. In 2003, Lavassas et al. [6] proposed another wind turbine finite element model, in which the tower body was a truncated cone shell with a pile base at the bottom; in this model, the authors simulated the impact of soil-structure interaction by setting a contact element between the foundation and the soil body. The common point of the two models above was to model the tower body and the foundation, but not the blade and cabin; therefore, it was difficult to use them for the refined analysis and design of the wind turbine tower system. Murtagh et al. [7] and [8] proposed a shear transfer-based blade and tower coupling finite element model, having clarified the coupling mechanism of the blade and tower body and used the time-history method to analyse the dynamic wind load on the structure. Nevertheless, the authors did not model the foundation, nor considered the effect of soil-structure interaction. Witcher [9] studied the seismic load and applied the seismic analysis method for the bridge and building structures to wind turbines, but he did not consider the soil-structure interaction. Kang et al. [10] used a nonlinear spring to simulate the soil-structure interaction and analysed the reliability of offshore wind turbine bases, but the applied load was static. Harte et al. [11] studied the wind turbine response to wind-induced buffeting and considered the soil-structure interaction in the model, but he did not carry out a seismic analysis. Lombardi et al. [12] conducted a series of experiments on the

*Corr. Author’s Address: Chongqing University, College of Mechanical Engineering, Chongqing 400044, China, jinxin191@hotmail.com

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

soil-structure interaction of wind turbines, in which only the natural vibration frequency of the structure was considered; however, no consideration was given to the wind load borne by the wind turbines, nor was any seismic analysis carried out. It is easy to conclude that to date there are few studies on modelling bladetower-foundation coupled with a multi-body system of wind turbines, particularly by using the time-history method to analyse the dynamic seismic load based on the coupled model. In consideration of the above, this paper uses the multi-body system dynamic theory as a basis for studying the dynamic response of wind turbines to earthquakes, while taking into account the soilstructure interaction. According to the basic theory of multi-body system dynamics, the wind turbine blade and tower system comprises a series of continuous, discrete units, while the soil-structure interaction of the wind turbine tower system is realized through the spring and damping set on the interface between the foundation and the soil body. In order to study the dynamic characteristics of a wind turbine, the time-history method is used to analyse the dynamic seismic impact on the wind turbine based on Eurocode 8, and an analysis model is established to study the seismic impact on load-bearing conditions of the wind turbine during operation, so as to provide references for designing key parts and control strategies of wind turbines for earthquake-prone regions. 1 COORDINATE SYSTEMS Coordinate systems are critical in structural analysis. A proper coordinate system may function to simplify

calculations; therefore, it is necessary to establish a wind turbine coordinate system before establishing a structural analysis model. As required for analysis, a dynamic coordinate system for a three-blade wind turbine can be established, as shown in Fig. 1. Fig 1a shows the inertia system Z (orthogonal coordinate axes z1, z2 and z3), wind turbine base system A (orthogonal coordinate axes a1, a2 and a3), tower system T (orthogonal coordinate axes t1, t2 and t3), tower top system B (orthogonal coordinate axes b1, b2 and b3). Fig 1b shows the principal axes system C (orthogonal coordinate axes c1, c2 and c3), rotor azimuth system E (orthogonal coordinate axes e1, e2 and e3), hub system G (orthogonal coordinate axes g1, g2 and g3), cone angle system I (orthogonal coordinate axes i1, i2 and i3), blade pitch system J (orthogonal coordinate axes j1, j2 and j3), blade local torsion angle system Lj (orthogonal coordinate axes Lj1, Lj2 and Lj3), and blade element local torsion angle system N (orthogonal coordinate axes n1, n2 and n3) etc. The details of the coordinate system J, Lj and N are shown in Fig. 1c. 2 DISCRETENESS MODEL Suppose the wind turbine blade and tower system comprises a flexible cantilever with evenly and continuously distributed mass and rigidity, as shown in Fig. 2. Deformation of the continuous blade and tower system is expressed with a series of linearly overlapped normalized vibration modes; so that the freedom of the blade and tower is reduced from infinity to N (N is the supposed number of modes selected for calculation). Then the horizontal deformation u(z,t) of

Fig. 1. Coordinate systems; a) tower coordinate systems b) drivetrain coordinate systems c) detail coordinate system of blade Wind Turbine Seismic Load Analysis Based on Numerical Calculation

639

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

the flexible cantilever, at any time and any place, is a series of linearly overlapped normalized modes ϕa(z), which associate with the generalized coordinates qa(t) as follows:

where Qa is the deformation of the cantilever free-end in the free mode a. By applying qa(t) in the equation Cb(t)=Cm,bqm(t), and applying the result obtained from the foregoing equation in Eq. (5), we can obtain:

u ( z , t ) = ∑ φa ( z ) qa ( t ), (1)

N + p −1

∑ ( −ω m

a =1

where ϕa(z) is the ath mode of the cantilever and a function of the cantilever longitudinal distance z; the generalized coordinate qa(t) associating with the mode a is a function of time, and it is commonly the free end deformation of the cantilever corresponding to the mode.

j= p

)

+ kij C j = 0,

( i = p, p + 1,, N + p − 1) ,

(6)

which can be transformed to be a matrix as follows:

( −ω [M ] + [K ]){C} = {0} , (7) 2

where the generalized mass matrix [M] and the generalized rigidity matrix [K] are N×N order matrixes; the coefficient vector {C} is an N×1 order vector. By analysing the characteristic values of the matrix equation, we can obtain the characteristic value ω2a and the characteristic vector {C}a. 2.1 Tower Modelling Fig. 2. Discreteness model

By using the Lagrange equation of a conservative and constant system, the kinetic equation of N freedom systems can be expressed as: N + p −1

∑ j= p

mij c j ( t ) +

N + p −1

∑ j= p

kij c j ( t ) = 0,

( i = p, p + 1, , N + p − 1) ,

(2)

where the generalized mass mij and the generalized rigidity kij can be defined by using the kinetic energy T and potential V as:

1 N + p −1 N + p −1 ∑ ∑ mij ci ( t ) c j ( t ), 2 i= p j= p

1 N + p −1 N + p −1 ∑ ∑ kij ci ( t ) c j ( t ). (4) 2 i= p j= p

(3)

When the cantilever vibrates at a specific inherent frequency ωa, suppose a = m, we can obtain:  Q sin (ωa t + ψ a ) qa ( t ) =  a 0 

The tower is simulated to be an inverted cantilever whose free end is fixed with a point mass MTop, representing the total mass of the base plate, the cabin, the hub and the blade. The generalized mass of the tower can be expressed as:

a=m , (5) a≠m

mij = M Top + ∫ µT ( h )φi ( h )φ j ( h ) dh, (8) 0

where μT(h) is the mass distribution line density of the tower. The tower potential comprises VBeam associating with distribution rigidity of the beam and VGravity associating with gravity. V = VBeam + VGravity , (9)

where the potential component associating with the distribution rigidity of the cantilever can be expressed as: VBeam =

d 2φi ( h ) d 2φ j ( h )  1 N + p −1 N + p −1  H dh  ci ( t ) c j ( t ),  ∫0 EIT ( h ) ∑ ∑  2 i= p j= p  dh 2 dh 2 

(10)

where EIT(h) is the distribution rigidity of the tower and H is the height of the tower.

The potential component associating with gravity can be expressed as:

H dφ ( h ) dφ j ( h ) H  h dφi ( h′ ) dφ j ( h′ )   1 N + p −1 N + p −1  i VGravity = − g ∑ ∑  M Top ∫ dh + ∫ µT ( h )  ∫ dh′  dh  ci ( t ) c j ( t ), 0 0 0  2 i = p j = p  dh dh dh′ dh′   

640

Jin, X. – Liu, H. – Ju, W.B.

(11)

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

where the minus sign indicates that gravity will reduce the generalized rigidity of the tower; the first item in the bracket is associated with gravity of the tower mass, and the second item is associated with gravity of the tower distribution mass, with the impact of the tower deformation on the gravitational potential considered. The generalized rigidity of the tower can be expressed as: H

kij = ∫ EIT ( h ) 0

d 2φi ( h ) dφ dh

( h ) dh −

dh 2

H H dφ ( h ) dφ j ( h ) dh. (12) − g ∫  M Top + ∫ µT ( h′ ) dh′ i 0  h    dh dh

2.2 Soil-Structure Interaction Modelling The interaction between the soil and the foundation is an interactive force on the contact surface, which is caused by different material properties of the soil body and the structure (mainly the elastic modulus). In civil engineering, that the soil-structure interaction impacts the dynamic response of structures is a generally known phenomenon. For wind turbine towers established on soft soil (Fig. 3), the soilstructure interaction is considered to be one of the key factors in dynamic analysis, and a more critical factor for wind turbine tower systems in earthquake-prone regions.

interactions. However, the continuum medium model is too complex, while the finite element model used to simulate soil-structure interactions is overly time-consuming and is thus disadvantageous for calculation. Therefore, to comprehend the most essential characteristics of soil-structure interaction in wind turbine tower systems, one simple and effective approach is to set a spring and damping on the interface between the foundation and the soil body (Fig. 4). For a three-dimensional soil-structure interaction model, the two horizontal displacements couple with the rotary movement. However, the coupling item can be neglected as it is of a relatively small value [13]; this is particularly the case for wind turbine tower systems based in shallow soil (Fig. 2). Therefore, all dynamic components are considered to not be subject to mutual coupling. For a rigid round foundation, the rigidity and damping coefficient can be determined based on the properties of the surrounding soil body and the foundation size [14] to [16], which can be expressed as:

Fig. 3. Wind turbine system

Fig. 4. Dynamic model of soil-structure interaction

The continuum medium model (analytical method), the discrete model and the finite element model can be used to study complex soil-structure

8Gs Rs   kx = k y = 2 − µ s   4Gs Rs kz =  1 − µs  ,  3 k = k = 8Gs Rs β  α 3(1 − µ s )  16  3  kγ = 3 Gs Rs

(13)

 4.6 Rs2 Gs ρ s  cx = c y = 2−µ   3.4 Rs2 Gs ρ s  cz = 1− µ , (14)   0.4 Rs4 cα = cβ = Gs ρ s 1− µ   4  cγ = 1.11Rs Gs ρ s

where kx and ky are horizontal rigidity coefficients; kz is the vertical rigidity coefficient; kα and kβ are the bending rigidity coefficients; kγ is the bending coefficient about the vertical axis. Similarly, ci (i = x, y, z, α ,β ,γ ) is the corresponding damping coefficient. Rs is the round foundation radius; Gs, μs and ρs are the shear modulus, the Poisson’s ratio and the density of the soil body, respectively.

Wind Turbine Seismic Load Analysis Based on Numerical Calculation

641

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

2.3 Blade Modelling Each rotor blade is simulated to be a rotary cantilever that has a point mass MTip fixed at the free end and rotating about an axis perpendicular to the rotor plane at the angular velocity Ω. Suppose the flexible part of each blade moves independently flap-wise (perpendicular to the wing-type string) and span-wise (parallel with the wing-type string). Meanwhile, blade deformation can be expressed as deformation towards the in-plane direction (parallel with the rotor rotating plane) and the out-of-plane direction (perpendicular to the rotor rotating plane). The relationship between the flap-wise-span-wise and in-plane-out-of-plane coordinate systems is shown in Fig. 5.

where EIB(h) is the flap-wise or span-wise blade distribution rigidity. The potential component produced by rotor rotation can be expressed as: VRotation =

1 2 Ω × 2

    R − RH dφi ( r ) dφ j ( r ) dr +  N + p −1  M Tip R ∫0   dr dr    × N + p −1 ∑     . (18) ′   ′ r dφi ( r ) dφ j ( r ) R − RH × ∑  j= p  + ∫ dr ′  dr   µ B ( r ) ( RH + r )  ∫ i= p   0 0 ′ ′ dr dr         ×ci ( t ) c j ( t ) 

The first item in the bracket is the potential component associating with the mass of the bladetip braking part; the second item is the potential component associating with the blade distribution mass. The positive sum of the centrifugal potential value of the two items is the generalized rigidity to be increased of the blade by the centrifugal force, i.e. centrifugal rigidity. Through simplification, the generalized rigidity of the blade can be expressed as: kij = ∫

R − RH

EI B ( r )

d 2φi ( r ) d 2φ j ( r ) dr 2

+Ω 2  M Tip R + ∫ r 

R − RH

Fig. 5. Relationship between coordinate systems

In the rotor rotation system, the dynamic energy of the blade is consistent with the dynamic energy of the tower. The generalized mass of the blade can be expressed as:

mij = M Tip + ∫

R − RH

µ B ( r )φi ( r )φ j ( r ) dr , (15)

where μB(r) is the blade distribution line density, R is the rotor radius, and RH is the hub radius. When the gravity is neglected, the potential of the blade comprises component VBeam associating with the blade distribution rigidity and component VRotation associating with centrifugal force produced by blade rotation.

V = VBeam + VRotation ,

(16)

The blade potential component associating with the distribution rigidity of the cantilever can be expressed as: VBeam

642

  R − RH d 2φi ( h ) d 2φ j ( h )   EI B ( h ) dh  × 1 N + p −1  N + p −1  ∫0  , (17) dh 2 dh 2 = ∑  ∑   2 i= p  j= p  ×ci ( t ) c j ( t )  

dφi ( r ) dφ j ( r ) dr

dr 2

dr +

µ B ( r ′ ) ( RH + r ′ )dr ′ . 

dr ,

(19)

3 DYNAMIC MODEL AND SOLUTION Suppose the wind turbine is a first order linear multibody system with N freedoms; the motion of the wind turbine system can be described by using N generalized coordinates qi(i = 1, 2, …, N), or by using N generalized velocities ur(r = 1, 2, …, N). The latter are N independent scalars selected arbitrarily from the module values of the rigid body angular velocities or the particle velocities comprising the system, and can be expressed as the linear combination of generalized coordinate differential coefficients qi (i = 1, 2, …, 15). For the cabin, it is simplified as a rigid model for calculation. N

ur = ∑ Yri q i + Z r , r = 1, 2, , N , (20) i =1

where Yri and Zr are the functions of the generalized coordinate qi and time t; if ur is an independent variable, the only solution qi to Eq. (20) can be obtained, and we obtain:

Jin, X. – Liu, H. – Ju, W.B.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

q i = ∑Wir ur + X i , i = 1, , N . (21)

r =1

After the generalized velocity is determined, the absolute angular velocity Eω Ni ( q , q, t ) and absolute linear velocity E v Ni ( q , q, t ) of the Nith rigid body in the wind turbine system corresponding to the inertial coordinate system E can be only expressed as a linear combination of ur.

ω Ni ( q , q, t ) = ∑ EωrNi ( q, t ) ur + EωtNi ( q, t ) , (22) r =1

v Ni ( q , q, t ) = ∑ E vrNi ( q, t ) ur + E vtNi ( q, t ) , (23) r =1

where Eω rNi ( q, t ) and E v rNi ( q, t ) are the rth angular velocity vector and the rth linear velocity vector, respectively, of the Nith rigid body in the inertia coordinate system E. Based on time derivation, we can obtain the angular acceleration Eε Ni ( q, q , q, t ) and the acceleration E a Ni ( q, q , q, t ) of the Nith rigid body in the inertial coordinate system E. After the partial velocity and angular velocity of each rigid body, as well as the corresponding generalized active force F and the generalized inertia force F* are determined, the Kane kinetic equation can be expressed as: ∗ r

Fr + F = 0,

( r = 1, 2,, N ) , (24)

Suppose the wind turbine system comprises w rigid bodies and for each rigid body Ni, the active force is applied on the centroid Xi; the generalized active force of the wind turbine system can be expressed as: w

Fr = ∑ E vrX i ⋅ F X i + EωrNi ⋅ M Ni , i =1

( r = 1, 2,,15) , (25)

The corresponding generalized inertia force can be expressed as: w

(

)

(

)

Fr* = ∑ E vrX i ⋅ −m Ni E a X i + EωrNi ⋅ − E H Ni , i =1

( r = 1, 2, ,15) ,

(26)

where

H Ni

′  H Ni + Eω Ni × E H Ni   =  or . (27)   I Ni ⋅ Eε Ni + Eω Ni × I Ni ⋅ Eω Ni

(

)

Substitute Eq. (25) and Eq. (26) into the Kane kinetic equation (Eq. ((24)), and we can obtain the kinetic equation of the wind turbine system, expressed in matrix as:

C ( q, t )  {q} + { f ( q , q, t )} = {0} , (28)

where [C(q,t)] is the system acceleration coefficient matrix, and { f ( q , q, t )} is the vectors relating to the system displacement and the velocity. To obtain a solution at each time step, the fourth-order Adams– Bashforth prediction-correction algorithm is first used at each time step to determine the value of the lower-order item, which constitutes the right item of the equation; then the Gauss elimination method is used to obtain the system freedom acceleration. The acceleration obtained through these calculations is then used to correct the estimated value and promote precision. Through several iterations, the fourth-order Adams-Bashforth prediction-correction algorithm is used to determine the acceleration and the final solution at the time step. As the prediction-correction algorithm is not spontaneous, solutions at the first four time steps shall be determined by using the fourthorder Runge-Kutta method. 4 EXCITATION EARTHQUAKE LOAD In this paper, the seismic load acts on the base of the wind turbine tower system in the form of an acceleration process, while acceleration is generated based on the acceleration response spectrum designed in the structural specifications. In engineering design, Eurocode 8 is widely applied throughout the world. To be universal, this paper obtains the seismic acceleration process based on Eurocode 8, according to which the designed acceleration response spectrum of a seismic load can be expressed [17] as shown in Fig. 6. The calculation formula is shown as follows: T   Se (T ) = ag ⋅ S ⋅ [1 + T ⋅ (η ⋅ 25 − 1)], 0 ≤ T ≤ TB B   Se (T ) = ag ⋅ S ⋅η ⋅ 25, TB ≤ T ≤ TC  , (29)  T Se (T ) = ag ⋅ S ⋅η ⋅ 25. C , TC ≤ T ≤ TD  T  TCTD   TD ≤ T ≤ 40 s, Se (T ) = ag ⋅ S ⋅η ⋅ 25. T 2 where Se(T) is the elastic response spectrum; T is the vibration period; TB and TC are the periodical constant range limits of the acceleration spectrum; TD is the start period constant value of the displacement response spectrum; ag is the designed ground

Wind Turbine Seismic Load Analysis Based on Numerical Calculation

643

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

acceleration of A-type sites; S is the soil-associated coefficient; η is the damping correction coefficient, according Eurocode 8, η = 10 / (5 + ξ ) ≥ 0.55 , η = 1 and damping rate ξ = 5 in this paper. Fig. 7 is the Acceleration process of an earthquake according Fig. 6.

paper. To validate the correctness of the model, the calculation results are compared with GH Bladed [18]. It can be seen from the comparison that the calculations using this model are correct, as shown in Table 1.

Fig. 6. Earthquake spectrum according Eurocode 8

Fig. 7. Acceleration process of an earthquake

5 EXAMPLE 5.1 Seismic Impact on Wind Turbine Performance To take into account the seismic impact on a working wind turbine, seismic analyses of the wind turbine are carried out through the establishment of a theoretical model. As most winds in the natural world prevail as irregular turbulent winds, if turbulent winds are adopted in the seismic analyses, the significance of a seismic impact can hardly be identified. Table 1. Dynamic motion comparison Deformation Deformation Deformation [m] velocity [m·s-1] acceleration [m·s-2] Blade Model 3.98 10.72 35.27 tip Bladed 4.17 11.92 37.05 Tower Model 0.61 1.26 2.47 top Bladed 0.69 1.45 2.63

Therefore, to better clarify impact of seismic excitation on the wind turbine, a steady wind at the speed of 11m/s is used in the case described in this 644

Fig. 8. Seismic impact on the performances of wind turbines; a) generator torque comparison, b) wind turbine speed comparison, and c) wind turbine power comparison

Fig. 8 shows the torque, speed and power of the wind turbine. In the figures, the part in red indicates normal generation without considering seismic impact, and the part in blue indicates normal generation under seismic impact. It can be seen from the three figures that upon application of seismic excitation at the 20th second, the torque, speed and power of the wind turbine are disturbed to different degrees. In Fig. 8a, the disturbance to the wind turbine torque is relatively great, about 6.01% as calculated, mainly due to minor changes of the incoming wind velocity relative to the blade element upon occurrence of a seismic shock, resulting in a change of the aerodynamic torque; in Fig. 8b, disturbance to the wind turbine speed is relatively small, about 1% as calculated, mainly due to

Jin, X. – Liu, H. – Ju, W.B.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

short seismic excitation and inertia of the rotor; in Fig. 8c, disturbance to the wind turbine power is mainly caused by joint action of the wind turbine speed and torque. In general, the seismic shock has no significant impact on wind turbine operation, provided the wind turbine structure is not damaged; whether the structure will be damaged mainly depends on the load caused by the earthquake. 5.2 Seismic Impact on Load-Bearing Conditions Generally, the load on the tower base, spindle and blade root of a wind turbine is the most representative in the load calculation. Therefore, to study the seismic impact on the major load on a wind turbine, three

corresponding load groups are described, i.e. the load on the tower base, load on the spindle and load on the blade root (in the seismic propagation order). Similarly, in the figures, the part in red indicates normal generation without considering seismic impact, and the part in blue indicates normal generation under seismic impact. Fig. 9 shows the load on the tower base; Fig. 9a is the tower base bending moment Mx, with the maximum load fluctuation caused by the seismic shock reaching 188%; Fig. 9b is the tower base bending moment My, with the maximum load fluctuation caused by the seismic shock reaching 108%; Fig. 9c is the tower base bending moment Mz, with the maximum load fluctuation caused by the seismic shock reaching 45%; Fig. 9d is load Fx on

Fig. 9. Seismic impact on load-bearing conditions of the tower base; comparison of: a) foundation bending moment Mx, b) foundation bending moment My, c) foundation bending moment Mz, d) load Fx acting on the base, e) load Fy acting on the base, f) load Fz acting on the base Wind Turbine Seismic Load Analysis Based on Numerical Calculation

645

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

Fig. 10. Seismic impact on load to be borne by the hub; comparison of: a) hub bending moment Mx , b) hub bending moment My , c) hub bending moment Mz , d) load Fx acting on the hub, e) load Fy acting on the hub, f) load Fz acting on the hub

the tower base, with the maximum load fluctuation caused by the seismic shock reaching 233%; Fig. 9e is load Fy on the tower base, with the maximum load fluctuation caused by the seismic shock reaching 500%; Fig. 9f is the tower base bending moment Mz, with the maximum load fluctuation caused by the seismic shock reaching 2.41%. Generally, as the seismic shock first arrives at the tower base, it causes great fluctuation of load on the tower base. Fig. 10 shows the load on the spindle hub; Fig. 10a is the bending moment Mx, with the maximum load fluctuation caused by the seismic shock reaching 6.1%; Fig. 10b is My, with the maximum load fluctuation caused by the seismic shock reaching 66.7%; Fig. 10c is the tower base bending moment 646

Mz, with the maximum load fluctuation caused by the seismic shock reaching 30%; Fig. 10d is load Fx on the tower base, with the maximum load fluctuation caused by earthquakes reaching 36%; Fig. 10e is load Fy on the tower base, with the maximum load fluctuation caused by the seismic shock reaching 150%; Fig. 10f is the tower base bending moment Mz, with the maximum load fluctuation caused by the seismic shock reaching 20%. Generally, as the seismic shock first arrives at the tower base, it causes great fluctuation of load on the tower base. The results show that as the tower is a flexible fine long piece, loads caused by seismic excitation will be reduced in the amount and magnitude when arriving at the spindle after passing the tower.

Jin, X. – Liu, H. – Ju, W.B.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

Fig. 11. Seismic impact on blade root load; a) blade flap-wise moment comparison, b) blade edge-wise moment comparison, c) blade spanwise moment comparison, e) blade flap-wise force comparison, d) blade edge-wise force comparison, f) blade span-wise force comparison

Fig. 11 shows the load on the blade root, specifically the flap-wise, edge-wise and span-wise moments and loads, respectively. It can be seen from the figures that, due to seismic impact, load fluctuations only occur in a) and e), but the load in other figures is not subject to significant impacts, mainly because the blade is a flexible fine long piece, just as the tower, which reduces the seismic impact. 6 CONCLUSION According to the multi-body dynamic theory this paper proposes a multi-body-system dynamic model based on blade-cabin-tower-foundation coupled multi-body system dynamic model with soil-structure interaction considered to study the load-bearing conditions of

wind turbines subject to seismic impact. Based on the basic theory of multi-body system dynamics, the wind turbine blade and tower system comprise a series discrete continuous units, while the soil-structure interaction of the tower system is realized via the spring and damping set on the interface between the foundation and the soil body; the cabin is simplified as a rigid model. Analyses show that a sudden earthquake occurring during normal operation of a wind turbine will disturb the wind turbine performance to some extent, but the generating performance will not fluctuate greatly, provided that the wind turbine structure is not damaged; if the wind turbine structure is damaged, the disturbance intensity will depend on the load caused by the seismic shock. Through the

Wind Turbine Seismic Load Analysis Based on Numerical Calculation

647

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 638-648

three typical load groups selected for the wind turbine, it can be seen that load on the tower base suffers the greatest fluctuation when it is the nearest to the foundation, which presents higher requirements for the tower base design; the seismic excitation transmits through the tower and then reaches the top of the tower, so the load fluctuation at the spindle reduces greatly; when it reaches the blade, the load fluctuation reduces more significantly as the blade is of a flexible structure. Generally, the seismic shock exerts the most significant impact on the tower base, but much less significant impact on the upper part, which is an advantage of the flexible structure. Therefore, when designing wind turbines for earthquake-prone regions, flexible structures should be considered for vibration relief. 7 ACKNOWLEDGEMENT This paper is sponsored by Natural Science Foundation of China (approval No.: 5100 5255) and the Fundamental Research Funds for the Central Universities. (Approval No.: CDJZR10 11 00 04). 8 REFERENCES [1] Stamatopoulos, G.N. (2013). Response of a wind turbine subjected to near-fault excitation and comparison with the Greek Aseismic Code provisions. Soil Dynamics and Earthquake Engineering, vol. 46, p. 77-84, DOI:10.1016/j.soildyn.2012.12.014. [2] Teng, W., Wang, F., Zhang, K.L., Liu, Y.B., Ding, X. (2014). Pitting fault detection of a wind turbine gearbox using empirical mode decomposition. Strojniški vestnik - Journal of Mechanical Engineering, vol. 60, no. 1, p. 12-20, DOI:10.5545/sv-jme.2013.1295. [3] Potočar, E., Širok, B., Hočevar, M., Eberlinc, M. (2014). Control of Separation Flow over a Wind Turbine Blade with Plasma Actuators. Strojniški vestnik - Journal of Mechanical Engineering, vol. 58, no.1, p. 37-45, DOI:10.5545/sv-jme.2011.016. [4] DNV/Risø National Laboratory. (2001). Guidelines for design of wind turbines. 2nd ed., Wind Energy Department, Roskilde. [5] Bazeos, N., Hatzigeorgiou, G.D., Hondros, I.D., Karamaneas, H., Karabalis, D.L., Beskos, D.E. (2002). Static, seismic and stability analyses of a prototype wind turbine steel tower. Engineering Structures, vol. 24, no. 8, p. 1015-1025, DOI:10.1016/S01410296(02)00021-4.

648

[6] Lavassas, I., Nikolaidis, G., Zervas P., Efthimiou, E., Doudoumis, I.N., Baniotopoulos, C.C. (2003). Analysis and design of the prototype of a steel 1-MW wind turbine tower. Engineering Structures, vol. 25, no. 8, p. 1097-1106, DOI:10.1016/S0141-0296(03)00059-2. [7] Murtagh, P.J., Basu, B., Broderick, B.M. (2005). Along-wind response of a wind turbine tower with blade coupling subjected to rotationally sampled wind loading. Engineering Structures, vol. 27, no. 8, p. 12091219, DOI:10.1016/j.engstruct.2005.03.004. [8] Murtagh. P.J., Basu, B., Broderick, B.M. (2005). Along-wind response of a wind turbine tower with blade coupling subjected to rotationally sampled wind loading. Engineering Structures, vol. 27, no. 8, p. 12091219, DOI:10.1016/j.engstruct.2005.03.004. [9] Witcher, D. (2005). Seismic analysis of wind turbines in the time domain. Wind Energy, vol. 8, no. 1, p.81-91, DOI:10.1002/we.135. [10] Kang, H., Li, Y., Wu, F., Guo, W., Huan, K. (2008). A system reliability analysis method for offshore wind turbine foundation. Electronic Journal of Geotechnical Engineering, vol. 13, bundle L, p. 1-10. [11] Harte, M., Basu, B., Nielsen, S.R.K. (2012). Dynamic analysis of wind turbines including soil-structure interaction. Engineering Structures, vol. 45, p. 509518, DOI:10.1016/j.engstruct.2012.06.041. [12] Lombardi, D., Bhattacharya, S., Wood, D.M. (2013). Dynamic soil–structure interaction of monopile supported wind turbines in cohesive soil. Soil Dynamics and Earthquake Engineering, vol. 49 p. 165-180, DOI:10.1016/j.soildyn.2013.01.015. [13] Newmark, N.M., Rosenblueth, E. (1971). Fundamentals of Earthquake Engineering. Prentice Hall, New Jersey. [14] Stejska, J.V., Valasek, M. (1996). Kinematics and dynamics of machinery. Marcel Dekker, Inc., New York. [15] Wolf, J.P. (1997). Spring-dashpot-mass models for foundation vibrations. Journal of Earthquake Engineering & Structural Dynamics, vol. 26, no. 9, p. 931-949, DOI:10.1002/(SICI)10969845(199709)26:9<931::AID-EQE686>3.0.CO;2-M [16] Mulliken, J.S., Karabalis, D.L. (1998). Discrete model for dynamic through-the-soil coupling of 3-D foundations and structures. Earthquake Engineering & Structural Dynamics, vol. 27, no. 7, p. 687-710, DOI:10.1002/(SICI)10969845(199807)27:7<687::AID-EQE752>3.0.CO;2-O. [17] European Committee for Standardization (2003). Eurocode8: Design of Structure for Earthquake Resistance. London South Bank University, London, [18] Bladed, Wind Turbine Design Software (2014). from http://www.gl-garradhassan.com/en/software/ GHBladed.php, accessed on 2014-01-04.

Jin, X. – Liu, H. – Ju, W.B.

Received for review: 2014-03-04 Received revised form: 2014-04-30 Accepted for publication: 2014-05-16

A Computational Model for Bending Fatigue Analyses of Sintered Gears Glodež, S. – Šori, M. – Verlak, T. Srečko Glodež1,* – Marko Šori2 – Tomaž Verlak1

1 University 2 University

of Maribor, Faculty of Mechanical Engineering, Slovenia of Maribor, Faculty of Natural Science and Mathematics, Slovenia

A computational model for determination of the fatigue life of sintered gears in regard to bending fatigue in a gear tooth root is presented. The proposed model is based on the stress-life approach in which the multi-axial state of stress, the mean stress effect, the influence of surface roughness, and the notch effect are studied when determining the fatigue life of a treated gear pair. The required material parameters (the fatigue strength coefficient σf' and the fatigue strength exponent b) are determined experimentally on a uni-axial tension/compression test machine with a load ratio of R = 0. Here, the influence of additional thermal treatment (after sintering) on the fatigue strength is studied. The model is used for the determination of the fatigue life of real spur gear made from a Höganäs Distaloy AB powder mixture, while the stress field in a gear tooth root is determined numerically using the FEM method. Keywords: sintered gears, fatigue, stress life approach, numerical modelling, experiments

0 INTRODUCTION Powder metallurgy (P/M) is useful in making parts that have irregular curves, or that are difficult to machine. It is suitable for high volume production with very little wastage of material. Powder metal gears were initially used only for light-duty applications, such as toys and power tools. Today, powder metal gears are a cost-efficient alternative for machined gears in larger series in the automotive industry (synchronizer gears, oil pump gears, engine gears, etc.). The next step should be power transmission gears [1]. A critical review regarding the application of P/M-sintered gears for transmissions and machinery was presented by Dizdar [2], who asserted that P/Msintered gears can reach relatively high levels of dynamic strength when compared to wrought steel gears. They also offer highly sustainable production, low cost and full recycling for a range of applications in the automotive, power tool and home appliance industries. In recent years, high-performance sintered steel gears have been extensively investigated by researchers, specifically with regards to some new technologies (surface densification, gear rolling, burnishing, shot peening, high density pressing, warm compaction, warm die pressing techniques, etc.) and their influence on gear characteristics [3] to [5]. The research work presented in this paper is focused on the fatigue failures of sintered gears. Although two kinds of fatigue failures (surface pitting and tooth breakage) should be taken into account when dimensioning gear drives, only the tooth breakage [6] and [7] is addressed in this paper. ISO 6336, the classic standardised procedure [8], is usually used to determine the bending load capacity of treated

gear pairs; however, it does not apply to gears finished by sintering. Some guidelines on how to calculate the bending load capacity of sintered external spur gears are described in the AGMA 930-A05 information sheet [9], although the required fatigue strength properties are not included in this sheet and should be taken from the available literature or determined experimentally. There is also a complicated geometrical analysis to determine the geometry factor for bending strength, which is usually quite impractical when calculating sintered gear pairs. In this paper, the stress-life approach is used to determine the fatigue life of sintered gears in regard to the bending stress in a gear tooth root. Because the stress field in the critical cross section in a gear tooth root is determined numerically using the FEM method, the proposed computational model can be applied to a wide range of gear pairs and is not limited by some geometrical parameters, as with the ISO [8] and AGMA [9] standards. 1 BENDING FATIGUE ANALYSIS USING STRESS LIFE APPROACH The stress-life (S-N) approach is usually used to determine the fatigue life of dynamically loaded machine parts, when stresses and strains are mostly elastic. This approach is based on the following Basquin equation: 1

 σ b N =  b a  , (1)  2 ⋅σ f '   

where N is the number of stress cycles to failure, σa is the alternating stress, σf' is the fatigue strength

*Corr. Author’s Address: University of Maribor, Faculty of Mechanical Engineering, Smetanova 17, 2000 Maribor, Slovenia, srecko.glodez@um.si

649

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 649-655

coefficient and b is the fatigue strength exponent. The material parameters σf' and b can be determined experimentally, usually by means of a rotating bending test under fully reversed uniaxial stressing (stress ratio R = −1). If this is not the case, the influence of mean stress and a multi-axial state of stress should be considered when determining the appropriate alternating stress σa in Eq. (1). In this paper, the material parameters σf’ and b have been determined with a uni-axial pulsating test machine under a stress ratio of R = 0 (see Section 3). Because the stress ratio of R = 0 corresponds to the operation of real gear pairs, the mean stress effect can be omitted in this case. Otherwise, the effect of multi-axial stress in a gear tooth root should be considered using the Mises hypothesis [10]. Once the equivalent alternating stress σa,eq is determined on the basis of alternating principal stresses in a gear tooth root σa1, σa2 and σa3 (see Section 3), it should be considered when determining the fatigue life according to Eq. (1). When determining the fatigue life of machine parts by using Eq. (1), the influence of the notch effect and the surface finish should also be considered, because the material parameters σf’ and b are usually determined using non-notched specimens with polished surfaces. If this is not the case, the appropriate notch factor Kf and surface factor Ks should be taken into account [11]. In the work presented in this paper, the test specimens used have been produced in the same way as actual sintered gears (compaction, sintering, additional thermal treatment, without additional polishing); see Section 2. Because the surface finish of the test specimens corresponds to the surface finish of real gears, the surface factor Ks = 1 can be assumed in such cases. The notch factor Kf is indirectly included in the alternating principal stresses σa1, σa2 and σa3, which are determined numerically in this study using the FEM method (see Section 3). 2 EXPERIMENTAL TESTING The Automatic Die Compaction and Sintering (ADC/ SINT) procedure has been used to prepare the test specimens. A base metal powder (iron) is mixed with alloying elements and lubricant. The alloying elements are added not only to improve material properties [12] and [13], but also to control dimensional change while sintering [14] and the final density of sintered component [15]. The powder mixture used in this study was Höganäs Distaloy AB with an addition of 0.58 wt% of Kenolube P11 and 0.3 wt% of carbon in the form of graphite UF4 (see Table 1). The powder mixture 650

had to be compressed into a desired shape. Before the compaction of specimens, the apparent density of the powder was 3.15 g/cm³ and the hall flow rate was 29 s per 50 g. Flat specimens (Fig. 1) were cold compacted at a compacting pressure of 485 MPa and then sintered for 30 minutes in a 10/90 hydrogen and nitrogen atmosphere at 1120 °C. After sintering, half of the specimens were subjected to the additional hardening (austenitization at 915 °C, oil-quenched and tempered for 1 h at 175 °C). Both sets of specimens had a final density of 7.07 g/cm³.

Fig. 1. Test specimen Table 1. Chemical composition of used powder mixture and comparison with standardized powders according to DIN 30910-4 [16] [wt%]

Specimens

SINT-D30 DIN 30910-4

Bal

C Cu Ni Mo

0.29 1.47 1.69 0.50

< 0.3 1.0 to 5.0 1.0 to 5.0 < 0.6

Kenolube

0.58

Additional grinding of specimens was done before experimental testing to remove the sharp edges that were a result of the compaction process and that could significantly affect the experimental results. However, the surface of the specimens was not additionally polished; therefore, surface roughness was measured at the locations and directions indicated in Fig. 2. Differences between both sets of specimens (with and without hardening) were negligible. Therefore, the average surface roughness of multiple measurements

Glodež, S. – Šori, M. – Verlak, T.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 649-655

from four specimens for given directions in Fig. 2 was Ra[1 - 5] = [0.83, 1.16, 0.80, 0.79 and 0.65] μm and the average surface roughness at the thinned section (Directions 1, 4 and 5) was Ra = 0.76 μm. Six measurements (three at Position 1 and three at Position 2 in Fig. 2) of surface hardness were taken on three randomly chosen specimens from each set. The results of these measurements were in the range 160 to 180 HV1 for sintered specimens and 310 to 340 HV1 for sintered and additionally hardened specimens.

Fig. 2. Directions of measured roughness

2.1 Static Tests Static properties of randomly chosen specimens from both sets were determined in a controlled environment at room temperature (20 °C) on a uni-axial pulling test machine with a data acquisition rate of 500 Hz. The displacement rate for all quasi-static tests was set to 0.5 mm/min. Fig. 3 shows the stress-strain diagram for both sets of specimens. The determined static material properties are shown in Table 2.

Fig. 3. Stress – strain diagram for both sets of specimens

Table 2. Material properties of test specimens by quasi-static loading Thermal treatment sintering sintering + hardening

Young’s modulus Ultimate strength E [GPa] Rm [MPa] 130 532 142

842

Elongation A [%] 2.16 0.86

2.2 Fatigue Tests As explained in Section 1, the rotating bending test [11] is usually used to determine fatigue properties (material parameters σf’ and b), which are needed to determine the fatigue life of dynamically loaded machine parts, using Eq. (1). Due to the rectangular cross-section of the treated test specimens, fatigue testing on a rotating beam machine was not possible. Therefore, it was performed on a uni-axial tension/ compression test machine with a load ratio of R = 0. In order to achieve this load ratio, the loadcontrol regime was induced in such a way that maximum load was set. To ensure that the load at the start of dynamic tests did not exceed this limit, at the first cycle, maximum load was not equal, but approximately 90% of what it was set to be. A similar situation was applied for the minimum load, which was not zero at first cycle, but around 10% of the specified maximum load. The maximum and minimum loads were then gradually altered to meet set specifications after 50 cycles. To avoid excessive heating of the specimens due to damping effects [17], testing was done at the loading frequency f = 10 Hz, because cooling of the specimens was not possible. Figs. 4 and 5 show the experimental results of fatigue tests for both sets of specimens. Data points (black dots) represent the situation when fracturing occurs after an appropriate number of stress cycles N when the test specimen is loaded with given alternating stress σa. The method of least squares was then used to determine the S–N curves and, consequently, the material parameters σf’ and b (see Table 3). Fig. 6 shows the comparison of S-N curves for both sets of specimens. When comparing the fatigue strength at 104 cycles, the calculated values from S-N curves are 192 MPa for hardened specimens and 162 MPa for unhardened specimens. It is also evident that the difference in fatigue strength gradually decreases when the number of stress cycles increases. Thus, the fatigue strength at 106 cycles would be almost the same for both sets of specimens. However, this is a rough assessment because there are no data points after 106 cycles (see Figs. 4 and 5). Therefore,

A Computational Model for Bending Fatigue Analyses of Sintered Gears

651

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 649-655

3 NUMERICAL MODEL

additional testing should be performed to determine the fatigue limit of discussed material. Table 3. Material properties by dynamic loading Thermal treatment sintering sintering + hardening

fatigue strength coefficient σf’ [MPa] 537

fatigue strength exponent b [-]

875

–0.153

–0.121

Fig. 4. S–N curve for unhardened specimens

Fig. 5. S–N curve for hardened specimens

Fig. 6. Comparison of S – N curves for both sets of specimens

652

The presented model has been used for the computational determination of the fatigue life of spur gears with module m = 4 mm, pressure angle at normal section αn = 25°, number of teeth z = 9 and gear width b = 10 mm. In the computational analysis, it was assumed that gears are made of the same material as the tested specimens (see Section 2). To determine the fatigue life of the treated gear using Eq. (1), the alternating principal stresses σa1, σa2 and σa3 in a gear tooth root should be known. Here, the principal stresses have been determined numerically using FEM-method in the framework of commercial software Abaqus [18]. As seen from Fig. 7, only one third of the discussed gear is modelled (Instance 1) and one tooth of a pairing gear (Instance 2) is added to the assembly in such a way that the contact between the two of them is at the outermost single contact point. The surfaces of shaft holes with a diameter of di = 15 mm and cut result surfaces were coupled to the reference points in such a way that all relative displacements and rotations between them were restricted. All degrees of freedom of both reference points were set to zero at the initial step. In the load step, rotation around the gear axis was applied to a reference point coupled to Instance 1 (RP1). Because of contact between the teeth, a reaction moment is calculated for the reference point, which is considered to be the torque load to the gear that causes the stress field in the tooth root. This approach was taken because of the significantly greater simulation stability in comparison to the torque boundary condition. The material of both sections was modelled as linearly elastic with Young’s modulus for hardened sintered steel, which was taken from data obtained by experimental testing of this material discussed in Section 2. Contact between the two sections was modelled as tangentially frictionless and as “hard” contact in the normal direction. The whole assembly was meshed with linear tetrahedral elements of type C3D4H. There were approximately 427,000 elements in Instance 1 and around 64,000 elements in Instance 2 (Fig. 8). Numeric simulation was done in 20 increments. The first increments were used to establish contact between two instances and the rest are used to obtain principal stress components in the tooth root of Instance 1 for different reaction moments. Although stress gradients are covered to consider stress intensity factor dependent on the notch geometry, the material notch sensitivity factor is neglected in this approach.

Glodež, S. – Šori, M. – Verlak, T.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 649-655

Table 4. Principal stresses in a gear tooth root for different torques Torque

Max. Principal [MPa]

Mid. Principal [MPa]

Min. Principal [MPa]

σa,eq

T [Nm]

σ1

σa1

σ2

σa2

σ3

σa3

[MPa]

40 45 50 55 60 75 80

220 247 275 302 330 412 439

110 124 137 151 165 206 220

50 56 62 68 74 93 99

25 28 31 34 37 47 50

8 9 10 11 11 14 15

4 4 5 5 6 7 8

97 109 122 134 146 182 195

Fig. 7. 3D model of treated gear pair

Fig. 9 shows a numerically determined Mises stress field in the gear half width plane for the 12th increment, where the torque is equal to T = 45 Nm. The light grey colour indicates areas with Mises stress higher than 220 MPa; those areas are subjected to compression stress and are not problematic for fatigue crack nucleation and propagation in a gear tooth root. Principal stress components and Mises stress according to different load moments are shown in Table 4. All the values in the table are taken from the same node that has maximum principal stress in the tooth root.

Fig. 9. Mises stress in a gear tooth root for the torque T = 45 Nm

Fig. 10. Fatigue life diagram of analysed gear

Fig. 8. FEM-mesh of analysed gear

Stress levels from FEM simulation and experimental data of dynamic tests are combined in Fig. 10, where dependence between load torque and the number of expected cycles before failure is presented. In the region between 103 and 105 cycles, sintered and additionally hardened gears should perform better than gears that are only sintered. However, Fig. 10 suggests that after 106 cycles, sintered-only gears would outperform additionally hardened ones, which could be explained by the greater ductility of sintered-only specimens, which causes longer crack propagation periods at lower stress levels, because the wider plastic zone at the crack tip decreases crack

A Computational Model for Bending Fatigue Analyses of Sintered Gears

653

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 649-655

growth rate. However, as mentioned in Section 2, the fatigue limit of the discussed materials should be found in order to prove or disprove this suggestion. 4 CONCLUSIONS A computational model for the analysis of the bending fatigue of sintered gears is presented. The proposed model is based on the stress-life approach, in which the required material parameters are determined experimentally on the uni-axial tension/ compression test machine with a load ratio of R = 0. The stress field in a gear tooth root is determined numerically using FEM. The proposed model is used to determine the fatigue life of a sintered spur gear made of Höganäs Distaloy AB powder mixture; the influence of additional thermal treatment after sintering on the fatigue strength is also studied. On the basis of experimental testing and comprehensive computational analyses, the following conclusions can be made: • Additional hardening significantly increases (by more than 35%) the static strength of treated sintered material (see Table 2). • Additional hardening improves fatigue strength by approximately 15% at 104 stress cycles. However, the difference in fatigue strength gradually decreases when the number of stress cycles increases (the fatigue strength at 106 cycles is almost the same for specimens with and without additional hardening). • As with sintered specimens, additional hardening also increases the determined operational time of sintered gears at given torque. • Existing procedures for the determination of load capacity of sintered gears require the calculation of many influential factors that are mainly dependent on the geometry of a gear. FEM analysis may be used to replace this procedure and numerically determine the stress field. Usually, dies for sinter-press technology are made by wire-EDM, and a 2D contour should be provided, which can be used for the rapid preparation of an FEM model. • Although some fatigue data for sintered steels can be found in the literature, experimental testing of the exact material to be used is preferable, because mechanical properties can be significantly affected by many variables: density, sintering temperature and time, type and quantity of lubricant, ratio of alloying elements and additional heat treatment. 654

The computational model used in this study considers only the final stage of the whole fatigue process, i.e. the occurrence of final breakage after an appropriate number of stress cycles. However, the fatigue process leading to breakage may be divided into the crack initiation (Ni) and crack propagation (Np) period, which enables the determination of the total service life as N = Ni+Np [19] and [20]. In further research, the strain-life approach may be used to determine the number of stress cycles Ni required for the fatigue crack initiation at the point of the largest stresses in a gear tooth root. When the initial crack of length ai is known, the appropriate crack growth model should be used to determine the number of stress cycles Np for crack propagation from the initial to the critical length when final fracture can be expected to occur. Here, some fracture mechanics parameters should be determined experimentally, because they are not known for treated sintered material. However, when gears operate for extended periods, failures other than tooth breakage may occur. Failures such as wear [21] and surface pitting usually prevail over fatigue tooth fracture and through hardening or surface hardening are well-established methods to improve surface hardness. The subject of further research work may also be the use of selective laser sintering (SLS) to produce small sintered gears. As described in [22], SLS is a new manufacturing technique, which uses a high power CO2 laser to melt or sinter metal powder particles into a mass that has a desired three-dimensional shape in precisely defined areas. 5 REFERENCES [1] Flodin, A., Brecher, C., Gorgels, C., Rothlingshofer, T., Henser, J. (2011). Designing powder metal gears. Gear Solutions, August, p. 26-35. [2] Dizdar, S. (2012). High-performance sintered-steel gears for transmissions and machinery: A critical review. Gear Technology, August, p. 60-65. [3] Sudhakar, K.V. (2000). Fatigue behaviour of a high density powder metallurgy steel. International Journal of Fatigue, vol. 22, no. 9, p. 729-734, DOI:10.1016/ S0142-1123(00)00067-0. [4] Sonsino, C.M., Mueller, F., Mueller, R. (1992). The improvement of fatigue behaviour of sintered steels by surface rolling. International Journal of Fatigue, vol. 14, no. 1, p. 3-13, DOI:10.1016/0142-1123(92)901475. [5] Koide, T., Ishizuka, I., Takemasu, T., Miyachika, K., Oda, S. (2008). Load bearing capacity of surfacerolled sintered metal gears. International Journal of Automation Technology, vol. 2, no. 5, p. 334-340.

Glodež, S. – Šori, M. – Verlak, T.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 649-655

[6] Glodež, S., Flašker, J., Kramberger, J. (2002). A computational model for calculating the bendingload capacity of gears. Strojniški vestnik – Journal of Mechanical Engineering, vol. 48, no. 5, p. 257-266. [7] Glodež, S., Flašker, J., Pehan, S. (1994). Prediction of service life of spur gears using statistical methods. Strojniški vestnik – Journal of Mechanical Engineering, vol. 40, no. 11-12, p. 393-400. [8] ISO 6336 (2006). Calculation of Load Capacity of Spur and Helical Gears, International Organization for Standardization, Geneva. [9] AGMA 930-A05 (2005). Calculated Bending Load Capacity of Powder Metallurgy (P/M). External Spur Gears, Information sheet, 2005. [10] Stephens, R.I., Fatemi, A., Stephens, R.R., Fuchs, H.O. (2001). Metal Fatigue in Engineering, John Wiley & Sons Inc., New York. [11] Dowling, N.E. (2007). Mechanical Behaviour of Materials. Pearson Prentice Hall, New York. [12] Candela, N., Velasco, F., Martinez, M.A., Torralba, J.M. (2005). Influence of microstructure on mechanical properties of molybdenum alloyed P/M steels. Journal of Materials Processing Technology, vol. 168, no. 3, p. 505-510, DOI:10.1016/j.jmatprotec.2004.02.066. [13] Polasik, S.J., Williams, J.J., Chawla, N. (2002). Fatigue crack initiation and propagation of binder-treated powder metallurgy steels. Metallurgical and Materials Transactions, a-Physical Metallurgy and Materials Science, vol. 33a, p. 73-81, DOI:10.1007/s11661-0020006-8. [14] Petrova, A.M., Stepichev, A.V. (1998). Effect of carbon on volume changes during the sintering of an ironchromiuim material. Powder Metallurgy and Metal Ceramics, vol. 37, no. 5-6, p. 270-273, DOI:10.1007/ BF02675860.

[15] Khraisat, W., Nyborg, L. (2004). Effect of carbon and phosphorus addition on sintered density and effect of carbon removal on mechanical properties of high density sintered steel. Materials Science and Technology, vol. 20, no. 6, p. 705-710, DOI:10.1179/026708304225017210. [16] DIN 30910-4 (2010). Sintered metal materials: Sintered material specifications – Part 4: Materials for structural parts, German standard, Berlin. [17] Dlapka, M., Danninger, H., Gierl, C., Klammer, E., Weiss, B., Khatibi, G., Betzwar-Kotas, A. (2012). Fatigue behaviour and wear resistance of sinterhardening steels. International Journal of Powder Metallurgy, vol. 48, no. 5, p. 49-60. [18] ABAQUS Version 6.12-1 (2012). Dassault Systèmes Simulia Corp., Providence. [19] Fajdiga, G., Flašker, J., Glodež, S., Ren, Z. (2000). Numerical simulation of the surface fatigue crack growth on gear teeth flanks. Strojniški vestnik – Journal of Mechanical Engineering, vol. 46, no. 6, p. 359-369. [20] Podrug, S., Glodež, S., Jelaska, D. (2011). Numerical modelling of crack growth in a gear tooth root. Strojniški vestnik – Journal of Mechanical Engineering, vol. 57, no. 7/8, p. 579-586, DOI:10.5545/sv-jme.2009.127. [21] Miltenović, A., Nikolić, V., Milovančević, M., Banić, M. (2012). Experimental and FEM analysis of sintered steel worm gear wear. Transaction of Famena, vol. 36, no. 4, p. 85-96. [22] Dobrzanski, L.A., Musztyfaga, M., Drygala, A. (2013). Final manufacturing process of front side metallisation on silicon solar cells using conventional and unconventional techniques. Strojniški vestnik – Journal of Mechanical Engineering, vol. 59, no. 3, p. 175-182, DOI:10.5545/sv-jme.2012.625.

A Computational Model for Bending Fatigue Analyses of Sintered Gears

655

Received for review: 2014-01-10 Received revised form: 2014-05-20 Accepted for publication: 2014-07-08

Optimization of Intervening Variables in MicroEDM of SS 316L using a Genetic Algorithm and Response-Surface Methodology Suresh, P. – Venkatesan, R. – Sekar, T. – Elango, N. – Sathiyamoorthy, V. Periyakgounder Suresh1,* – Rajamanickam Venkatesan1 – Tamilperruvalathan Sekar2 – – Natarajan Elango3 – Varatharajan Sathiyamoorthy4 1 Sona

College of Technology, Department of Mechanical Engineering, India College of Engineering, Department of Mechanical Engineering, India 3 Kolej University Linton, School of Mechanical Engineering, Malaysia 4 Mahendra Engineering College, Department of Mechanical Engineering, India

2 Government

This research paper attempts to investigate the optimum values of the major intervening parameters in micro-Electric Discharge Machining (microEDM) of Stainless Steel (SS) 316L. Experiments are conducted using a 400 micrometre brass electrode. The discharge current, pulseon time and pulse-off time with three levels are selected as significant intervening parameters. The Taguchi method is initially applied to determine the optimum process parameters and the number of experiments required to model the responses. The response-surface methodology (RSM) is applied to correlate between intervening parameters, and the selected objectives to maximize the material removal rate (MRR) and to minimize the tool wear rate (TWR) in the machining of SS 316 L. The mathematical model obtained from RSM is used as a fitness function to multi-objective optimization using a genetic algorithm (GA). The results reveal that the resulting optimal intervening parameters improve the chosen objectives significantly. The confirmation results prove that the better developed mathematical model yields deviate within 5% of the experiment. Keywords: response-surface methodology, genetic algorithm, stainless steel 316L, Taguchi method

0 INTRODUCTION Electrical discharge machining (EDM) is the most widely used and most successfully applied method to machine conductive hard materials. It is a nontraditional machining process in which metal is removed by producing powerful electric spark discharge between the tool electrode and the work material. Both the work piece and the tool are submerged in a dielectric fluid and a servo-mechanism is employed to maintain the spark gap.

Fig. 1. Schematic layout of microEDM

A high-power spark is produced when the voltage across the gap becomes sufficiently large. Hence, the dielectric fluid breaks down and the gap is ionized. Thousands of sparks occur per second at the spark gap 656

and make the work-piece metal melt and erode. The removed metal is carried away by the dielectric fluid circulated around it, as shown in Fig. 1 [1] and [2]. In EDM, the problem of cutting force and vibration is avoided since the tool does not contact the work piece directly. In spite of many advantages, it has some limitations, such as longer lead time, lower productivity and higher energy consumption. Therefore, recent research focuses on optimizing the process parameter to increase the productivity and the capability of the process. The experimental methods increase the cost of investigation, and performing all the experiments is not feasible, particularly when the number of parameters and their levels are high. The Taguchi method has evolved to become the most powerful way to improve the productivity of EDM [3] and [4]. It was used for experimental design to optimize the cutting parameters of the turning of E0300 alloy steel [5]. Natarajan and Arunachalam [6] applied this method and the grey relational analysis to optimize the process parameters of stainless steel grade 304 with brass electrodes 500 µm in diameter. Dhanabalan et al. [7] optimized the process parameters of titanium grades in EDM. Mukherjee and Ray [8] presented a generic framework for parameter optimization in metal cutting processes for the selection of an appropriate approach. The response-surface methodology (RSM) explores the relationships between several explanatory variables and one or more response variables. It will successfully relate the input process parameters and

*Corr. Author’s Address: Sona College of Technology, Junction Main Road, Salem-5, TN, India, suresh_p_g@yahoo.com

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

output response variables [9] and [10]. It is a statistical method that uses quantitative data from experiments to determine and simultaneously solve multi-variant equations. Karthikeyan et al. [11] conducted general factorial experiments for microEDM in order to present an exhaustive study of parameters on the material removal rate (MRR) and the tool wear rate (TWR). Kung et al. [12] introduced powder-mixed EDM when machining cobalt-bonded tungsten carbide. The RSM was used to plan and analyse the experiments in terms of MRR and electrode wear ratio (EWR). They concluded that the aluminium powder mixed with dielectric fluid increases the MRR and reduces the EWR. Genetic algorithms (GA) and artificial neural networks (ANN) are popular software technologies used for the optimization of machining parameters. Samtas et al. [13] investigated the effects of cutting parameters and deep cryogenic treatment on the thrust force in the drilling of AISI 316 stainless steel. Saric et al. [14] used neural networks to predict and simulate the surface roughness of the steel, by using back-propagation neural networks, modular neural networks, and radial basis function neural networks in the process of modelling. Kao and Hocheng [15] applied grey relational analysis for optimizing the electro-polishing of 316L stainless steel with multiple performance characteristics. Lee et al. [16] studied the process of ball burnishing AISI 316L stainless steel, in which they used Taguchi techniques for the statistical design of experiments for achieving good surface finish on flat specimens. Pushpendra et al. [17] developed an artificial neural network model for the experimental values and then applied a non-dominated sorting genetic algorithm (NSGA II) to predict the MRR and surface roughness (SR) for Inconol 718. They concluded experimental results with a set of pareto-optimal solutions. Baraskar et al. [18] developed empirical models relating the surface roughness and MRR of EN8 steel with the process parameters such as pulse-on time, pulse-off time, and discharge current. They used a multi-objective optimization tool, NSGA II, to obtain the pareto-optimal set of solutions. Though much research has been done in the field of the machining of stainless steel, the optimization of machining parameters of microEDM of SS316L has not been addressed. 1 EXPERIMENTAL DETAIL The stainless steel (316L) considered in this research is a metal used in pharmaceuticals, marine and medical applications. It has a significant role in medical implants, including pins, screws and

orthopaedic implants, such as total hip and knee replacements, due to various mechanical properties, such as high oxidation resistance, corrosive resistance and hardness. Though there are many process parameters that influence the machinability criteria of microEDM, this research dealt with three important processes: parameter-discharge current, pulse-on time Ton, and pulse-off time Toff. The Taguchi method was initially applied to determine the optimum process parameters and the number of experiments required to model response functions. RSM was then successfully applied to relate the input process parameters and the output responses of the selected material. The mathematical model obtained from RSM was then used as a fitness function for GA multi objective optimization. A schematic of the experiment was performed in a SPARKONIX microEDM machine as shown in Fig. 2 with a brass electrode (diameter: 400 µm) and deionized water as a dielectric fluid for machining the selected 316L stainless steel work piece.

Fig. 2. Sparkonix microEDM machine

1.1 Design of Experiment (DOE) The Taguchi method is a powerful approach that provides a simple, efficient and systematic approach to determine the optimum process parameters, which drastically reduces the number of experiments that are required to model response functions [7] and [8]. It is a method based on orthogonal array (OA) experiments, which provide the much-reduced variance for the experiment resulting in the optimum setting of process control parameters. The major influencing parameters and their levels considered are listed in Table 1. The selection of the

Optimization of Intervening Variables in MicroEDM of SS 316L using a Genetic Algorithm and Response-Surface Methodology

657

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

orthogonal array is based on the number of process parameters and their levels. In the current research, the L9 orthogonal array with three rows and nine columns is selected as given in Table 2. The tool wear rate for each experiment are calculated as: TWR =

Initial weight of tool − Final weight of tool . (1) Machining time

The material removal rate for each experiment is calculated as: Weight before machining − Weight after machining MRR = . (2) Machining time

Indentify the response functions and the process parameters

↓ Determine the number of levels for the process parameters and possible interactions between them

↓ Select the appropriate orthogonal array and conduct the experiment accordingly

↓ Analyse the experimental results and select the optimum level of process parameters

1.2 Response-Surface Methodology (RSM) In most RSM problems, the form of the relationship between the response and the independent variables is unknown. Thus, the first step in RSM is to find a suitable approximation for the actual relationship between the response and the process parameters. The quantitative form of relationship between the desired response and independent input variables can be represented as:

y = f ( x1 , x2 , x3 ,..., xn ) + ε , (3)

where, y is the desired response, f is the response function (or response-surface), x1, x2, x3, …, xn are the independent input variables, and is the fitting error. The appearance of the response function looks like a surface curve while plotting the expected response of f. The identification of suitable approximation for f will determine whether or not the application of RSM is successful. The necessary data for building the response model are generally collected from the design of experiments. In the current research, the experimental data were fitted into a two-factor interaction (2FI) regression model. The general form of 2FI model is:

↓ Verify optimal process parameters through confirmation experiment

Fig. 3. Steps involved in the Taguchi method

Table 1. Machining parameters and their levels Parameters Discharge current [A] Pulse-on time [µs] Pulse-off time [µs]

Level 1 6 3 3

Level 2 9 6 6

Level 3 12 9 9

i =1

i< j

f = β 0 + ∑β i xi + ∑β ij xi x j + ε , (4)

where, f is the desired response, βi represents the linear effect of xi , βij, represents the quadratic effect of xi. They are cross-product terms that reveal a linearby-linear interaction between xi and yi. ε is a statistical error term. Design Expert R7.0 software was used to obtain regression models for two responses separately.

Table 2. Experimental design using L9 orthogonal array Experiment 1 2 3 4 5 6 7 8 9

658

Machining Parameter Level Discharge current Pulse-on time Pulse-off time [A] [µs] [µs] 6 3 3 6 6 6 6 9 9 9 3 6 9 6 9 9 9 3 12 3 9 12 6 3 12 9 6

Total machining time [s]

TWR [mg/s]

MRR [mg/s]

306.45 140.30 108.09 109.84 76.30 74.20 51.34 67.81 42.54

0.02576 0.07466 0.08077 0.09538 0.17163 0.25884 0.25506 0.32185 0.26678

0.64545 0.94500 1.35233 1.36724 1.86400 1.37924 2.58245 2.06539 3.20636

Suresh, P. – Venkatesan, R. – Sekar, T. – Elango, N. – Sathiyamoorthy, V.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

The mathematical model correlating MRR with the process control parameters is obtained as: MRR = 0.66885 + 0.076855 x1 – 0.15883 x3 – – 0.31122 x1 + (0.011663 x1 x2) + + (0.028113 x1 x3) + (0.028438 x2 x3).

(5)

The mathematical model correlating TWR with the process control parameters is obtained as: TWR= – 0.384951429 + 0.053079841 x1 + + 0.05632381 x1 – 0.003937302 x3 – – (0.003577937 x1 x2) + (0.000419524 x1 x3) – (6) – (0.001784127 x2 x3) , where x1 is discharge current, x2 pulse-on time, and x3 pulse-off time. Figs. 4 and 5 show the linear correlation between the predicted values and the actual values of MRR and TWR. In the ANOVA test, if p value is less than 0.05, the developed model is significant; otherwise, it is insignificant. The coefficient of determination (R²) and Adj. R² from the ANOVA test in MRR are observed to be 0.9833 and 0.9331, respectively. Similarly, from the ANOVA test, in TWR, R² = 0.9656 and Adj. R² = 0.8622, which proves that the developed model is statistically considerable.

Fig. 4. Linear correlation between actual values and predicted values of MRR

1.3 The Effect of Discharge Current on MRR and TWR with Various Pulse-on Time Experiments were conducted on the chosen stainless steel 316L with 400 μm brass electrode. The discharge current, pulse-on time and pulse-off time with three levels were selected as major influencing parameters. The effect on MRR and TWR when Ton =3, 6 and 9 μs are presented in Fig. 6a, b and c, respectively. From Fig. 6a, it is observed that when Ton = 3 µs, the MRR increases linearly from 0.645 mg/s to 2.582 mg/s. The rate of increase in MRR varies on the discharge current. The rate of change in MRR is 0.18 mg/s in the range of 6 to 9 A, and it is 0.61 mg/s in the range of 9 to 12 A. In the case of TWR, it is 0.02576 mg/s at 6 A and 0.2551 mg/s at 12 A, but it is evident that there is a drastic linear increase in TWR between 9 and 12 A. In the case of Ton = 6 µs, MRR is 0.945 to 2.06539 mg/s at 6 to 12 A. It increases linearly at the rate of 0.22975 mg/s from 6 to 9 A, beyond which, there is no significant increase in MRR. The TWR is 0.0746 mg/s at 6 A, and it increases linearly at the rate of 0.02424 mg/s until 9 A, beyond which a sudden increase is noticed.

Fig. 5. Linear correlation between actual values and predicted values of TWR

When Ton = 9 µs, there is no effect of the change in MRR between 6 and 9 A, but it suddenly increases to 3.2 mg/s at 12 A. In the case of TWR, the linear increase from 0.0807 to 0.2588 mg/s is noticed from 6 to 9 A, beyond which it is found to be almost constant. 1.4 Effect of Discharge Current on MRR and TWR with Various Pulse-off Time The effect on MRR and TWR with respect to Toff = 3, 6 and 9 μs is plotted in Fig. 7a, b and c respectively.

Optimization of Intervening Variables in MicroEDM of SS 316L using a Genetic Algorithm and Response-Surface Methodology

659

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

Fig. 6. Effect of discharge on MRR and TWR when; a) Ton=3 µs, b) Ton= 6 µs, and c) Ton=9 µs

In the case of 3 μs, the MRR at 6 A is 0.6455 mg/s and it increases very linearly to 2.0654 mg/s at 12 A. The TWR is 0.02576 mg/s at 6 A and increases to 0.2588 mg/s at 9 A and then increases to 0.32185 mg/s at 12 A. 660

Fig. 7. Effect of discharge on MRR and TWR when; a) Toff=3 µs, b) Toff=6 µs, and c) Toff = 9 µs

In the case of 6 μs, the MRR and TWR at 6 A are 0.945 mg/s and 0.07466 mg/s respectively. The trend of increase in both MRR and TWR are almost same. A sudden increase is observed from 9 to 12 A.

Suresh, P. – Venkatesan, R. – Sekar, T. – Elango, N. – Sathiyamoorthy, V.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

In the case of 9 μs, the MRR is 1.352 mg/s at 6 A, 1.864 mg/s at 9 A and 2.5824 mg/s at 12 A. The TWR is 0.08077 mg/s at 6 A, 0.1716 mg/s at 9 A and 0.25506 mg/s at 12 A. 2 MULTI-OBJECTIVE OPTIMIZATION USING A GENETIC ALGORITHM The optimization seeks to minimize or maximize the value of a function in a given search space. Evolutionary algorithms are popular as robust and effective methods for solving optimization problems. These algorithms apply the principle of survival of the fittest to find the best approximations. A new set of approximations is created at each generation by the process of selecting individual potential solutions (individuals) according to their level of fitness in the problem domain and breeding them together using operators borrowed from natural genetics. This process leads to the evolution of populations of individuals that are better suited to their environment. A wide range of evolutionary algorithms for multi-objective optimization is available. An NSGA is one of the second generation evolutionary algorithms proposed by Deb et al. [19] and [20]. Many authors have discussed evolutionary algorithms in their research [21] to [23]; the multi-objective problem [24] to [26] is comprehensively dealt with. In recent years, several other algorithms, such as ant colony optimization (MOACO) [27], artificial immune systems [28], and particle swarm optimization (MOPSO) [29] have also been used in multi-objective

optimization. These kinds of algorithms have also been applied in manufacturing processes [30] to [32]. These heuristic algorithms [33] to [35] are mainly applied for optimal search. Optimization based on using meta-heuristic algorithms starts with an initial set of independent variables and then evolves to obtain the global minimum/maximum of the objective (fitness) function. The objective function is a mathematical model (function) that assigns a value to each solution in the search space. Starting from an initial solution built with some heuristics, meta-heuristics improve it iteratively until a stopping criterion is met. The NSGA-II considered in this paper is a fast non-dominated sorting approach with computational complexity is introduced, where is the number of objectives and is the population size. It is a steady-state genetic algorithm, which is more suitable for machining applications. The MATLAB GA multi objective tool box was applied to predict the optimum process parameters. The mathematical models developed using RSM were used in tool box as fitness functions. The objectives are to maximize the MRR minimize the TWR. In order to convert objective for minimization, it is suitably modified. The objective functions are framed as;

Objective 1 = 1 / MRR,

Objective 2 = TWR,

subject to:

6 ≤ x1 ≥ 12,

3≤ x2 ≥9,

3≤ x3 ≥9.

Fig. 8. Obtained optimal solutions from GA Optimization of Intervening Variables in MicroEDM of SS 316L using a Genetic Algorithm and Response-Surface Methodology

661

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

The GA generally includes three fundamental genetic operations of selection, crossover and mutation. These operations are used to modify the chosen solutions and select the most appropriate offspring to pass on to the succeeding generations. A population size of 45, a cross-over fraction of 0.8 and a scattered cross-over function were selected from the tool. The tool considered a two-point cross-over function by default. The mutation rate was observed to be 0.01. 2.1 Results from the Multi objective GA The observed responses corresponding to control parameters are listed in Table 3. The multi-objective GA predicts low MRR of 0.4352 mg/s and TWR of 0.0122 mg/s corresponding to Ton = 3.3608 μs, Toff = 8.6356 μs and discharge current is 6.0263 A. The high MRR is observed with TWR = 0.2391 mg/s at the condition Ton = 8.9999 μs, Toff = 8.9185 μs, and the discharge current is 11.9991 A. Table 3. Process decision variables corresponding to each of optimal solution point and the predicted responses using GA Control Parameters Pulse-on Pulse-off Discharge time (Ton) time (Toff) [A] 3.3608 8.6356 6.0263 8.9999 8.9185 11.9991 8.8931 8.9147 10.3863 7.2814 8.8935 8.3505 8.6525 8.8813 10.8190 8.7242 8.8797 9.8730 7.8617 8.8742 8.6471 8.3824 8.6987 7.2569 8.1827 8.8591 9.8165 4.0178 8.6923 6.4205 8.2307 8.8723 9.3832 5.4919 8.8640 7.0884 8.1926 8.8996 8.9381 3.4702 8.7186 7.1761 8.9546 8.9166 11.0506 4.9293 8.8463 6.9434

Responses MRR TWR [mg/s] [mg/s] 0.4352 0.0122 3.9366 0.2391 3.2150 0.1991 2.0248 0.1315 3.3404 0.2100 2.9507 0.1859 2.2571 0.1464 1.7457 0.1169 2.8097 0.1819 0.6819 0.0167 2.6393 0.1701 1.1875 0.0659 2.4486 0.1572 0.8655 0.0403 3.5157 0.2157 1.0325 0.0519

It is observed that Ton and discharge are to be set low for low MRR and must be set high for high MRR. It is also observed that when MRR increases, TWR also increases correspondingly. However, the objective of this research is to maximize MRR and minimize TWR. Hence, the obtained optimal solutions from GA is presented in Fig. 8. 662

2.2 Confirmation Test The confirmatory experiments were further conducted for the optimal parameters obtained from the MATLAB multi-objective GA. The error between optimum values from GA and the confirmation test was derived by considering Serial No. 6 from the Table 3, at the condition Ton = 8.7 μs, Toff = 8.9 μs and discharge 9.87 A, and is shown in Table 4. Table 4. Error between optimum values from GA and confirmation test value Obtained from GA [mg/s] MRR TWR 2.9506 0.1859

Confirmation test value [mg/s] MRR TWR 2.831 0.1954

Error [%] MRR 4.06

TWR 5

The average prediction error for MRR is 4.06%, and TWR is 5%. Thus, the GA predicted results are within the acceptable limits, thereby establishing the validity of the method proposed. 3 CONCLUSION A new attempt to optimize the intervening parameters in microEDM of Stainless Steel 316L using a 400 μm brass electrode was done. It was intended to obtain better MRR and TWR simultaneously. The discharge current, pulse-on time and pulse-off time with three levels were considered to be the major intervening parameters in microEDM of SS316L. The mathematical model was derived from RSM, and the result of it was used as a fitness function for multiobjective optimization using GA. The results reveal that the developed mathematical models significantly improve the chosen objectives of obtaining the better MRR and TWR. The multi-objective optimization processes have categorically revealed the interaction effects among the chosen intervening parameters. The optimization model was developed by simultaneously considering the maximization of MRR and minimization of TWR, which is highly useful for real life applications. It is evident from the confirmation results that the developed mathematical model yields the results with a deviation of 5% from the experimentation. 4 ACKNOWLEDGEMENT We acknowledge the financial assistance provided by All India Council for Technical Education (AICTE), New Delhi under the RPS File no 8023/BOR/RID/ RPS-96 /2009-10 and Sona College of Technology.

Suresh, P. – Venkatesan, R. – Sekar, T. – Elango, N. – Sathiyamoorthy, V.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

We also thank the management of Sona College of Technology for the continuous encouragement and support. 5 REFERENCES [1] Mahendran, S., Devarajan, R., Nagarajan, T., Majdi, A. (2010). A review of microEDM. Proceedings of the International MultiConference of Engineers and Computer Scientists, vol. 2, Hong Kong. [2] Joshi, S.N., Pande, S.S. (2011). Intelligent process modeling and optimization of die – sinking electric discharge machining. Applied Soft Computing, vol. 11, no. 2, p. 2743-2755, DOI:10.1016/j.asoc.2010.11.005. [3] Mohamad, A.B., Siddique, A.N., Quadir, G.A., Khan, Z.A., Saini, V.K. (2012). Optimization of EDM parameters using Taguchi method. Proceeding of International Conference on Applications and Design in Mechanical Engineering, Penang. [4] Rose, P.J. (2005). Taguchi Techniques for Quality Engineering (2nd ed.) Tata McGraw Hill, New Delhi. [5] Manna, A., Salodkar, S. (2008). Optimization of machining conditions for effective turning of E0300 alloy steel. Journal of Materials Processing Technology, vol. 203, no. 1-3, p.147-153, DOI:10.1016/j. jmatprotec.2007.09.052. [6] Natarajan, N. Arunachalam, R.M. (2011). Optimization of microEDM with multiple performance characteristics using Taguchi and Grey relational analysis. Journal of Scientific and Industrial Research, vol. 70, no. 7, p 500-505. [7] Dhanabalan, S., Sivakumar, K., Sathiya Narayanan, C. (2012). Optimization of EDM process parameters with multiple Performance characteristics for Titanium Grades. European Journal of Scientific Research, vol. 68, no. 3, p. 297-305. [8] Mukherjee, I., Ray, P.K. (2006). A review of optimization techniques in metal cutting processes. Computers & Industrial Engineering, vol. 50, no. 1-2, p. 15-34, DOI:10.1016/j.cie.2005.10.001. [9] Chiang, K.-T., Chang, F.-P. (2006). Application of response surface methodology in the parametric optimization of a pin-fin type heat sink. International Communications in Heat and Mass Transfer, vol. 33, no. 7, p. 836-845, DOI:10.1016/j. icheatmasstransfer.2006.04.011. [10] Rajesh, R., Dev Anand, M. (2012). The optimization of the electro-discharge machining process using response surface methodology and genetic algorithms. Procedia Engineering, International Conference on Modelling, Optimization and Computing, vol. 38, p. 3941-3950, DOI:10.1016/j.proeng.2012.06.451 [11] Karthikeyan, G., Ramkumar, J., Dhamodaran, S., Aravindan, S. (2010). Micro electric discharge milling process performance: an experimental investigation. International Journal of Machine Tools & Manufacture, vol. 50, no. 8, p. 718-727, DOI:10.1016/j. ijmachtools.2010.04.007.

[12] Kung, K.Y., Horng, J.T., Chiang, K.T. (2009). Material removal rate and electrode wear ratio study on the powder mixed electrical discharge machining of cobaltbonded tungsten carbide. International Journal of Advanced Manufacturing Technology, vol. 40, no. 1-2, p. 95-104, DOI:10.1007/s00170-007-1307-2. [13] Samtaş, G., Çiçek, A., Kıvak, T., Çay, Y. (2012). Modeling of thrust forces in drilling of AISI 316 stainless steel using artificial neural network and multiple regression analysis. Strojniški vestnik – Journal of Mechanical Engineering, vol. 58, no. 7-8, p. 492-498, DOI:10.5545/sv-jme.2011.297. [14] Saric, T., Simunovic, G., Simunovic, K. (2013). Use of neural networks in prediction and simulation of steel surface roughness. International Journal of Simulation Modelling, vol. 12, no. 4, p. 225-236, DOI:10.2507/ IJSIMM12(4)2.241. [15] Kao, P.S., Hocheng, H. (2003). Optimization of electrochemical polishing of stainless steel by grey relational analysis. Journal of Materials Processing Technology, vol. 140, no. 1-3, p. 255-259, DOI:10.1016/ S0924-0136(03)00747-7. [16] Lee, S.S.G., Tam, S.C., Loh, N.H. (1993). Ball burnishing of 316L stainless steel. Journal of Materials Processing Technology, vol. 37, no. 1-4, p. 241-251, DOI:10.1016/0924-0136(93)90094-M. [17] Bharti, P.S., Maheshwari, S., Sharma, C. (2012). Multiobjective optimization of electric-discharge machining process using controlled elitist NSGA-II. Journal of Mechanical Science and Technology, vol. 26, no. 6, p. 1875-1883, DOI:10.1007/s12206-012-0411-x. [18] Baraskar, S.S., Banwait, S.S., Laroiya, S.C. (2013). Multi objective optimization of electrical discharge machining process using a hybrid method. Materials and Manufacturing Processes, vol. 28, no. 4, pp. 348354, DOI:10.1080/10426914.2012.700152. [19] Deb, K., Agrawal, S., Pratap, A. (2000). A fast elitist non-dominated sorting genetic algorithm for multiobjective optimization: NSGA-II. Proceedings of the 6th International Conference on Parallel Problem Solving from Nature, Paris, p. 849-858. [20] Deb, K., Agrawal, S., Pratap, A., Meyarivan, C. (2002). A fast and elitist multiobjective genetic algorithm: NSGA-II. Evolutionary Computation, vol. 6, no. 2, p. 182-197, DOI:10.1109/4235.996017. [21] Goldberg, D.E. (1989). Genetic Algorithms in Search, Optimization & Machine Learning. Addison-Wesley, Indianapolis. [22] Srinivas, N., Deb, K. (1994). Multiobjective optimization using nondominated sorting in genetic algorithms. Evolutionary Computation, vol. 2, no. 3, p. 221-248, DOI:10.1162/evco.1994.2.3.221. [23] Horn, J., Nafpliotis, N., Goldberg, D.E. (1994). A niched Pareto genetic algorithm for multiobjective optimization. Proceedings of the 1st IEEE Conference on Evolutionary Computation, Orlando, p. 82-87. [24] Bentley, P.J., Wakefield, J.P. (1997). Finding acceptable solutions in the Pareto-optimal range using

Optimization of Intervening Variables in MicroEDM of SS 316L using a Genetic Algorithm and Response-Surface Methodology

663

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 656-664

multiobjective genetic algorithms. Proceedings of the 2nd On-Line World Conference on Soft Computing in Engineering Design and Manufacturing, p. 231-240. [25] Zitzler, E., Thiele, L. (1999). Multiobjective evolutionary algorithms: A comparative case study and the strength Pareto approach. IEEE Transactions on Evolutionary Computation, vol. 3, no. 4, p. 257-271, DOI:10.1109/4235.797969. [26] Deb, K. (1999). Multi-objective genetic algorithms: Problem difficulties and construction of test problem. Journal of Evolutionary Computation, vol. 7, no. 3, p. 205-230, DOI:10.1162/evco.1999.7.3.205. [27] Garca, C., Cordn, O., Herrera, F. (2004). An empirical analysis of multiple objective ant colony optimization algorithms for the Bi-criteria TSP. ANTS Workshop, p. 61-72. [28] Cruz-Corts, N., Coello, C.A. (2003). Multiobjective optimization using the clonal selection principle immune system. 9th Electrical Engineering Conference, p. 470-477. [29] Durillo, J., Garca, J., Nebro, A., Coello, C., Luna, F., Alba, E. (2009). Multi-objective particle swarm optimizers: An experimental comparison. 5th International Conference on Evolutionary MultiCriterion Optimization, Nantes, p. 495-509. [30] Ciurana, J., Arias, G., Özel, T. (2009). Neural network modelling and particle swarm optimization (PSO) of process parameters in pulsed laser micromachining of hardened AISI H13 steel. Materials and

664

Manufacturing Processes, vol. 24, no. 3, p. 358-368, DOI:10.1080/10426910802679568. [31] Zain, A.M., Haron, H., Sharif, S. (2010). Application of GA to optimize cutting conditions for minimizing surface roughness in end milling machining process. Expert Systems with Applications, vol. 37, no. 6, p. 4650-4659, DOI:10.1016/j.eswa.2009.12.043. [32] Rao, R.V., Pawar, P.J., Davim, J.P. (2010). Optimisation of process parameters of mechanical type advanced machining process using a simulated annealing algorithm. International Journal of Materials and Product Technology, vol. 37, no. 1-2, p. 83-101, DOI:10.1504/IJMPT.2010.029461. [33] Samanta, S., Chakraborty, S. (2011). Parametric optimization of some non-traditional machining processes using artificial bee colony algorithm. Engineering Applications of Artificial Intelligence, vol. 24, no. 6, p. 946-957, DOI:10.1016/j. engappai.2011.03.009. [34] Madić, М., Marković, D., Radovanović, M. (2012). Optimization of surface roughness when turning polyamide using ANN-IHSA approach. International Journal of Engineering & Technology, vol. 1, no. 4, p. 432-443, DOI:10.14419/ijet.v1i4.378. [35] Yildiz, A.R. (2013). Cuckoo search algorithm for the selection of optimal machining parameters in milling operations. The International Journal of Advanced Manufacturing Technology, vol. 64, no. 1-4, p. 55-61, DOI:10.1007/s00170-012-4013-7.

Suresh, P. – Venkatesan, R. – Sekar, T. – Elango, N. – Sathiyamoorthy, V.

Received for review: 2014-01-17 Received revised form: 2014-04-20 Accepted for publication: 2014-05-07

Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve Liu, L. – Zhang, D. – Zhao, J. Liu Lei1 – Zhang Desheng2,3 – Zhao Jiyun1,* 1 China

University of Mining and Technology, School of Mechanical and Electrical Engineering, China 2 Coal Mining Technology Department, Tiandi Science and Technology, China 3 China University of Mining and Technology Beijing, School of Mechanical Electronic & Information Engineering, China The basic requirements of the valve-control coupling for control valves are quick response, high flow capacity, and anti-blocking ability. In this article, a pure water hydraulic test-bed is built, and the pressure distribution characteristics between the throttle nozzle and the pilot valve, as well as the flow characteristics in the throttle orifice are researched. Following that, the water medium low pressure large-flow pilot operated solenoid valve is designed. The influences of throttle nozzles’ diameter and other parameters on the dynamic and static characteristics of the valve are analysed via AMESim simulation software. The optimized combination of parameters of the solenoid valve group is determined. The experiment results show that the normal working pressure drop of the designed solenoid valve group is approximately 0.07 MPa, the opening time is 0.3 to 0.4 s, the closing time is approximately 1 s; it possesses better low pressure characteristics and rapid response characteristic, which can satisfy the requirements of valve-control coupling. Keywords: water medium, low pressure, large-flow, pilot operated solenoid valve, throttle orifice, main valve structure

0 INTRODUCTION In hydraulic systems, a pure water hydraulic system is a new development direction in fluid transmission and control field [1] and [2]. Compared with conventional mineral hydraulic oil, water as a medium has certain unique advantages, such being clean/non-polluting, easily available and resistant to explosion [3]; it has been increasingly used widely in food, fire, high pressure cleaning, reactor, mining and other industries [4] and [5]. Recent research on water electromagnetic valves has mostly focussed on high-pressure valves [6], rather than on low-pressure large-flow valves. Meanwhile, most of the research is on the flow field characteristics of existing valves, but little has been done on the structural design of valves. Low-pressure large-flow valves with fast response characteristics are highly valued in industrial applications, such as high-power valve-control coupling, in which low-pressure large-flow solenoid valves are a core component. It helps to achieve gentle starts of valve-control coupling by controlling the liquid-filling processes. The working liquid can be replaced in a timely manner according to the liquid temperature and pressure in the chamber to control the speed and limit the working temperature. Valvecontrol coupling has been widely used in heavy scraper conveyors of coal mines, belt conveyers, pump, draught fans and other heavy equipment; it plays an important role in improving working conditions and saving energy [7] and [8]. The lowpressure large-flow valve group is one of the key factors of high-power valve-control coupling.

The control valve for valve-control coupling can be classified as a solenoid valve. Low-pressure largeflow solenoid valves with piston liftd and diaphragms are both using pilot valves structure. Some research about solenoid valve for valve-control coupling provides a good basis for this study [9] to [11]. Other related research about valve design [12], simulation [13] and improvement [14] is also encouraging. However, the application requirement of the solenoid control valve group is stricter in some difficult operating environments. This article describes low-pressure large-flow pure pilot-operated water solenoid valves (the flow is larger than 240 L/min) according to the working requirements of the valve-control coupling. The effect of key parameters on valve group characteristics is analysed using AMESim in order to seek reasonable parameters. The pure water hydraulic test platform is set up to carry out experimental verification. 1 WORKING PRINCIPLE OF THE SOLENOID VALVE GROUP 1.1 Working Principle of Differential Pressure Type Pilot Operated Solenoid Valve The differential pressure type pilot-operated solenoid valve is constituted with a main valve and a solenoid pilot valve, as seen in Fig. 1. In the pilot hydraulic half-bridge, R1 and R2 represent the throttle orifice fluid resistance and the pilot valve fluid resistance, respectively. Both are connected in series. The supply fluid pressure is p1,

*Corr. Author’s Address: School of Mechanical and Electrical Engineering, China University of Mining and Technology, Xuzhou, 221116, China, jyzhao@cumt.edu.cn

665

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

and upper chamber pressure of the main spool is p2. The relationship between them is:

p2 =

R2 p1. (1) R1 + R2

The role area of the control chamber (upper chamber) is A2. The role area of the high-pressure chamber (lower chamber) is A1. Here A1 < A2. The spring stiffness in the upper chamber is k.

Define

A2 R = k1 , 1 = k2 , thus A1 R2  k A1 p1 1 − 1 1 + k2 

  = kx0 + Ff . (4) 

Ignore the spring force and the frictional force. Thus: kx0 + Ff = 0. And open the parameters condition: R1 > ( k1 − 1) R2 .

Then it is expressed by pressure: ∆p1 > ( k1 − 1) ∆p2 . (5)

If the pilot valve flow capacity is strong, which means R2 is 0, the opening pressure parameter condition is:

Fig. 1. Working principle of pilot operated solenoid valve

When the pilot valve is closed, the working medium flows into the upper chamber through the orifice, and it cannot continue because of resistance. The fluid resistance R2 is infinite at this time. From Eq. (1) it can be known that p1 is equal to p2. The greater the work pressure, the tighter seal. When the pilot valve is opened, the pressure drops are formed at the orifice and the pilot valve. So R2 is small, and by the Eq. (1) it can be known that p1 is greater than p2. When p1 can overcome, the summation of p2, spring force and friction brought by spool movement, the main spool will open. 1.2 Spool Motion Equation The Spool motion equation is:

p1 A1 − p2 A2 = mx − D1 x + k ( x0 + x ) + Ff + Fs . (2)

Here m is the spool’s mass, x is spool displacement, and x0 is spring precompression. D1 is damped coefficient. Ff is the friction force. Fs is flow force. The static equilibrium condition before main valve opening is: 666

p2 A2 + kx0 + Ff = p1 A1. (3)

p1 >

( kx

+ Ff A1

) . (6)

As can be seen from Eq. (6), the spring mainly affects the opening pressure. In consideration of the reset, the smaller pre-tightening force is required because of the reducing opening pressure. 1.3 Characteristics of Throttle Orifice The diameter d and length l are two basic parameters of the throttle nozzle. When the l/d (length diameter ratio) is between 0.5 and 4, it is called a ‘short orifice’. When l/d is greater than 4, it is called a ‘thin long orifice’. The short orifice flow calculation equation is similar to that of the thin-walled holes.

q = Cd a0

2∆p . (7) ρ

In the equation, a0 is the cross-sectional area of the orifice. Cd is flow coefficient, which is approximately 0.8 when the Reynolds number is larger. Δp is the pressure difference at orifice. Regarding laminar flow in the thin long orifice, the flow through circular tube calculation equation is:

πd4 ∆p. (8) 128µ l

Therefore, the flow rate equation of the orifice Eqs. (1) and (2) can be summarized as: q = Ca0 ∆p m . (9)

Liu, L. – Zhang, D. – Zhao, J.

StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

The coefficient C is determined by the shape and dimensions of the orifice and the general nature of the liquid. The coefficient m (0.5 < m < 1) is determined by the length diameter ratio of the orifice. The orifice is at the transitional state of laminar flow and viscous flow. 1.4 Fluid Resistance Characteristic According to the working principle of the pilot solenoid valve, the pressure distribution test system is shown in Fig. 2. The pilot valve and throttle nozzle are connected in a series to simulate the relationship between the throttle nozzle and the pilot valve. The characteristics of â&#x20AC;&#x2DC;pressure difference-flowâ&#x20AC;&#x2122; and the pressure distribution rule for the throttle nozzle and pilot valve are investigated using this test system. The solenoid pilot valve is turned on while electrified. Fig. 3. shows the throttle nozzle structure.

4 is pressure-flow curves of different nozzle lengths, which shows that with increasing length of throttle nozzle, the flow capacity of the orifice is weakened. However, with further increases, the rate of flow decreases become slower. This means that sensitivity of the flow capacity is reduced according to the length. When the flow coefficient Cd is 0.7, the theoretical value of throttle nozzle is calculated by using Eq. (7) as plotted in Fig. 4. It shows that the theoretical value is significantly larger. Although the effect of length on the flow characteristics is not considered in Eq. (7), it can also be seen that the overall trend of each experimental curve substantially presents the exponent distribution law, which means that the flow of the water medium in the orifice is a turbulent-state flow. As the viscosity of water is small, the resistance loss along the way of the orifice is also low. The flow through the throttle nozzle is hardly affected by the viscosity in the test range, which means it is insensitive to change for water temperature. The temperature of water in the valve-control coupling changes tremendously in start and overloading conditions. Therefore, the feature is particularly suitable for valve-control coupling.

Fig. 2. Water hydraulic system for pressure dividing test

Adopt 2, 1.8, 1.5 and 1.2 mm four different throttle nozzles matching 3 mm pilot valve. Select 1.2 mm diameter throttle nozzle matching 1.5 mm pilot valve. Then compare the matching effect.

Fig. 3. Throttle nozzle structure

1.5 Flow Characteristics of Throttle Nozzle The diameters of throttle nozzles are 1.2 mm, and the lengths are 3.5, 5, 6 and 7 mm respectively. Fig.

Fig. 4. Pressure-flow curves of different nozzle lengths

When diameters of the throttle nozzle are 1.2, 1.5, 1.8 and 2 mm, the length diameter ratio is 3.5, and the corresponding lengths are 4.2, 5.2, 6 and 7 mm, respectively. Characteristics of flow in each throttling nozzle are shown in Fig. 5. It can be seen that in case of the same length diameter ratio, the larger the diameter, the stronger flow capacity. In accordance with the principles of the least squares method, fitting every pressure-flow curve according to the exponent function, here the exponent is approximately 0.5, which is closer to Eq. (9) and meets the flow characteristics of a short orifice.

Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve

667

StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

Fig. 5. Pressure-flow curves of different nozzle diameters

1.6 Differential Pressure Distribution Rule Adopt four different throttle nozzles (2, 1.8, 1.5 and 1.2 mm) with matching 3 mm pilot valves; the lengths of throttle nozzles are all 4 mm. Fig. 6 shows pressure difference-flow curve. Adopting a 1.2 mm orifice, when the system pressure is 1.9 MPa, the flow is 2.15 L/min. The pressure drop at the orifice accounts for the majority of the total pressure. As a result of the small flow, the

pressure drop at the pilot valve is very small. When the pressure is 2 MPa at orifice diameters of 1.5 mm, the flow is 3.8 L/min. For a 1.8 mm orifice, when the pressure is 2 MPa, the flow is 5.5 L/min. When the pressure is 2 MPa at orifice diameters of 2 mm, the flow is 6 L/min. Therefore, when increasing the orifice diameter, the pressure drop of the pilot valve at the same flow rate is also increasing. Thus, the corresponding solenoid valve upper chamber pressure is increased. If the diameter of the orifice is further increased, the open condition of the Eq. (5) will not be satisfied. In comparing Fig. 7 with Fig. 6, the pressure drop brought by a 1.5 mm diameter pilot valve nearly accounts for total pressure drop of 1/3 for 1.2 mm throttle nozzle. However, the pressure drop proportion in total pressure brought by 3 mm diameter pilot valve is almost negligible. In considering the working pressure, the bigger diameter pilot valve should be selected. 2 DESIGN OF THE MAIN VALVE Fig. 8 shows a hydraulic system of valve-control hydrodynamic coupling. It comprises a control system with an electro-hydraulic valve group as the main part

a) d = 1.2 mm

c) d = 1.8 mm d) d = 2 mm Fig. 6. Pressure distribution between 3 mm pilot valve and different throttle nozzles

668

Liu, L. â&#x20AC;&#x201C; Zhang, D. â&#x20AC;&#x201C; Zhao, J.

b) d = 1.5 mm

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

and a hydrodynamic system taking coupling as the core part.

Fig. 7. Pressure distribution between 1.5 mm pilot valve and throttle nozzle (d = 1.2 mm)

Fig. 8. Hydraulic system of valve-control hydrodynamic coupling; 8-1 flow controller, 8-2 filling valve, 8-3 solenoid pilot valve (2-position 2-way), 8-4 circulating valve, 8-5 solenoid pilot valve (2-position 3-way), 8-6 discharging valve, 8-7 coupling, 8-8 cooler, 8-9 relief valve, 8-10 motor, 8-11 load, 8-12 pressure transducer, 8-13 temperature transducer, 8-14 velocity transducer

The control valve group works according the pressure-difference pilot principle. It consists of the filling valve (8-2), the liquid-discharged valve (8-6) and the circulating valve (8-4). The pilot valve (8-3) controls the filling valve (8-2). The solenoid valve (83) is used as a pilot valve, and the filling valve (8-2) is used as the main valve. There is a damping orifice on the main spool. When the solenoid valve (8-3) is opened, there will be a pressure difference between the upper and lower cavities, the main valve will open under the action of pressure difference, and then the coupling will be filled with liquid. The two-position three-way solenoid pilot valve (8-5) can control both the on-off state of the circulating valve (8-4) and

the on-off state of the discharging valve (8-6). The circulating valve and the liquid-discharged valve have the same principle as the liquid-filled valve (82). The movement of the main spool is also caused by the pressure difference between the upper and lower cavities. The circulating valve is mainly to control the on-off state of coupling circuits, and the liquiddischarged valve is used to discharge the liquid inside coupling. Only one of the two valves could open under the action of the pilot valve (8-5). The motor rotating speed is 1490 r/min. The essence of the valve-controlled coupling catheter is a rotary jet pump [15]. Therefore, the theoretical maximal value that the coupling catheter can provide is approximately 0.6 to 0.88 MPa [16]. For a pure plane structure (Fig. 9a), during the process of fast reversing of the spool, a sudden block will cause sharply pressure rising inside the tube and lead to strong shock and noise. In order to relieve shock and noise, a cone is set at outlet surface of the plane valve. The fixed damping orifice is put on the main spool. The parameters of the tail cone structure are shown in Table 1.

a) Pure plane structure b) Tail cone structure Fig. 9. Structure of the main spool Table 1. Lists of parameters Parameters Value [mm]

α [°]

45°

3 SIMULATION OF THE CHARACTERISTICS OF SOLENOID VALVE BASED ON AMESIM The liquid-filled valve is studied as a sample in order to investigate the influences of the throttle nozzle, spring and other parameters on the dynamic characteristics, static characteristics and opening pressure. Following that, the combination of optimized parameters of the solenoid valve group can be determined, and whether spool structure could satisfy the function requirements of valve-control coupling can also be checked.

Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve

669

StrojniĹĄki vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

3.1 Building a Simulation Model The simulation model of the liquid-filled valve is established in AMESim, as shown in Fig. 10. It is composed of a pilot-operated solenoid valve and an external fluid supply system. The fluid supply system is provided liquid by a quantitative pump, and the fluid supply pressure is regulated by a relief valve. The stationary liquid resistance represents the characteristics of a damping orifice. The combination of the plane valve and poppet valve represents the main spool with tail cone structure.

opened. In order to obtain a smaller opening pressure, considering the need of the anti-block ability, a 1.5 mm diameter throttle nozzle is selected.

Fig. 11. Comparisons of opening pressure of different nozzle diameters

Fig. 10. Simulation model

3.2 Simulation Results and Analysis 3.2.1 Opening Pressure The signal source is set to ensure that the opening pressure of relief valve increases linearly from 0 to 1.6 MPa. In addition, the valve remains open while the pump flow rate is set to 240 L/min. The stationary liquid resistance is given by the pressure differenceflow data of a 3 mm pilot valve. The main spool spring stiffness is 10 N/mm. The pressure difference-flow data (obtained by the test) of 1.2, 1.5, 1.8 and 2 mm throttle nozzles are assigned to the liquid resistance. It can be seen from Fig. 11 that before opening, the pressure at position A increased synchronously with the relief valve setting pressure. However, once opened, it declines sharply. The value of the inflection point is the minimum opening pressure. The opening pressure is about 0.16 MPa when using a 1.2 mm throttle nozzle, 0.19 MPa when using a 1.5 mm throttle nozzle and 0.225 MPa when using a 1.8mm throttle nozzle. The minimum opening pressure increases to 0.23 MPa when using a 2 mm throttle nozzle. If the throttle nozzle is further enlarged, the opening pressure will rise further until it cannot be 670

As a result of the main spool with a tail cone, it is necessary to observe the pressure at inlet B. Fig. 12 shows the pressure curves of position A and position B when using a 1.5 mm throttle nozzle. When the pressure is 0.13 MPa, the pressure at B begins to rise, and the pressure at A continues to increase with an increasing relief valve. When the pressure exceeds 0.13 MPa, the spool is slightly opened, and the relief valve is still working. However, the flow through the main spool is small, which is in the non-effective open state. The setting pressure of the relief valve continues to rise. When the opening pressure of the spool is 0.19 MPa, the relief valve closes. Thus, all liquid comes through the main spool, the distance of the spool is further enlarged, the pressure drop appears at the valve orifice, which reduces until stability is attained. The pressure difference between position A and position B is the pressure drop produced by the plane valve. When spool is slightly opened, pressure drop is greater in the plane valve. When the relief valve is closed, the pressure drop at the plane circular throttle nozzle is less than 0.02 MPa.

Fig.12. Comparison of opening pressures at two stages

Liu, L. â&#x20AC;&#x201C; Zhang, D. â&#x20AC;&#x201C; Zhao, J.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

3.2.2 Stable Pressure Pump flow is set to 240 and 400 L/min. The relief valve pressure is set to 1.6 MPa. The spring stiffness is set to 10 N/mm. Simulation time is 2 s. Then, steady pressure at position A is observed. Fig. 13 shows that the greater the flow and orifice, the greater the pressure drop at position A. When the orifice is 1.2 mm and flow is 240 L/min, the pressure drop is minimum of 0.051 MPa. When the orifice is 2 mm and flow is 400 L/min, the pressure drop is maximum of 0.095 MPa. The pressure drop changes in a small range for each combination of parameters, which satisfies the requirements of low pressure and large flow.

In the Fig. 15, it can be seen that the response curves of opening and closing process are overlapping. The closing time is about 1.2 s and the opening time is less than 0.5 s. However, the opening time is close to closing time at different supply pressure. Furthermore, the fluctuation of fluid supply pressure has little effect on the overall response characteristics, i.e. this valve has good stability at low pressure.

Fig. 14. Response characteristics (different throttle nozzles)

Fig. 13. Pressure drop (different flows and throttle nozzles); 1-Q = 240 L/min, d = 1.2 mm; 2-Q = 240 L/min, d = 1.5 mm; 3-Q = 240 L/min, d = 1.8 mm; 4-Q = 240 L/min, d = 2.0 mm; 5-Q = 400 L/min, d = 1.2 mm; 6-Q = 400 L/min, d = 1.5 mm; 7-Q = 400 L/min, d = 1.8 mm; 8-Q = 400 L/min, d = 2.0 mm;

3.2.3 The Dynamic Response Characters The pressure of relief valve is set to 1.6 MPa; the flow is set to 240 L/min. The spring stiffness is set to 10 N/mm. The reversing valve of an on-off signal is given, and then the opening and closing time of the spool is studied. In order to measure and compare conveniently, the pressure at position A is taken as a reference (it would close down when it reaches to 1.6 MPa). The comparison of response time of throttle nozzles in different diameters (1.2, 1.5, 1.8 and 2 mm) are shown in Fig. 14; the blue line represents the input signal of the pilot valve, and high position represents ‘on’, zero position represents ‘off’. Therefore, the smaller the throttle nozzle, the longer the closing time. The closing time is 2.6 s by using 1.2 mm throttle nozzle, and the closing time reduces to 0.75 s by using a 2 mm throttle nozzle. The open time is within 0.5 s, which is hardly affected by the throttle nozzle. The effect of different fluid supply pressure on the opening and closing characteristics are shown in Fig. 15. The throttle nozzle is 1.5 mm in diameter.

Fig. 15. Response characteristics (different supply pressure)

4 EXPERIMENT AND DISCUSSIONS

Fig. 16. Experimental system of valve group; 1 water pump; 2 throttle valve; 3 filter; 4 flowmeter; 5 pressure transducer; 6 pressure transducer; 7 filling valve; 8 solenoid pilot valve; 9 circulating valve; 10 solenoid pilot valve; 11 discharging valve; 12 solenoid pilot valve

Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve

671

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

The principle of the valve group test system is shown in Fig. 16. The system pressure could be adjusted by manual throttle valve (2). The flow is 20 m³/h. The head of delivery is 162 m, the rotation rate is 2900 r/ min, and the power is 22 kW. The pressure, flow, and voltage signals are all collected with a data acquisition card. The collecting frequency is 1000 Hz. In order to prevent the impurities blocking the main valve and pilot valve orifice, a filter (3) is set in the system 4.1 Opening Pressure After adjusting the opening of the throttle valve (162) to maximum, the supply pressure reaches to the minimum; keep the pilot valve (16-8) in the electric state. The pilot valve is opened at this time. Slowly adjust the throttle valve so as to gradually increase the pressure. The change of pressure and flow can be observed. When the liquid flows out from the discharging valve, the opening pressure is minimum. When the valve opens, the pressure decreases. Therefore, it can be judged according to the record of the point of inflexion of pressure curve. The entrance pressure change of filling the valve is shown in Fig. 17. When the pressure is between 0.22 MPa and 0.23 MPa, a significant pressure drop appears, which means the valve is opening. Therefore, the opening pressure of filling the valve is approximately 0.22 MPa.

Fig. 17. Opening pressure

4.2 Response Characteristics The inlet pressure is adjusted to 0.8, 1.2 and 1.5 MPa. The pressure change is studied on the process of the opening and closing of the spool under different pressures of the liquid supply, which are response characteristics. The opening and closing process are shown in two images. Fig. 18 shows that the opening (left) and closing (right) process under the different inlet pressures. 672

In the process of opening, the solenoid pilot spool opens immediately after the solenoid pilot valve is electrified. The inlet position of the main spool produces small pressure fluctuations along with sudden change of the solenoid valve, and the fluctuation is the stability process of the pilot spool. With the main spool opening, the pressure declines rapidly. When spool attains the maximum opening, an inflection point happens, after which the pressure gently declines. The process of the gentle decline is a new process of pressure balance that the pump reestablishes. Therefore, the opening process isfrom the electrified point to the pressure inflection point in Fig. 18a, b and c. In the closing process, with the pilot valve closed, a pressure pulse appears first, and then the pressure rises after a straight section. A rapid decline and fluctuation appear after the high-point, and then the pressure gently rises. After the spool is completely shut down, it cannot be immediately restored to the pressure before opening. The process of the gentle rise after the pressure decline is a new process of pressure balance that the pump re-establishes. The fluctuation of the high-point is the impact and liquid returning as a result of the spool being completely closed. Thus, the closing process is from the electrified point to the highest pressure point in Fig. 18a, b and c. The opening process is between 0.3 and 0.4 s. The bigger the inlet pressure, the shorter the opening time; however, the changing amplitude is small, which is basically identical with Fig. 15. The time of closing process is around 1 s, which contains two stages, straight and rise, and the trend and time are essentially identical with the AMESim simulation results in Fig. 15. In Fig. 18, when the spool opens, it maintains in long period. Before closing, the pressure is under a stable condition, which is the low pressure flow characteristic. In different supply pressures of 0.8, 1.2 and 1.5 MPa, the stable flows are, respectively, 140, 160 and 190 L/min. Before closing, the pressure is maintained approximately 0.1 MPa. When the flow is 190 L/min, the pressure loss is minimum, which is about 0.07 MPa. The value is close to that of pressure drop of 240 L/min in the simulation process. When the flow is low, the fluid power is also small. In order to overcome the spring force, a larger pressure difference is needed. There will be a larger pressure drop, but the pressure is in the lower state. The opening pressure is 0.22 MPa. When flow is 190 L/min, the stable pressure is 0.07 MPa. Meanwhile, the opening time is about 0.3 to 0.4 s.

Liu, L. – Zhang, D. – Zhao, J.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

Fig. 18. Opening (left) and closing (right) characteristics at different inlet pressures; a) 0.8 MPa; b) 1.5 MPa; and c) 1.2 MPa

The closing time is about 1 s. The circulating liquid entrance response time is about 1.1 s. The test results show that the designed solenoid valve possesses better low-pressure characteristics and rapid response characteristics. 5 CONCLUSIONS Based on the logic relations of the filling valve, circulating valve and discharging valve, we designed the pilot-operated solenoid valve group. The radial seal of main valve is adopted co-axial seal (Gelai ring), and the end face seal is adopted a plane soft seal, in order to adapt to the characteristics of the water medium. Two-stage throttling of plane and cone structure can reduce the impact during the process of opening and closing. The simulation model of filling the valve is established via AMESim. The influence of

liquid damping on static and dynamic characteristics for the control valve is studied. The simulation results show that the response time is decided by the diameter of throttle nozzle and the spring stiffness (the bigger the diameter of the throttle nozzle and the spring rigidity, the faster the response), and with little influence of the supply pressure. The opening pressure and stable working pressure of the single valve are both small, which satisfies the demand of low pressure and high flow. 6 ACKNOWLEDGMENTS This work is supported by Specialized Research Fund for the Doctoral Program of Higher Education (20130095110012).

Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve

673

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 665-674

7 REFERENCES [1] Scheffcls (1996). Developments in water hydraulics. Hydraulics & Pneumatics, no. 12, p. 33-34. [2] Koskinen, K.T., Vilenius, M.J. (1998). Water hydraulics-A versatile technology. Journal of the Japan Hydraulics and Pneumatics Society, vol. 29, no. 7, p. 597-603. [3] Yang, H.Y., Gong, Y.J., Zhou, H. (2004). Development review of water hydraulic valve. China Mechanical Engineering, vol.15, no. 15, p. 1400-1404. (in Chinese) [4] Urata, E. (1999). Technological aspects of the new water hydraulic. The 6th Scandinavian International Conference on Fluid Power, Tampere, p. 21-34. [5] Backe, W. (1999). Water-or-oil hydraulics in the future. Proceedings of 6th Scandinavian International Conference on Fluid Power, Tampere, p. 51- 64. [6] Tang, G.Q., Tao, J., Wang, J., Zhu, Y.Q. (2008). Development of a pure water pilot-operated on-off solenoid hydraulic valve. Chinese Hydraulics & Pneumatics, no. 7, p. 68-70. (in Chinese) [7] Song, W.G. (2007). Control-valve liquid type hydraulic coupling based on the different needs. Coal Engineering, no. 9, p. 107-110. (in Chinese) [8] Lauhoff, H. (2005). Speed control on belt conveyor does it really save energy. Bulk Solids Handling, vol. 25, no. 6, p. 368-377. [9] Majdič, F., Pezdirnik, J., Kalin M. (2011). Experimental validation of the lifetime performance of a proportional 4/3 hydraulic valve operating in water.

674

Tribology International, vol. 44, no. 12, p. 2013-2021, DOI:10.1016/j.triboint.2011.08.020. [10] Majdič, F., Velkavrh, I., Kalin, M. (2013). Improving the performance of a proportional 4/3 water hydraulic valve by using a diamond-like-carbon coating. Wear, vol. 297, no. 1-2, p. 1016-1024, DOI:10.1016/j. wear.2012.11.060. [11] Majdič, F., Kalin, M. (2014). Test rig and comparison of pressure changes at transient phenomena in waterand oil-based power-control hydraulics. Journal of Vibroengineering, vol. 16, no. 1, p. 401-411. [12] Šimic, M., Debevec, M., Herakovič, N. (2014). Modelling of hydraulic spool-valves with special designed metering edges. Strojniški vestnik - Journal of Mechanical Engineering, vol. 60, no. 2, p. 77-83, DOI:10.5545/sv-jme.2013.1104. [13] Tič, V., Lovrec, D. (2012). Design of Modern Hydraulic Tank Using Fluid Flow Simulation. International Journal of Simulation Modelling, vol. 11, no. 2, p. 7788, DOI:10.2507/IJSIMM11(2)2.202. [14] Jošt, D., Škerlavaj, A., Lipej, A. (2014). Improvement of efficiency prediction for a Kaplan turbine with advanced turbulence models. Strojniški vestnik Journal of Mechanical Engineering, vol. 60, no. 2, p. 124-134, DOI:10.5545/sv-jme.2013.1222. [15] Wang, C.M., Li, C.X., Yin, X. (2000). Experimental research on the hydraulic performance of jet pump. Pump Technology, no. 5, p. 3-7, p. 26. (in Chinese) [16] Zhang, D.S. (2011). Research on the Design Theory and Key Technologies of Large Power Valve-Control Hydrodynamic Coupling. China University of Mining & Technology, Xuzhou. (in Chinese)

Liu, L. – Zhang, D. – Zhao, J.

Received for review: 2013-02-25 Received revised form: 2013-06-13 Accepted for publication: 2013-08-20

Optimization of Machining Performance in High-Pressure Assisted Turning of Ti6Al4V Alloy Çolak, O. Oğuz Çolak* Süleyman Demirel University, CAD/CAM Research and Application Center, Turkey In this study, a genetic algorithm has been employed to determine optimum cutting parameters in the turning of Ti6Al4V alloy under conventional and high pressure cooling conditions. Three machining performance measures, i.e. surface roughness, material removal rate and cutting power, are considered as optimization criteria. First, with multi-regression analysis of experimental responses, empirical equations are defined and, by using these equations, objective functions are constructed for each pressure level, based on a hybrid model. Objective functions are maximized by means of a genetic algorithm and optimum machining parameters are determined. Moreover, tool wear tests are carried out at a cutting condition that is close to the optimum machining parameters. Optimization results show that optimum cutting parameters and their responses, particularly in P = 6 and 150 bar cooling conditions, are quite similar, but tool life is significantly different. Maximum tool life is achieved in the highest pressure level (P = 300 bar). Keywords: high pressure cooling, optimization, tool life

0 INTRODUCTION Titanium alloys have found wide application in areas such as chemical processing equipment, surgical implants, and prosthetic devices, due to their excellent corrosion resistance, as well as in the automotive industry in engine components such as valves, connecting rods, drive shafts, crankshafts, and suspension assemblies, owing to their unique characteristics, including low density or high strengthto-weight ratio (density of titanium is about 60% of that of steel or nickel-based super alloys) [1] and [2]. They are considered to be difficult-to-machine materials due to the inherent material properties of high chemical reactivity and low thermal conductivity [3]. The major problems during machining are high temperatures and stresses close to the tool nose resulting in rapid tool wear. This is partly due to the poor thermal conductivity of titanium alloys, which implies that a considerable proportion of heat generated (about 80%) during the machining process is conducted into the cutting tool [2] and [4]. Recently, various cooling-lubrication techniques have been developed to improve the machinability of titanium, nickel alloys and some other materials [5]. High-pressure jet-assisted cooling (HPJAC) is one of the main methods that aims to increase machining performance by using the thermal and mechanical properties of high-pressure jet water or emulsion injected into the cutting zone [6]. The application of a high-pressure water jet to the tool-chip interface during the machining of titanium, nickel alloys, in particular, which have superb properties has significant benefits in terms of machining performance, such

as providing control of the chip shapes, better chip breakability, improved chip removal, considerably reduced temperatures in the cutting zone, resulting in prolonged tool life (5 to 15 times) [7] and [8]. It can also improve the surface integrity of workpieces [9]. Additionally, HPJAC enhances production efficiency compared to conventional cooling by increasing the cutting speed [10]. Çolak [6] studied the machinability of Inconel 718 under conventional and high pressure cooling conditions at various cutting speeds, feed rates, depths of cut, and pressure levels with a (Ti,Al) N+TiN-coated carbide cutting tool. He found that the injection of high-pressure coolant to the tool-chip interface reduces cutting force components, provides desirable chip breakability and lower cutting tool wear, especially flank face wear, due to the efficient lubrication and cooling than conventional cooling. Palanisamy et al. [7] conducted an investigation on HPJAC in the turning of Ti6Al4V alloy. A series of experiments were carried out at various cutting parameters and pressure levels with uncoated straight tungsten carbide inserts. The investigation showed that the application of high-pressure directly on the tool-chip interface provides smaller chips due to the mechanical effect of high pressure, which generates a more efficient chip evacuation process. They also found that HPJAC increases tool life by almost three times in comparison to conventional cooling. In many manufacturing industries, parameter setting is made based on the skill of the operator or on handbook recommendations. Consequently, optimum parameter setting is not achieved, which leads towards reduced production, poor quality and

*Corr. Author’s Address: Süleyman Demirel University, CAD/CAM Research and Application Center, 32260, Isparta, Turkey, kadirkiran@sdu.edu.tr

675

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 675-681

increased product costs [11]. Therefore, optimization of machining performance has great importance in terms of machining quality and cost for manufacturing industries. Many studies in the literature have been performed on the optimization of machining operations and parameters by using various methods and models. Cus and Balic [12] conducted an investigation on cutting parameter optimization in milling by means of a genetic algorithm (GA) approach. The results of their proposed approach were compared with results of other approaches. They concluded that the GA approach can be integrated online with an intelligent manufacturing system for automated process planning; it can also use parameter selection of complex machined parts, which results in decreasing production cost, time and improving product quality. Raja and Baskar [13] studied machining parameters optimization to obtain the desired surface roughness in face milling by using a particle swarm optimization technique. The proposed approach reduces the time and cost of the trial for surface roughness prediction. It can also be utilized in industrial applications due to its predicting ability and accuracy. Da et al. [14] developed a hybrid model to optimize cutting parameters in turning by taking into account machining-performance factors, such as cutting force (Fc), tool-life (T), surface roughness (Ra), material removal rate (MR) and chip breakability (CB). Hagiwara et al. [15] also performed the same approach for the contour finish turning operations. Surface roughness and chip breakability were chosen as optimization criteria due to their significance to finish turning. They stated that optimum cutting parameters, determined via the GA approach, yielded better chip breakability and surface quality. Several investigations that used the same approach can be found in the literature [16] to [18]. Furthermore, the present approach has been utilized in this research to determine optimum cutting parameters during the turning of Ti6Al4V under HPJAC and conventional cooling conditions based on three machining performances, i.e. the material removal rate (MR), cutting power (Pc) and surface roughness (Ra). Therefore, a number of machining tests with Ti6Al4V were carried out in conventional and various high pressure levels of cooling conditions. The experiments were designed based on a Taguchi L9 orthogonal array [23] at three different cutting speeds (Vc), feed rates (f ) and pressure (P) levels. Cutting forces (Fc, Fp, Ff) and surface roughness (Ra) were recorded during the experiments at a constant depth of cut (ap); moreover, the material removal rate 676

(MR) and cutting power (Pc) were calculated according to cutting parameters and experimental responses. In order to construct objective functions for each pressure level, empirical equations that indicate the relation between cutting conditions and experimental responses were obtained via multi-regression analysis. Objective functions were maximized by means of GA, and optimum machining parameters were determined. Finally, tool wear tests were carried out at a cutting condition that is close to the optimum machining parameters. 1 EXPERIMENTAL PROCEDURE The experiments were designed according to plan of a Taguchi L9 orthogonal array at three different cutting speeds, feed rates and pressure levels, and depths of cut were kept constant during the tests. Each experiment was performed with new cutting edge to compare results. Cutting parameters and their levels are given in Table 1. Table 1. The levels of cutting parameters Level

Vc [m/min] f [mm/rev] P [bar] ap [mm]

III

0.1

0.15

0.2

Conv. (6) 2

150 -

300 -

The experiments were conducted on an ALEX ANL-75 CNC lathe machine that has a variable spindle speed (50 to 4000 rpm) and a 15 kW motor drive that is equipped with the high-pressure plunger pump of maximum 350 bar pressure and 21 l/min volumetric flow rate capacity (Fig. 1). The cutting fluid used in experiments was a chemical-based 6 to 7% concentration water soluble oil (Swisslube Blaser BCool 650). The high pressure cutting fluid was injected between the cutting tool-chip interface at a low angle (about 5 to 6° with the cutting tool rake angle), as shown in Fig. 1. A CNMG0812 (Ti,Al)N+TiN-coated carbide cutting tool has been chosen for the experiments. The tool has rε = 0.8 mm nose radius. It was mounted on a SECO Jet stream PCLNR tool holder, which resulted in: cutting rake angle, γa = –6°, back rake angle, γb = –6°, approach angle, Kr = 95°, and d = 0.8 mm nozzle diameter. All the experiments were carried out using the Ti6Al4V alloy supplied as bars (80 mm diameter and 200 mm long) with a hardness of 292 do 407 HV100. The standard chemical composition is 0.08% C, 5.5 to 6.75% Al, 3.5 to 4.5% V, 0.03% Çolak, O.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 675-681

Fig. 1. Photographic view of experimental set-up and illustration of the high-pressure injection system

N, 0.14 to 0.23% O, 0.01% H, 0.3% Fe, 50 ppm %Y, balance Ti. The mechanical properties of Ti6Al4V are tensile strength: 900 to 1160 MPa, yield strength: 830 MPa, elongation: 8% [2]. 2 REGRESSION ANALYSIS In order to construct a hybrid model [14] for a singlepass straight-turning operations based on the multiple machining prformance measures, empirical equations have been obtained by multi-regression analysis of experimental responses shown in Table 2. Fc, Fp, Ff and Ra were measured during the machining. The others MR, Pc, were calculated based on cutting parameters and main cutting forces (Fc), respectively and inserted as responses. The generated equations are presented in Eq. (1). with R2 = 0.99, 1 and 0.99, respectively:

Ra = –0.42 –0.002 Vc + 10.31 f + 0.0004 P ,

MR = 6.5·10–16 Vc + 1.6·10–13 f + 2 Vc f , (1)

Pc = –0.052 + 3.1·10–3 Vc + 0.31 f + 5·10–5 P + 0.043 Vc f .

3 OPTIMIZATION CRITERION A hybrid model developed by Da et al. [14] was utilized for single-pass straight turning optimization problems. The parameters, Ra, Pc, and MR denote the surface roughness, cutting power and material removal rate, respectively. Corresponding constraints on these machining performance measures are assumed as, Ra’, Pc’ and MR’. The objective function can be constructed as seen in Eq. (2).  R' − R U (Vc , f ) = CR  a ' a  Ra

  Pc' − Pc  + CP  '   Pc  M − M'  +CM  R ' R  ,  MR 

 +  (2)

where each term is normalized by using user-provided information concerning machining performance requirements. Ci(i = R, P and M) are weighting factors considered as the contribution coefficient of ith machining performance variable to the value of the operation. These weighting factors satisfy two conditions:

Table 2. The experiment results No 1 2 3 4 5 6 7 8 9

Vc [m/min] f 50 50 50 70 70 70 90 90 90

[mm/rev] 0.1 0.15 0.2 0.1 0.15 0.2 0.1 0.15 0.2

P [bar]

Fc [N]

Fp [N]

Ff [N]

Ra [µm]

MR [cm3/min]

Pc [kW]

6 150 300 150 300 6 300 6 150

424.1 583.1 740.8 436.5 605.8 725.7 445.3 580.8 720.5

95.3 102.3 121.5 81.9 111.1 120.5 91.8 134.9 122

95.9 108.3 119.7 82.7 110.8 121.6 92.6 135.8 121.5

0.45 1.16 1.66 0.54 1.06 1.51 0.58 0.96 1.49

10 15 20 14 21 28 18 27 36

0.35 0.49 0.62 0.51 0.71 0.85 0.67 0.87 1.08

Optimization of Machining Performance in High-Pressure Assisted Turning of Ti6Al4V Alloy

677

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 675-681

(P = 6 bar) and high pressure cooling (P = 150 and P = 300 bar) conditions, respectively. Feasible regions in HPJAC conditions are smaller than that in conventional cooling, as seen in Figs. 2 to 4, which is due to the increasing effect of HPJAC on surface roughness. Surface quality is reduced with rising pressure levels, which also could be shown in optimization results for each pressure level given in Tables 5 to 7. This observation confirms the experiments of Courbon et al. [10], who stated that this may result from the fact that the chips damage the machined surface with high pressure when the velocity of the chip is relatively low. Similar observations have been also reported in [1]. In contrast, there is no significant difference in the optimization results, Pc, which were calculated based on the main cutting force for each pressure level. It can be assumed that HPJAC has not considerably affected the main cutting forces in this set of experiments. Optimum cutting parameters are also quite similar, especially at P = 6 and 150 bar cooling conditions; the responses (Pc, MR), are almost same. However, as discussed in the next section, tool life is remarkably improved by HPJAC.

CR + CP + CM = 1, 0 ≤ Ci ≤ 1 (i = R, P, M ). (3)

Corresponding constraints on these machining performance measures are assumed as Ra′, Pc′ and MR′. Therefore, the constraint conditions are:

Ra ≤ Ra' , Pc ≤ Pc' , M R ≥ M R' , Vc min ≤ Vc ≤ Vc max ,

f min ≤ f ≤ f max . (4)

Hence the optimization problem becomes: U (Vc , f ) , Maximize With respect to Vc , f , Subject to Ra ≤ Ra' , Pc ≤ Pc' , M R ≥ M R' , Vc min ≤ Vc ≤ Vc max , f min ≤ f ≤ f max .

The material removal rate and cutting power have priority in the rough turning, because the main purpose is to remove the maximum material per unit of time in the rough turning. Thus, the weighting factors CM and CP are set equal to 0.45 and weighting factor for surface roughness, CR, is made equal to 0.1 as the surface roughness is not given precedence in the rough turning. The weighting factors and constraints are shown in Table 3.The parameters used GA are also given in Table 4.

4.2 Tool Wear Test Results After the optimization of machining performance, tool wear tests were conducted at constant cutting parameters, which are close to the optimum point for each pressure level, Vc = 70 m/min, f = 0.15 mm/rev, ap = 2 mm, under conventional and highpressure cooling conditions. Tool wear limits were considered to be average tool flank wear, VBB = 0.3 mm, maximum tool flank wear, VBBmax = 0.6 mm, and notch wear, VBN = 1 mm, according to ISO 3685:1993 standards [24]. Tool life for conventional and high-pressure cooling conditions are shown in Fig. 5. It can be clearly seen that high pressure cooling conditions have an increasing effect on tool life compared to conventional cooling, which is in agreement with the experiments performed by Nandy et al. [1], Ezugwu et al. [2] and Palanisamy et al. [7]. Hong et al. [19] stated that titanium and its alloys are poor thermal conductors. As a result, the heat generated when machining titanium cannot dissipate quickly; rather, most of the heat is concentrated on the cutting edge and tool face, which causes rapid tool wear. The injection of high pressure coolant to the tool-chip interface provides efficient lubrication and cooling by penetrating the cutting zone, which results in a considerable reduction in the temperature. Hence,

Table 3. Weighting factors and constraints for optimization Constraints

Rough turning

CM CP CR

0.45

[m/min]

50 to 90

[mm/rev]

0.1 to 0.2

Weighting factors

Vc f ap MR′ Pc′ Ra′

0.45 0.1

[mm]

[cm3/min]

[kW]

0.7 1

Table 4. Parameters used in GA Population size Crossover fraction Max number of generations

300 0.75 2000

4 RESULTS AND DISCUSSIONS 4.1 Optimization Results Figs. 2 to 4 show the contour plots with feasible regions and optimum points for conventional cooling 678

Çolak, O.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 675-681

Fig. 2. Illustration of feasible region and optimum point for conventional cooling (P = 6 bar)

Fig. 3. Illustration of feasible region and optimum point for high pressure cooling (P = 150 bar)

Fig. 4. Illustration of feasible region and optimum point for high pressure cooling (P = 300 bar)

cutting tool wear caused by high temperatures could be reduced or entirely prevented, leading to extended

tool life in comparison to conventional cooling. This conclusion has also been stated in [20] to [22].

Optimization of Machining Performance in High-Pressure Assisted Turning of Ti6Al4V Alloy

679

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 675-681

Fig. 5. Tool life for conventional and high pressure cooling conditions at Vc = 70 m/min, f = 0.15 mm/rev, ap=2 mm Table 5. Optimization results for conventional cooling (P = 6 bar)

[m/min] 71

[mm/rev] 0.141

[mm] 2

[µm] 0.89

[kW] 0.64

[cm3/min] 20

Table 6. Optimization results for high pressure cooling (P=150 bar)

[m/min] 71.3

[mm/rev] 0.14

[mm] 2

[µm] 0.94

[kW] 0.65

[cm3/min] 20

Table 7. Optimization results for high pressure cooling (P=300 bar)

[m/min] 75

[mm/rev] 0.133

[mm] 2

[µm] 0.92

[kW] 0.66

[cm3/min] 20

5 CONCLUSIONS In this study, machining performance optimization was experimentally investigated in the turning of Ti6Al4V under HPJAC and conventional cooling conditions. Therefore, a number of machining tests with Ti6Al4V were carried out in conventional and various high-pressure levels of cooling conditions. By means of multi-regression analysis, empirical equations were obtained, and objective functions for optimization were constructed using these equations based on the hybrid model. Furthermore, tool wear tests were carried out for each cooling condition at constant machining parameters. The following conclusions can be drawn from this work: 1. The hybrid model used in this research for the optimization of machining performance is very useful for determining optimum cutting 680

parameters according to the given optimization criteria. Feasible regions in HPJAC conditions are smaller than that in conventional cooling, because of the increasing effect of HPJAC on surface roughness. There are no significant change optimization results, Pc, in all the cooling conditions. This is due to the fact that high pressure cooling has not considerably influenced the main cutting forces in this set of experiment. Although the optimum cutting parameters and its responses are quite similar, tool life is remarkably different in each cooling condition. Tool life is about 47% and 112% higher than conventional cooling in P = 150 and 300 bar, respectively. The application of high pressure cooling during the machining of hard-to-cut materials supports sustainability in manufacturing by increasing tool life, thus resulting in lower machining cost. 6 ACKNOWLEDGEMENTS

This study was supported by the Scientific and Technological Research Council of Turkey (TÜBİTAK) and the Slovenian Research Agency (ARRS). The authors also would like to thank the companies SECOTOOLS, BLASER SwissLube and TAI-TUSAŞ A.Ş. for their support to this study. 7 REFERENCES [1] Nandy, A.K., Gowrishankar, M.C., Paul, S. (2009). Some studies on high-pressure cooling in turning of Ti–6Al–4V. International Journal of Machine Tools & Manufacture, vol. 49, no. 2, p. 182-198, DOI:10.1016/j. ijmachtools.2008.08.008. Çolak, O.

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, 675-681

[2] Ezugwu, E.O., Bonney, J., Da Silva, R.B., Machado A.R., Ugwoha, E. (2009). High productivity rough turning of Ti-6Al-4V alloy, with flood and high-pressure cooling. Society of Tribologists and Lubrication Engineers, Tribology Transactions, vol. 52, no. 3 p. 95400, DOI:10.1080/10402000802687866. [3] Machai, C., Biermann, D. (2011). Machining of β-titanium-alloy Ti–10V–2Fe–3Al under cryogenic conditions: Cooling with carbon dioxide snow. Journal of Materials Processing Technology, vol. 211, no. 6, p. 1175-1183, DOI:10.1016/j.jmatprotec.2011.01.022. [4] Hong, S.Y., Ding, Y. (2001). Cooling approaches and cutting temperatures in cryogenic machining of Ti-6Al-4V. International Journal of Machine Tools & Manufacture, vol. 41, no. 10, p. 1417-1437, DOI:10.1016/S0890-6955(01)00026-8. [5] Sharma, V.S., Dogra, M., Suri, N.M. (2009). Cooling techniques for improved productivity in turning. International Journal of Machine Tools & Manufacture, vol. 49, no. 6, p. 435-453, DOI:10.1016/j. ijmachtools.2008.12.010. [6] Çolak, O. (2012). Investigation on Machining Performance of Inconel 718 under High Pressure Cooling Conditions. Strojniški vestnik - Journal of Mechanical Engineering, vol. 58, no. 11, p. 683-690, DOI:10.5545/sv-jme.2012.730. [7] Palanisamy, S., McDonald, S.D., Dargusch, M.S. (2009). Effects of coolant pressure on chip formation while turning Ti6Al4V alloy. International Journal of Machine Tools & Manufacture, vol. 49, no. 9, p. 739743, DOI:10.1016/j.ijmachtools.2009.02.010. [8] Ezugwu, E.O., Bonney, J. (2004). Effect of highpressure coolant supply when machining nickel-base, Inconel 718, alloy with coated carbide tools. Journal of Materials Processing Technology, vol. 153-154, p. 1045-1050, DOI:10.1016/j.jmatprotec.2004.04.329. [9] Ezugwu, E.O., Bonney, J., Silva, R.B.D., Çakir, O. (2007). Surface integrity of finished turned Ti–6Al– 4V alloy with PCD tools using conventional and high pressure coolant supplies. International Journal of Machine Tools & Manufacture, vol. 47, no. 6, p. 884891, DOI:10.1016/j.ijmachtools.2006.08.005. [10] Courbon, C., Kramar, D., Krajnik, P., Pusavec, F., Rech, J., Kopac, J. (2009). Investigation of machining performance in high-pressure jet assisted turning of Inconel 718: an experimental study. International Journal of Machine Tools & Manufacture, vol. 49, no. 14, p. 1114-1125, DOI:10.1016/j. ijmachtools.2009.07.010. [11] Venkata, R.R., Kalyankar, V.D. (2012). Parameter optimization of modern machining processes using teaching–learning-based optimization algorithm. Engineering Applications of Artificial Intelligence, vol. 26, no. 1, p. 524-531, DOI:10.1016/j. engappai.2012.06.007. [12] Cus, F., Balic, J. (2003). Optimization of cutting process by GA approach. Robotics and Computer

Integrated Manufacturing, vol. 19, no. 1-2, p. 113-121, DOI:10.1016/S0736-5845(02)00068-6. [13] Raja, S.B, Baskar, N. (2012). Application of Particle Swarm Optimization technique for achieving desired milled surface roughness in minimum machining time. Expert Systems with Applications, vol. 39, no. 5, p. 5982-5989, DOI:10.1016/j.eswa.2011.11.110. [14] Da, Z.J., Sabler, J.P., Jawahir, I.S. (1996). Multiple criteria optimization of finish turning operations based on a hybrid model. Proceedings of the ASME Design Engineering Technical Conferences and Computers in Engineering Conference, Irvine. [15] Hagiwara, M., Chen, S., Jawahir, I.S. (2009). Contour finish turning operations with coated grooved tools: Optimization of machining performance. Journal of Materials Processing Technology, vol. 209, no. 1, p. 332-342, DOI:10.1016/j.jmatprotec.2008.02.023. [16] Jawahir, I.S., Wang, X. (2007). Development of hybrid predictive models and optimization techniques for machining operations. Journal of Materials Processing Technology, vol. 185, no. 1-3, p. 46-59, DOI:10.1016/j. jmatprotec.2006.03.133. [17] Wang, X., Da, Z.J., Balaji, A.K., Jawahir, I.S. (2007). Performance-based predictive models and optimization methods for turning operations and applications: Part 3—optimum cutting conditions and selection of cutting tools. Journal of Manufacturing Processes, vol. 9, no. 1 p. 61-74, DOI:10.1016/S1526-6125(07)70108-1. [18] Kardekar, A.D. (2005). Modeling and Optimization of Machining Performance Measures in Face Milling of Automotive Aluminum Alloy A380 under Different Lubrication/Cooling Conditions for Sustainable Manufacturing, M.Sc. Thesis, University of Kentucky, Lexington. [19] Hong, S.Y., Markus, I., Jeong, W. (2001). New cooling approach and tool life improvement in cryogenic machining of titanium alloy Ti-6Al-4V. International Journal of Machine Tools & Manufacture, vol. 41, no. 15. p. 2245-2260, DOI:10.1016/S0890-6955(01)00041-4. [20] Kaminski, J., Alvelid, B. (2000). Temperature reduction in the cutting zone in water-jet assisted turning. Journal of Materials Processing Technology, vol. 106, no. 1-3, p. 68-73, DOI:10.1016/S0924-0136(00)00640-3. [21] Dahlman, P., Escursell, M. (2004). High-pressure jetassisted cooling: a new possibility for near net shape turning of decarburized steel. International Journal of Machine Tools & Manufacture, vol. 44, no. 1, p. 109115, DOI:10.1016/S0890-6955(03)00058-0. [22] Kramar, D., Krajnik, P., Kopac, J. (2010). Capability of high pressure cooling in the turning of surface hardened piston rods. Journal of Materials Processing Technology, vol. 210, no. 2, p. 212-218, DOI:10.1016/j. jmatprotec.2009.09.002. [23] Yang, K., El-Haik, B. (2003). Design for Six Sigma. The McGraw-Hill, New York. [24] ISO 3685:1993 (1993). Tool-life testing with singlepoint turning tools. International Organization for Standardization, Geneva.

Optimization of Machining Performance in High-Pressure Assisted Turning of Ti6Al4V Alloy

681

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10 Vsebina

Vsebina Strojniški vestnik - Journal of Mechanical Engineering letnik 60, (2014), številka 10 Ljubljana, oktober 2014 ISSN 0039-2480 Izhaja mesečno

Razširjeni povzetki Uroš Stritih, Andrej Bombač: Opis in analiza adsorpcijskega shranjevalnika toplote Rita Ambu, Andrea Manuello Bertetto, Costantino Falchi: Zasnova prototipnega sistema za lunarno okolje Xin Jin, Hua Liu, Wenbin Ju: Analiza seizmične obremenitve vetrne turbine na osnovi numeričnega izračuna Srečko Glodež, Marko Šori, Tomaž Verlak: Računski model za analizo upogibne trdnosti sintranih zobnikov Periyakgounder Suresh, Rajamanickam Venkatesan, Tamilperruvalathan Sekar, Natarajan Elango, Varatharajan Sathiyamoorthy: Optimizacija vmesnih spremenljivk pri mikroelektroerozijski obdelavi nerjavnega jekla 316L s pomočjo genetskega algoritma in metodologije odzivne površine Liu Lei, Zhang Desheng, Zhao Jiyun: Snovanje in raziskava vodnega nizkotlačnega pilotnega elektromagnetnega ventila za velike pretoke Oğuz Çolak: Optimizacija struženja zlitine Ti6Al4V s podporo visokotlačne hladilne tekočine Osebne vesti Doktorske disertacije, diplomske naloge

SI 117 SI 118 SI 119 SI 120

SI 121 SI 122 SI 123

SI 124

Prejeto v recenzijo: 2014-03-26 Prejeto popravljeno: 2014-05-26 Odobreno za objavo: 2014-06-24

Opis in analiza adsorpcijskega shranjevalnika toplote Stritih, U. – Bombač, A. Uroš Stritih* – Andrej Bombač

Univerza v Ljubljani, Fakulteta za strojništvo, Slovenija

Namen raziskave je narediti in eksperimentalno ovrednotiti poskusni model adsorpcijskega shranjevanilnika toplote, s katerim smo pokazali način shranjevanja toplote sprejemnikov sončne energije. Kot adsorbent smo uporabili aluminosilikat Na2O Al2O3 * 2SiO2, ki je proizvod slovenskega podjetja. Pri poskusnem modelu smo analizirali vpliv parametrov na uparjanje in kondenzacijo vodne pare kot so: temperatura adsorbenta, temperatura na vstopu in izstopu iz prenosnika toplote (ki se nahaja v samem shranjevalniku), pretok vode skozi prenosnik toplote in razmere v shranjevalniku toplote. S pomočjo izmerjenih parametrov smo adsorpcijski shranjevalnik toplote ovrednotili glede na latentni in senzibilni delež toplote ter analizirali zadostnost shranjene toplote v ogrevalni sezoni. Pri izdelavi načrta smo se zgledovali po evropskem projekta MODESTORE (Modular High Energy Density Sorption Heat Storage) kjer sta adsorber in uparjalnik / kondenzator v eni posodi. V adsorberju (zgornji del posode) je shranjen adsorbent. Skozi vertikalen kanal v sredini posode dovajamo paro. Adsorbent je obdan s perforirano bakreno pločevino na katerem je integriran cevni prenosnik toplote za odvod in dovod toplote. V uparjalniku / kondenzatorju (spodnji del posode) je grelnik, ki služi kot nizko temperaturni vir energije. Na dnu posode je dovodna / odvodna cev za kondenzat in ventil. Pri preizkusu smo uporabili 5700 g adsorbenta. Delovni tlak pri adsorpciji je bil med 20 in 60 mbar absolutnega tlaka. V fazi adsorpcije toplotni tok narašča sorazmerno z naraščanjem temperature vode, ki izstopa iz prenosnika toplote. Največji toplotni tok 123,2 W dosežemo pri največji temperaturni razliki vstopne in izstopne vode iz prenosnika toplote. Povprečni toplotni tok znaša 12,8 W. Kot rezultat prenesene toplote iz shranjevalnika toplote na maso adsorbenta dobimo specifično energijo 48,86 kWh/kg. Rezultat desorpcije pokaže, da je najvišja temperatura adsorbenta znašala 85 °C , iz adsorbenta pa se je izločilo 4758 g vode. Toplotni shranjevalnik ima to dobro lastnost, da shranjena toplota lahko ostane v shranjevalniku poljubno dogo - dokler sta adsorbent in adsorbat ločena - in je zato zelo primeren kot sezonski shranjevalnik toplote, kar je pomembna prednost pred senzibilnimi in latentnimi shranjevalniki. Keywords: shranjevalnik toplote, termo-kemično shranjevanje, adsorbat, adsorbent, aluminosilikat, eksperiment

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

SI 117

Prejeto v recenzijo: 2014-01-30 Prejeto popravljeno: 2014-05-14 Odobreno za objavo: 2014-06-16

Zasnova prototipnega sistema za lunarno okolje Rita Ambu, R. – Manuello Bertetto A. – Falchi, C. Rita Ambu* – Andrea Manuello Bertetto – Costantino Falchi Univerza v Cagliariju, Oddelek za strojništvo, kemijsko tehnologijo in materiale, Italija

Veliko zanimanje za raziskovanje Lune je v preteklih letih privedlo do več lunarnih misij. Informacije, zbrane na teh misijah, govorijo v prid priložnostim za vzpostavitev človeške prisotnosti na našem naravnem satelitu in izkoriščanje morebitnih koristi. Za postavitev baze na Luni pa bi bila potrebna obsežna dela s pripravo lokacije, izkopavanjem in premikanjem materiala. Delovne operacije strojev za izvajanje teh delovnih operacij bi bile sicer značilne za gradbeno mehanizacijo, za učinkovito delo na Luninem površju pa bi bile potrebne drugačne konstrukcijske rešitve strojev. Članek obravnava raziskave vozil za delovne misije na Luninem površju. Problem rokovanja s skalami med miniranjem in gradnjo na Luni je zelo kompleksen, članek pa skuša podati izviren prispevek na tem raziskovalnem področju. Konkretno predstavlja zasnovo osnovnega dela roverja, ki bi odstranjeval skale iz področij na Luninem površju, namenjenih objektom in infrastrukturi. Glavna naloga roverja je prijemanje in premikanje skal, zasnovan pa je kot prijemalni mehanizem na štirih stebrih, povezanih s pari koles. Članek obravnava komponente stebra in posebna pozornost je posvečena dvigalnemu mehanizmu. Sistem mobilnosti je zaokrožen s predlogom spiralnega vzmetnega kolesa. Študija vključuje analizo primerne kombinacije in izbire materialov komponent za konstrukcijo minimalne mase in velikosti za delo na Luninem površju. Geometrija komponent je bila določena s parametričnim CADmodeliranjem, konstrukcijske značilnosti najpomembnejših delov v pogojih izjemnih obremenitev pa so bile ovrednotene z numerično simulacijo po metodi končnih elementov. Analogen pristop je bil uporabljen tudi za opredelitev lastnosti in zmogljivosti kolesa. Končno je bil izdelan tudi prototip stebra in kolesa za preverjanje predlagane konfiguracije. Stebri roverja so zasnovani za izvajanje različnih osnovnih delovnih nalog. Omogočajo pogon vozila, dviganje bremen in nadzorovanje ravnotežja na razgibanem površju. Zasnova omogoča določitev takšne lege okvirja na roverju, ki ustreza velikosti in geometriji skale. Ker ima vsak steber svoj neodvisen pogon, lahko vozilo deluje tudi na zahtevnem terenu in z neokrnjenim ravnotežjem prečka ovire v velikosti koles. Tako se zmanjša tveganje prevračanja pri vožnji po terenu. Konstrukcija kolesa sledi dobrim praksam iz predhodnih vozil, specifičnim zahtevam predlaganega vozila in nedavni študiji nepnevmatskih koles, katerih konstrukcija izkorišča možnosti naprednih kompozitnih materialov. Predlagani pristop je obetaven za Lunarne aplikacije, saj je prilagojen posebnim značilnostim tega okolja. Konstrukcija vozila je enostavna in lahka, kar prinaša številne prednosti tako pri transportu kot pri delu na Lunini površini. Ker je vozilo zmožno premagovati ovire z vzdolžnim spreminjanjem položaja parov koles, lahko deluje na neravnem terenu. Numerične simulacije so pokazale, da so lahko glavni deli izpostavljeni tudi velikim obremenitvam. Nadaljnje delo bo usmerjeno v eksperimentalno verifikacijo prototipnega roverja v okolju, ki simulira značilnosti Luninega površja. Ključne besede: zasnova, delovanje, lunarno okolje, steber, kolo, simulacija po MKE

SI 118

*Naslov avtorja za dopisovanje: Univerza v Cagliariju, Oddelek za strojnišvo, kemijo in inženiring materialov, Via Marengo 2, 09123 Cagliari, Italija, ambu@unica.it

Prejeto v recenzijo: 2014-01-04 Prejeto popravljeno: 2014-04-01 Odobreno za objavo: 2014-05-07

Analiza seizmične obremenitve vetrne turbine na osnovi numeričnega izračuna Jin, X. – Liu, H. – Ju, W.B. Xin Jin1,* – Hua Liu2 – Wenbin Ju1

1 Univerza

v Chongqingu, Kolidž za strojništvo, Kitajska 2 Dongfang elektro podjetje, China

Vetrne turbine so zadnjih 10 let rasle predvsem po Severni Evropi, ki ni potresno ogrožena oz. kjer so potresi šibkejši, zato seizmičnim vplivom pri snovanju vetrnih turbin ni bila posvečena večja pozornost. V zadnjem času pa je bilo vse več vetrnih turbin postavljenih tudi na potresno ogroženih območjih, zato se je pojavila potreba po analizi dinamičnega odziva vetrnih turbin na potrese ter po vključitvi seizmičnih dejavnikov v konstrukcijski proces. Pri pregledu razpoložljive literature se je izkazalo, da so študije modeliranja sklopljenega sistema več teles lopatice-steber-fundament pri vetrnih turbinah redke, še posebej pa to velja za študije, ki bi uporabljale tudi metodo časovne zgodovine za analizo dinamičnih seizmičnih obremenitev na osnovi sklopljenega modela. Teorija dinamike sistema več teles je bila uporabljena za preučitev dinamičnega odziva vetrnih turbin ob potresu, pri čemer so bile vključene interakcije med tlemi in konstrukcijo. Postavljen je bil analitični model za preučitev seizmičnega vpliva na obremenitve vetrne turbine med obratovanjem, ki bo osnova za snovanje ključnih komponent in strategij krmiljenja vetrnih turbin na potresno ogroženih območjih. Lopatice in steber vetrne turbine po osnovni teoriji dinamike sistemov več teles tvorijo vrsto zveznih diskretnih enot, medtem ko je interakcije med stebrno konstrukcijo vetrne turbine in tlemi mogoče popisati kot sistem vzmeti in dušilk na stiku med fundamentom in maso tal. Za preučitev dinamičnih lastnosti vetrne turbine in analizo dinamičnega seizmičnega vpliva na vetrno turbino po standardu Eurocode 8 je bila uporabljena metoda časovne zgodovine. Analiza je pokazala, da nenaden potres med normalnim delovanjem vetrne turbine do določene mere zmanjša učinkovitost vetrne turbine. Učinkovitost proizvodnje električne energije ne zaniha v večji meri, če se ne poškoduje konstrukcija vetrne turbine. V primeru poškodb konstrukcije pa je intenziteta motenj odvisna od obremenitve, ki jo povzroči potresni sunek. Z izbiro treh značilnih obremenitvenih sestavov vetrne turbine je bilo mogoče pokazati, da obremenitev najbolj niha na podnožju stebra, ki je najbližje temelju. To pomeni, da so zahteve pri konstrukciji podnožja stebra večje. Seizmično vzbujanje se prenaša po stebru in nato doseže njegov vrh, nihanje obremenitev pa se do gredi že močno zmanjša. Nihanja obremenitev se še dodatno zadušijo na lopaticah, ki predstavljajo fleksibilno konstrukcijo. Potresni sunki v splošnem najbolj vplivajo na podnožje stebra, bistveno manj pa na njegov zgornji del. To je prednost fleksibilne konstrukcije. Pri snovanju vetrnih turbin za potresno ogrožena območja torej pridejo v poštev fleksibilne konstrukcije za dušenje vibracij. Prihodnje raziskave na podlagi izsledkov tega članka bodo usmerjene v optimalno konstrukcijo stebra. Seizmične analize na osnovi modelov morajo upoštevati vse obremenitvene primere. Članek podaja predlog modela po teoriji dinamike sistema več teles za lopatice, vrh, steber in temelj vetrne turbine, vključujoč interakcijo med tlemi in konstrukcijo za preučitev obremenitvenih razmer na vetrnih turbinah ob potresu. Rezultati predstavljene raziskave bodo uporabljeni pri snovanju vetrnih turbin za potresno ogrožena območja. Ključne besede: vetrna turbina, potres, dinamika več teles, interakcija tal in konstrukcije, Eurocode 8, obremenitvene razmere

*Naslov avtorja za dopisovanje: Univerza v Chongqingu, Kolidž za strojništvo, Chongqing 400044, Kitajska, jinxin191@hotmail.com

SI 119

Prejeto v recenzijo: 2014-03-04 Prejeto popravljeno: 2014-04-30 Odobreno za objavo: 2014-05-16

Računski model za analizo upogibne trdnosti sintranih zobnikov Srečko Glodež1,* – Marko Šori2 – Tomaž Verlak1 1 Univerza

2 Univerza

v Mariboru, Fakulteta za strojništvo, Slovenija v Mariboru, Fakulteta za naravoslovje in matematiko, Slovenija

V času, ko je potrebno zagotavljati kakovostne izdelke z nizko ceno in čim manj odpada, je tehnologija avtomatskega stiskanja kovinskih prahov v končno ali skoraj končno obliko in sintranja (ADC/SINT – Automatic die compaction and sintering) postala zelo zanimiva za različna področja industrije, še posebej za avtomobilsko industrijo. Že nekaj časa je prašna metalurgija prisotna pri različnih delih volanskega sklopa, zobniških črpalkah in različnih pomožnih delih, zadnje čase pa se pojavlja kot zanimiva, cenovno ugodnejša alternativa pri izdelavi zobnikov za menjalnike. Obstoječi ISO-standardi za preračun zobnikov ne predvidevajo materialov iz skupine prašne metalurgije, tako da obstajajo le informacijska priporočila po AGMA, ki podajajo oceno vrtilnega momenta, ki ga lahko sintran zobnik prenaša. Ta priporočila zajemajo upoštevanje podatkov o materialih, obsežen in zapleten del teh priporočil pa upošteva geometrijo zob zobnika. Zaradi kompliciranosti postopka pri pridobivanju koeficientov geometrije zoba, je bil v okviru te raziskave razvit model, ki napove obratovalno dobo določenega sintranega zobnika, če so znane njegove materialne lastnosti pri dinamičnih obremenitvah. V okviru eksperimentalnega dela so v sklopu raziskav najprej s kvazi-statičnim nateznim preizkusom pri sobni temperaturi določene statične mehanske lastnosti uporabljenega sintranega materiala: modul elastičnosti, meja tečenja in natezna trdnost. Sledi dinamično testiranje pri različnih napetostih z obremenitvenim razmerjem R = 0, na osnovi katerega se določi vpliv amplitudne napetosti na dobo trajanja preskušanca. Končni rezultat raziskav je določitev dveh osnovnih materialnih parametrov za dimenzioniranje dinamično obremenjenih komponent z napetostno metodo: koeficienta trdnosti pri utrujanju sf‘ in eksponenta trdnosti pri utrujanju b. V sklopu raziskave so bile določene lastnosti materiala v dveh različnih stanjih. Polovica preskušancev je bila posintrana, druga polovica dodatno zakaljena. Preskušanci so bili izdelani in toplotno obdelani po priporočilih proizvajalca uporabljenega kovinskega prahu. Izkazalo se je, da dodatna toplotna obdelava materiala bistveno poveča natezno trdnost in zmanjša raztezek ob prelomu. Rahlo se poveča tudi modul elastičnosti. Občutna razlika je tudi pri dinamični trdnosti, vendar se ta razlika manjša in skoraj izgine pri nižjih napetostih oziroma pri daljši dobi trajanja, kar se kaže tudi kot večji naklon premice Basquin-ove enačbe v dvojnem logaritmičnem diagramu. Glavni fokus raziskave je določitev upogibne trdnosti zobnika v korenu zoba, ki je lahko pri sintranih zobnikih poljubne oblike. Matrica, v kateri se stiskajo sintrani zobniki, je običajno izdelana z žično erozijo (wire EDM), ki omogoča poljubno obliko korena zoba. Ker se pri tem postopku izdelave matrice potrebuje vsaj 2D digitalni model zobnika, se lahko isti model uporabi za hitro pripravo MKE modela, s pomočjo katerega se določi napetostno polje v korenu zoba. V numerični simulaciji so vključeni trije zobje obravnavanega zobnika in en zob zobnika, ki je togo vpet v prostoru. Obravnavan zobnik je postavljen k togo vpetemu zobniku tako, da je kontakt med njima na zunanji točki enojnega ubiranja obravnavanega zobnika. Obremenitev je izvedena z rotacijo obravnavanega zobnika okoli svoje osi tako, da se nasloni na tog zob. Celotna rotacija obravnavanega zobnika je izvedena v več korakih, kar omogoča boljšo stabilnost numerične simulacije. Napetosti v korenu zoba pri različnih obremenitvenih primerih so potem preko Basquin-ove enačbe s predhodno določenimi materialnimi parametri pretvorjene v predvideno dobo trajanja zobnika oziroma število nihajev obremenitve do poškodbe v korenu zoba. Rezultati računskih analiz po predlaganem modelu kažejo, da imajo dodatno zakaljeni zobniki predvsem v območju časovne trdnosti občutno večjo upogibno dinamično trdnost v primerjavi z nekaljenimi zobniki. Ta razlika pa je vse manjša, čim bolj se približujemo območju trajne trdnosti (106 ali več nihajev obremenitve). Vendar pa je potrebno poudariti, da ta metoda zajema zgolj trdnost korena zoba in ne obrabe zobnih bokov, ki bo predvidoma pri samo sintranih zobnikih bistveno večja. Nadaljnje raziskovalno delo bo zajemalo dejansko preskušanje sintranih zobnikov na ustreznem preskuševališču, ki bodo izdelani na enak način in iz iste prašne mešanice, kot so bili izdelani natezni preskušanci. Potrebno bo določiti, koliko obratovalnih ciklov zdrži posamezen zobnik ob določenem momentu. Ti rezultati bodo pokazali, ali je predlagana metoda primerna oziroma dovolj natančna za določitev nosilnosti korena zoba sintranih zobnikov. Ključne besede: sintrani zobniki, napetostna metoda, numerično modeliranje, preskusi SI 120

*Naslov avtorja za dopisovanje: Univerza v Mariboru, Fakulteta za strojništvo, Smetanova ulica 17, 2000 Maribor, Slovenija, srecko.glodez@um.si

Prejeto v recenzijo: 2014-01-10 Prejeto popravljeno: 2014-05-20 Odobreno za objavo: 2014-07-08

Optimizacija vmesnih spremenljivk pri mikroelektroerozijski obdelavi nerjavnega jekla 316L s pomočjo genetskega algoritma in metodologije odzivne površine Suresh, P. – Venkatesan, R. – Sekar, T. – – Elango, N. – Sathiyamoorthy, V. Periyakgounder Suresh1,* – Rajamanickam Venkatesan1 – Tamilperruvalathan Sekar2 – – Natarajan Elango3 – Varatharajan Sathiyamoorthy4 1 Tehniški

kolidž Sona, Oddelek za strojništvo, Indija kolidž za inženirstvo, Oddelek za strojništvo, Indija 3 Univerza Linton, Fakulteta za strojništvo, Malezija 4 Tehniški kolidž Mahendra, Oddelek za strojništvo, Indija

2 Vladni

Cilj predstavljene raziskave je določitev optimalnih vrednosti glavnih vmesnih spremenljivk pri mikroelektroerozijski obdelavi nerjavnega jekla 316L. EDM je nekonvencionalen postopek obdelave, pri katerem kovino odstranjuje močna električna iskra med elektrodo in materialom obdelovanca. Postopek se uspešno uporablja pri trdih prevodnih materialih. Nerjavno jeklo 316L, ki je obravnavano v tej raziskavi, se uporablja v farmaciji, ladjedelništvu in medicini. Zaradi svojih lastnosti, kot so visoka obstojnost proti oksidaciji in koroziji ter trdota, se je uveljavilo pri medicinskih vsadkih, kot so čepi in vijaki, ter pri ortopedskih vsadkih, kot so totalne kolčne in kolenske proteze. Na obdelovalnost pri postopku mikroEDM sicer vplivajo mnogi parametri procesa, v tej raziskavi pa so bili obravnavani trije pomembnejši parametri: jakost toka, vklopni čas impulza (Ton) in izklopni čas impulza (Toff). Metoda Taguchi je najzmogljivejše orodje za izboljšanje produktivnosti, zato je bila v preteklosti že uporabljena tudi na področju obdelave različnih materialov. Gre za enostaven, učinkovit in sistematičen pristop k določanju optimalnih procesnih parametrov, ki drastično zmanjša število potrebnih eksperimentov za modeliranje odzivnih funkcij. Metoda temelji na eksperimentih v ortogonalnem polju (OA), ki zagotavljajo bistveno manjšo varianco eksperimentov ter optimalno nastavitev parametrov krmiljenja procesa. Metoda Taguchi je bila najprej uporabljena za določitev optimalnih parametrov procesa in števila eksperimentov, potrebnih za modeliranje odzivnih funkcij. Izbrano je bilo ortogonalno polje L9 s tremi vrsticami in devetimi stolpci. Opravljeni so bili eksperimenti s 400-mikrometrskimi medeninastimi elektrodami. Glavni vmesni parametri so bili jakost toka, vklopni čas impulza in izklopni čas impulza na treh ravneh. Nato je bila uporabljena metodologija odzivne površine (RSM) za določitev korelacije med vmesnimi parametri ter izbranimi cilji doseganja maksimalne stopnje odvzema materiala in minimalne stopnje obrabe orodja pri obdelavi izbranega materiala. Podatki iz eksperimentov so bili nato vstavljeni v regresijski model dvofazne interakcije (2FI). Matematični model iz RSM je bil uporabljen kot funkcija uspešnosti za večciljno optimizacijo z genetskim algoritmom (GA). Optimizacija na osnovi metahevrističnih algoritmov se začne z začetnim naborom neodvisnih spremenljivk in se nato razvije za pridobitev globalnega minimuma/maksimuma ciljne funkcije (uspešnosti). Cilja optimizacije pri tej raziskavi sta bila maksimalna stopnja odvzema materiala in minimalna stopnja obrabe orodja. Uporabljena je bila zbirka orodij za večciljne analize MATLAB GA in matematični modeli iz RSM so bili uporabljeni za napovedovanje optimalnih parametrov procesa. V orodju je bila izbrana populacija velikosti 45, navzkrižni delež 0,8 in razpršena funkcija križanja. Orodje privzeto uporablja dvotočkovno navzkrižno funkcijo. Stopnja mutacije je bila 0,01. Preučen in prikazan je vpliv na stopnjo odvzema materiala in stopnjo obrabe orodja pri eksperimentalnih vrednostih Ton = 3, 6 in 9 μs ter pri Toff = 3,6 in 9 μs. Optimalne rešitve za obdelavo izbranega materiala so zbrane v preglednici in sledi diskusija. Večciljni GA napoveduje majhno stopnjo odvzema materiala 0,4352 mg/s in stopnjo obrabe orodja 0,0122 mg/s pri Ton = 3,3608 μs, Toff = 8,6356 μs in jakosti toka 6,0263 A. Visoka stopnja odvzema materiala je zabeležena pri stopnji obrabe orodja 0,2391 mg/s ob vrednostih Ton = 8,9999 μs, Toff = 8,9185 μs in jakosti toka 11,9991 A. Potrditveni poskusi so pokazali, da rezultati optimizacije odstopajo od eksperimentalnih vrednosti za manj kot 5 %. Ključne besede: metodologija odzivne površine, genetski algoritem, nerjavno jeklo 316L, metoda Taguchi

*Naslov avtorja za dopisovanje: Tehniški kolidž Sona, Oddelek za strojništvo, Junction Main Road, Salem-5, TN, Indija, suresh_p_g@yahoo.com

SI 121

Prejeto v recenzijo: 2014-01-17 Prejeto popravljeno: 2014-04-20 Odobreno za objavo: 2014-05-07

Snovanje in raziskava vodnega nizkotlačnega pilotnega elektromagnetnega ventila za velike pretoke Liu, L. – Zhang, D. – Zhao, J. Liu Lei1 – Zhang Desheng2,3 – Zhao Jiyun1,*

1 Šola

za strojništvo in elektrotehniko, Kitajska rudarska in tehniška univerza, Kitajska 2 Oddelek za premogovništvo, Tiandi Znanost in tehnologija, Kitajska 3 Šola za strojništvo, elektroniko in informatiko, Kitajska rudarska in tehniška univerza, Kitajska

Sklopke s krmilnimi ventili so razširjene pri težkih premogovniških skreperjih (transporterjih), tračnih transporterjih, črpalkah, sesalnih ventilatorjih in drugi težki opremi, saj imajo pomembno vlogo pri izboljševanju obratovalnih pogojev in pri varčevanju z energijo. Sestav nizkotlačnega ventila za velike pretoke je eden ključnih delov sklopke s krmilnimi ventili za velike moči. Osnovne zahteve, ki jih mora izpolnjevati sklopka s krmilnimi ventili, so odzivnost, velika zmogljivost pretoka in varnost pred blokiranjem. Cilj raziskave je zasnova vodnega nizkotlačnega pilotnega elektromagnetnega ventila za velike pretoke. Raziskana je porazdelitev tlaka med dušilno šobo in pilotnim ventilom ter pretočne značilnosti na dušilni odprtini. Vpliv premera dušilne šobe in drugih parametrov na dinamične in statične lastnosti ventila je bil analiziran s programsko opremo za simulacije AMESim. Določeni so bili optimalni parametri sestava elektromagnetnega ventila. Vpliv premera dušilne šobe in drugih parametrov na dinamične in statične lastnosti ventila je bil analiziran s programsko opremo za simulacije AMESim. Postavljeno je bilo hidravlično preizkuševališče na čisto vodo. Raziskana je bila tudi porazdelitev vlaka med dušilno šobo in pilotnim ventilom ter pretočne značilnosti v dušilni odprtini. Sestav pilotnega elektromagnetnega ventila je bil zasnovan na osnovi logičnih povezav pri delovanju polnilnega, cirkulacijskega in praznilnega ventila. Glavni ventil ima radialno tesnilo, čelna površina pa mehko ravno tesnilo. Obe tesnili sta prilagojeni značilnostim vode. Dvostopenjsko dušenje ravne in konične konstrukcije lahko zmanjša udar pri odpiranju in zapiranju. Simulacijski model polnilnega ventila je bil postavljen s programsko opremo AMESim. Raziskan je bil vpliv tekočinskega blaženja na statične in dinamične značilnosti krmilnega ventila. Rezultati simulacije kažejo, da na odzivni čas močno vplivata premer dušilne šobe in togost vzmeti. Večja kot sta premer dušilne šobe in togost vzmeti, hitrejši je odziv, medtem ko tlak tekočine nima večjega vpliva. Odpiralni tlak in stabilen delovni tlak ventila sta majhna, s čimer je izpolnjena zahteva po nizkem tlaku in velikem pretoku. Ni posebnih omejitev, razen natančnosti obdelave konusa glavnega drsnika. Postavljeno je bilo hidravlično preskuševališče na čisto vodo. Zasnovani sestav elektromagnetnega ventila ima boljše nizkotlačne lastnosti in hitrejši odziv, s čimer so izpolnjene zahteve za sklopko s krmilnimi ventili. Novost, ki je predstavljena v članku, je zasnova vodnega nizkotlačnega pilotnega elektromagnetnega ventila za velike pretoke. Določeni so bili optimalni parametri sestava elektromagnetnega ventila. Ključne besede: voda kot medij, nizek tlak, velik pretok, pilotni elektromagnetni ventil, dušilna odprtina, zgradba glavnega ventila

SI 122

*Naslov avtorja za dopisovanje: Šola za strojništvo in elektrotehniko, Kitajska rudarska in tehniška univerza, Xuzhou, 221116, Kitajska, jyzhao@cumt.edu.cn

Prejeto v recenzijo: 2013-02-25 Prejeto popravljeno: 2013-06-13 Odobreno za objavo: 2013-08-20

Optimizacija struženja zlitine Ti6Al4V s podporo visokotlačne hladilne tekočine Oğuz Çolak Univerza Süleymana Demirela, Center za raziskave in aplikacije CAD/CAM, Turčija

Zlitina Ti6Al4V je razširjena predvsem v letalski in vesoljski industriji ter pri izdelavi medicinskih izdelkov, zaradi svojih lastnosti pa spada med materiale, ki jih je najtežje obdelovati z odrezavanjem. Obdelava te zlitine brez pomožnega hlajenja je zelo težavna zaradi visokih temperatur v območju rezanja, ki so delno posledica slabe toplotne prevodnosti materiala in povzročajo hitro obrabo orodja. Iskanje rešitev za odpravo tega problema in izboljšanje obdelovalnosti zlitine Ti6Al4V gre v smeri razvoja različnih tehnik hlajenja in mazanja. En glavnih pristopov je hlajenje s pomočjo visokotlačnega curka. Osnovni princip je izkoriščanje mehanskega in toplotnega učinka hladilne tekočine pod visokim tlakom, ki se brizga na stik orodja in odrezka za boljše mazanje in hlajenje. Znano je, da imajo operacije obdelave številne vhodne parametre, ki vplivajo na stroške obdelave in na kakovost izdelkov. Izbira optimalnih parametrov obdelave je zato zelo pomembna za industrijo. Večja učinkovitost obdelave z visokotlačnim hlajenjem se dosega z optimizacijo parametrov visokotlačnega hlajenja in odrezavanja. Članek obravnava optimizacijo parametrov obdelave ob upoštevanju ukrepov za izboljšanje struženja zlitine Ti6Al4V na konvencionalen način in z visokotlačnim hlajenjem. V ta namen je bil v eksperimentalni študiji uporabljen genetski algoritem za iskanje optimalnih parametrov odrezavanja. Privzeti so bili trije optimizacijski kriteriji za izboljšanje učinkovitosti obdelave: površinska hrapavost, stopnja odvzema materiala in moč odrezavanja. Opravljena je bila vrsta preskusov obdelave Ti6Al4V pri konvencionalnih in različnih visokotlačnih pogojih hlajenja. Za preskuse je bilo izbrano orodje CNMG0812 s prevleko (Ti,Al)N+TiN. Eksperimenti so bili zasnovani na podlagi Taguchijevega ortogonalnega polja L9 s tremi različnimi rezalnimi hitrostmi, podajanji in ravnmi tlaka. Med eksperimentalno obdelavo pri konstantni globini reza so bile merjene komponente rezalne sile in površinska hrapavost, prav tako pa sta bili izračunani stopnja odvzema materiala in rezalna moč glede na parametre odrezavanja in eksperimentalni odziv. Za sestavljanje ciljne funkcije pri vsaki ravni tlaka so bile z multiregresijsko analizo pridobljene empirične enačbe, ki popisujejo odvisnost med pogoji odrezavanja in eksperimentalnim odzivom. Maksimum ciljnih funkcij je bil ugotovljen z genetskim algoritmom in opredeljeni so bili optimalni parametri obdelave. Opravljeni so bili tudi preskusi obrabe orodja pri pogojih odrezavanja, ki so blizu optimalnim parametrom obdelave. Na podlagi hibridnega modela so bila izrisana sprejemljiva območja z mejami ukrepov za učinkovitost obdelave pri vsakokratnih pogojih hlajenja. Rezultati optimizacije so pokazali, da so sprejemljiva območja v pogojih visokotlačnega hlajenja nekoliko manjša kot pri konvencionalnem hlajenju. Vzrok za to je naraščajoči vpliv visokotlačnega hlajenja na površinsko hrapavost – kakovost površine se zmanjšuje z naraščanjem tlaka. Optimalni parametri odrezavanja in odziv pri pogojih hlajenja P = 6 in 150 bar so si sicer precej podobni, kar pa ne velja za življenjsko dobo orodja pri različnih pogojih hlajenja. Življenjska doba orodij pri tlaku P = 150 bar oz. P = 300 bar je za približno 47 oz. 112 % daljša kot pri konvencionalnem hlajenju. To je mogoče pojasniti z dejstvom, da brizganje visokotlačne hladilne tekočine na stik orodja in odrezka zagotavlja učinkovito mazanje in hlajenje s penetracijo v območje rezanja, zaradi česar se občutno zniža temperatura. Obrabo orodja zaradi visokih temperatur je zato mogoče zmanjšati ali v celoti preprečiti, življenjska doba orodja pa se podaljša v primerjavi s konvencionalnim hlajenjem. Optimizacija po drugi strani ni dala pomembnejših sprememb pri rezalni moči, ne glede na pogoje hlajenja. Visokotlačno hlajenje pri eksperimentih namreč ni pomembneje vplivalo na glavne rezalne sile. Hibridni model, ki je bil uporabljen v raziskavi za optimizacijo učinkovitosti obdelave, je zelo primeren za opredeljevanje optimalnih parametrov odrezavanja glede na dane kriterije optimizacije. Uporaba visokotlačnega hlajenja pri obdelavi težavnih materialov gre tudi v prid trendu trajnostnosti proizvodnje zaradi daljše življenjske dobe orodij in manjših stroškov obdelave. Ključne besede: visokotlačno hlajenje, optimizacija, življenjska doba orodja, zlitina Ti6Al4V, hibridni model, obdelava

*Naslov avtorja za dopisovanje: Univerza Süleyman Demirel, CAD/CAM razisiskovalno aplikativni center, 32260, Isparta, Turčija, kadirkiran@sdu.edu.tr

SI 123

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, SI 124-130 Osebne objave

Doktorske disertacije, diplomske naloge

DOKTORSKE DISERTACIJE Na Fakulteti za strojništvo Univerze v Ljubljani so obranili svojo doktorsko disertacijo: ● dne 9. septembra 2014 Žiga GOSAR z naslovom: »Preskok sistema plitke dvoslojne dvakrat povezane lupine« (mentor: prof. dr. Franc Kosel); V doktorskem delu je obravnavan preskok sistema dvakrat povezane plitke dvoslojne lupine v homogenem temperaturnem polju. Lupina ima zvezdasto odprtino v temenu. Obravnava problema zahteva uporabo teorije tretjega reda po Chwalli, kar pomeni, da obravnava ravnovesje sil in momentov na deformiranem sistemu ob sočasnem upoštevanju teorije velikih premikov. Na osnovi izbrane nelinearne teorije je bilo izdelan fizikalno matematični model, s katerim je popisana geometrija sistema, premično in napetostno termo-mehansko stanje. Model lupine je izdelan z metodo končnih elementov. Preskok sistema plitke lupine je obravnavan za različne načine vpetja v homogenem temperaturnem polju. Lupina je poleg temperaturnih obremenitev dodatno obremenjena z zunanjimi obremenitvami, ki delujejo na notranjem in/ ali zunanjem robu lupine. Točnost rezultatov sferičnih dvakrat povezanih plitkih bimetalnih lupin so preverjeni z eksperimenti, izvedenimi v laboratoriju; ● dne 18. septembra 2014 Tadej KRANJC z naslovom: »Karakterizacija dinamičnega obnašanja kompleksnih struktur z analizo specifične deformacije v elastičnem območju« (mentor: prof. dr. Miha Boltežar, somentor: izr. prof. dr. Janko Slavič); Raziskava se osredotoča na analizo kompleksnih struktur. Izvedeni sta teoretična in eksperimentalna primerjava klasične eksperimentalne modalne analize (EMA) in eksperimentalne modalne analize, ki temelji na merjenju specifičnih deformacij (EMASD). EMASD ima v primerjavi z EMA določene prednosti, poleg tega pa tudi nekatere pomanjkljivosti. Pomembnejša pomanjkljivost je, da identificiranih lastnih oblik ni mogoče masno normirati. V ta namen je razvit pristop masnega normiranja, ki temelji na strukturni modifikaciji na osnovi dodajanja mas. Poleg modalne analize se raziskovalno delo osredotoča tudi na identifikacijo podstruktur na osnovi identifikacije povezovalnih sil z merjenjem specifičnih deformacij; ● dne 19. septembra 2014 David KOBLAR z naslovom: »Identifikacija parametrov vibroizolacije za numerično napoved njene učinkovitosti« (mentor: prof. dr. Miha Boltežar); V raziskavi je predstavljena identifikacija parametrov vibroizolacije za uporabo pri numerični SI 124

napovedi njene učinkovitosti. Na različnih enostavnih sistemih z eno prostostno stopnjo (SDOF sistem) smo izpeljali enačbe za identifikacijo parametrov vibroizolacijskega materiala. Tako identificirane frekvenčno odvisne parametre, frekvenčno odvisna modul elastičnosti in faktor dušenja, smo nato uporabili v modelu iz končnih elementov. To nam omogoči napoved učinkovitosti vibroizolaicje tudi na zahtevnejših geometrijah vibroizolacijskega materiala. Numerično napoved učinkovitosti vibroizolacije smo izračunali za realno kompleksno strukturo in jo preverili z eksperimentom; ● dne 23. septembra 2014 Boštjan PARADIŽ z naslovom: »Emisije dioksinov in furanov pri zgorevanju premoga v peči« (mentor: prof. dr. Vincenc Butala, somentor: izr. prof. dr. Ivan Bajsić); Pri zgorevanju črnega premoga z visoko vsebnostjo klora v peči za ogrevanje gospodinjstev smo izmerili zelo visoke koncentracije PCDD/F v dimnih plinih. Te so najvišje do sedaj objavljene za to vrsto naprav in kar tri razrede velikosti višje, kot v sodobnih sežigalnicah odpadkov. Ugotovili smo izrazit vpliv temperaturnega profila dimnih plinov v dimniku na emisije PCDD/F, kar nakazuje njihov nastanek v dimniku. Emisije PCDD/F so bile en velikostni razred večje pri toplotno izoliranem dimniku, kot pri toplotno neizoliranem, toplotna izolacija dimnika pa ni vplivala na emisije CO, NOx, delcev in PAH. Pri premogu z nizko vsebnostjo klora so bile emisije PCDD/F dva velikostna razreda nižje, kot pri premogu z visoko vsebnostjo klora, hkrati pa je bil vpliv temperaturnega profila v dimniku manj izrazit. V laboratorijskih pogojih se toplotne lastnosti dimnika običajno razlikujejo od tistih v stavbah. Zato je potrebno preveriti reprezentativnost emisijskih faktorjev za pripravo evidenc emisij PCDD/F, ki so bili pridobljeni z meritvami v laboratorijskih pogojih; ● dne 26. septembra 2014 Jan ŠKOFIC z naslovom: »Karakterizacija vibro-akustičnega obnašanja mehatronskih sistemov s koračnim motorjem« (mentor: prof. dr. Miha Boltežar); Doktorska naloga obravnava celostni pristop k vibro-akustični karakterizaciji mehatronskih sklopov. Zaradi obsežnosti področja mehatronskih sklopov, ki povezuje tako mehanske kot tudi elektronske komponente, je problematiksa predstavljena na realnih industrijskih izdelkih, saj bi splošna obravnava krepko presegla razpoložljive vire in časovne omejitve. Obravnavani sklop sestoji iz koračnega motorja s trajnim magnetom, zobniškega sistema in ohišja ter tako tvori majhen mehatronski reduktor za

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, SI 124-130

krmiljenje nove generacije pametnih avtomobilskih luči. Na tem primeru je bilo obravnavano rotacijsko in aksialno gibanje rotorja koračnega motorja, generacija dinamičnih sil na ležajna mesta zobniškega sistema ter nazadnje vibro-akustični odziv sklopa glede na določene parametre obratovanja; ● dne 29. septembra 2014 Matjaž ČEBRON z naslovom: »Geometrijska optimizacija izdelka iz gradiva z utrjevalno karakteristiko« (mentor: prof. dr. Franc Kosel); Doktorska naloga obravnava opis utrjevanja čistih kovinskih materialov s ploskovno centrirano kubično rešetko po teoriji translacijskih dislokacij. Dva dislokacijska modela utrjevanja sta vgrajena v Taylorjev visko-plastični model deformiranja polikristalov in uporabljena za opis natezne napetostno-deformacijske krivulje čistega bakra pri temperaturi zraka v prostoru in nizkih hitrostih deformiranja. Predlagana je uporaba energijskih modelov dislokacijske strukture in diferenčne dinamične kalorimetrije kot hiter in preprost način zagotavljanja fizikalno sprejemljivih izračunanih vrednosti gostote dislokacij. V drugem delu je po evolucijski optimizacijski strategiji izvedena geometrijska optimizacija izdelka iz gradiva z utrjevalno karakteristiko, v kateri je upoštevano povečanje meje tečenja zaradi namensko povzročene plastične deformacije.

Matic SAJOVIC z naslovom: »Konstruiranje pomožne mize za napravo za merjenje zaostalih napetosti X-stress 3000« (mentor: prof. dr. Nenad Gubeljak). * Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv magister inženir strojništva: dne 1. septembra 2014: Jernej LUPŠINA z naslovom: »Stroj za masno uravnoteženje zavornih diskov železniških vozil« (mentor: prof. dr. Janez Kopač); Marko MRAK z naslovom: »Razvoj optičnega merilnika topografije kože« (mentor: doc. dr. Matija Jezeršek); * Na Fakulteti za strojništvo Univerze v Mariboru je pridobil naziv univerzitetni diplomirani gospodarski inženir: dne 25. septembra 2014: Igor KOLARIČ z naslovom: »Optimizacija zalog vhodnega materiala v podjetju CM-OS Mastnak« (mentorja: doc. dr. Iztok Palčič, prof. dr. Vojko Potočan). *

DIPLOMSKE NALOGE Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv univerzitetni diplomirani inženir strojništva: dne 1. septembra 2014: Denis MAVRIČ z naslovom: »Optimizacija parametrov laserskega rezanja jeklene pločevine« (mentor: doc. dr. Matija Jezeršek); Matjaž ŽONTA z naslovom: »Informacijski sistem za obvladovanje tveganj in podporo odločanju pri projektnem vodenju« (mentor: prof. dr. Peter Butala). * Na Fakulteti za strojništvo Univerze v Mariboru sta pridobila naziv univerzitetni diplomirani inženir strojništva: dne 8. septembra 2014: David NOVINA z naslovom: »Inovativna oblikovna zasnova dirkalnika na podlagi magnetne levitacije« (mentor: izr. prof. Vojmir Pogačar, somentor: prof. dr. Jana Padežnik Gomilšek); dne 11. septembra 2014:

Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv magister inženir strojništva: dne 18. septembra 2014: Primož ŠTEFANE z naslovom: »Lomno obnašanje konstrukcijskega jekla v temperaturno prehodnem območju« (mentor: prof. dr. Nenad Gubeljak, somentor: izr. prof. dr. Jožef Predan); dne 24. septembra 2014: Boštjan FERK z naslovom: »Merjenje indeksa pretoka taline materialov za lasersko sintranje« (mentor: izr. prof. dr. Igor Drstvenšek, somentor: dr. Tomaž Brajlih); Peter JURGEC z naslovom: »Projektiranje in optimizacija montažne linije« (mentor: doc. dr. Iztok Palčič, somentor: prof. dr. Borut Buchmeister); Peter KIRBIŠ z naslovom: »Razvoj nanostrukturnega bainitnega jekla« (mentor: prof. dr. Ivan Anžel, somentor: doc. dr. Mihael Brunčko); Iztok ŠOSTER z naslovom: »Modeliranje in dinamična analiza več-masnih sistemov s programom Maple« (mentor: izr. prof. dr. Karl Gotlih); Edis VEHABOVIĆ z naslovom: »Integracija hidravličnih dovodnih kanalov v ohišje univerzalnega prijemala« (mentor: izr. prof. dr. Bojan Dolšak, somentor: prof. dr. Srečko Glodež); SI 125

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, SI 124-130

dne 25. septembra 2014: Janez JELEN z naslovom: »Primerjava SPTE sistemov za termično izrabo biomase v hotelskem kompleksu« (mentor: prof. dr. Aleš Hribernik); Blaž OREŠNIK z naslovom: »Numerična analiza mešanja sipkega materiala« (mentor: prof. dr. Matjaž Hriberšek, somentor: dr. Matej Zadravec); dne 29. septembra 2014: Gašper ZDOVC z naslovom: »Zasnova montažne linije mikrokogeneracijskih enot v podjetju Indop« (mentor: prof. dr. Borut Buchmeister somentor: doc. dr. Iztok Palčič). * Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv magister inženir tehniškega varstva okolja: dne 3. septembra 2014: Marko BUDLER z naslovom: »Izzivi in inovativni postopki pri reciklaži redkozemeljskih magnetov SmCo« (mentorja: prof. dr. Ivan Anžel, izr. prof. dr. Zdenka Ženko); dne 3. septembra 2014: Franci GOLEŽ z naslovom: »Vpliv znanstvenega in tehnološkega razvoja na uspešnost organizacij« (mentorja: doc. dr. Marjan Leber, prof. dr. Vojko Potočan). * Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv magister inženir tehniškega varstva okolja: dne 10. septembra 2014: Ana KRAČUN z naslovom: »Analiza filtrskega prahu iz podjetja Talum Livarna d.o.o.« (mentor: prof. dr. Ivan Anžel, somentorica: izr. prof. dr. Lidija Fras Zemljič). * Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv magister inženir oblikovanja izdelkov: dne 24. septembra 2014: Valentina URBANČIČ z naslovom: »Oblikovanje naprave za gojenje vrtnin v stanovanju« (mentor: izr. prof. Vojmir Pogačar). * Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv diplomirani inženir strojništva (UN): dne 1. septembra 2014: Maša TOŠKAN, Janez TRČEK, Bernard TRČEK in Matej ŽUPAN; SI 126

* Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv diplomirani inženir strojništva (UN): dne 4. septembra 2014: Mihael BREZNIK z naslovom »Parametrična analiza napetostnega stanja upogibno in torzijsko obremenjenega prehoda premera gredi« (mentor: prof. dr. Zoran Ren, somentor: dr. Matej Borovinšek); Aljaž GORČAN z naslovom »Konstruiranje transportnega vozička za prevoz strojnih delov« (mentor: prof. dr. Zoran Ren, somentor: dr. Matej Borovinšek); Kevin HOJNIK z naslovom »Modeliranje in optimizacija topologije aluminijastega stenskega nosilca« (mentor: doc. dr. Boštjan Harl, somentor: izr. prof. dr. Marko Kegl); Marko HORVAT z naslovom »Obnova stranskega dela orodja za litje aluminija« (mentor: izr. prof. dr. Miran Ulbin); Anton KOČAR z naslovom »Določitev meje tečenja lotus poroznega železa z računalniškimi simulacijami« (mentor: prof. dr. Zoran Ren, somentor: Aljaž Kovačič); Tomaž STOJAN z naslovom »Vakuumska cementacija jekel« (mentor: doc. dr. Mihael Brunčko, somentor: doc. dr. Tomaž Vuherer); Rok ŠTANCER z naslovom »Konstruiranje hidravličnega cepilca 17T z mobilno črpalko« (mentor: prof. dr. Srečko Glodež, somentor: doc. dr. Janez Kramberger); Blaž ZAVRATNIK z naslovom »Računalniško podprto konstruiranje hidravlične vpenjalne priprave« (mentor: izr. prof. dr. Miran Ulbin); dne 5. septembra 2014: Miran KLANČNIK z naslovom »Konstruiranje fitnes igrala« (mentor: doc. dr. Aleš Belšak, somentor: izr. prof. dr. Miran Ulbin); Simon KORAŽIJA z naslovom » Oblikovanje armaturne plošče v osebnem avtomobilu« (mentor: doc. dr. Andrej Skrbinek); Blaž KOŠIR z naslovom »Zmanjšanje prekomernega prašenja pri tračnih transporterjih« (mentor: prof. dr. Iztok Potrč, somentor: izr. prof. dr. Tone Lerher); Benjamin KUHAR z naslovom »Dimenzioniranje izvozne mize stiskalnice« (mentor: doc. dr. Janez Kramberger, somentor: prof. dr. Srečko Glodež); Matija MLAKAR z naslovom »CFD simulacija vbrizgavanja zraka v aeracijski bazen centralne čistilne naprave Ptuj« (mentor: prof. dr. Matjaž Hriberšek, somentor: dr. Matej Zadravec); Primož SKERBIŠ z naslovom »Numerična analiza kavitacije v toku okoli valjaste ovire« (mentor:

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, SI 124-130

doc. dr. Ignacijo Biluš, somentor: prof. dr. Brane Širok); Jakob STRAŠEK z naslovom »Projektiranje 0,5 l tlačnega reaktorja« (mentor: prof. dr. Srečko Glodež); Jan ŠIMIĆ z naslovom »Numerična simulacija procesa zgorevanja v dizelskem motorju ob uporabi različnih modelov zgorevanja« (mentor: prof. dr. Breda Kegl, somentor: Luka Lešnik); Žiga VOLK z naslovom »Dimenzioniranje nosilne konstrukcije kogeneracijske enote« (mentor: prof. dr. Srečko Glodež, somentor: doc. dr. Janez Kramberger); Jani ZAGORŠEK z naslovom »Razvoj upogibnega orodja za izdelavo avtomobilske komponente« (mentor: izr. prof. dr. Stanislav Pehan); dne 8. septembra 2014: Rok HRIBERNIK z naslovom »Analiza gospodarnosti proizvodnje pri predelavi lesa« (mentor: doc. dr. Iztok Palčič); Uroš KOREN z naslovom »Kreativno oblikovanje inovacijskih idej pri razvoju novega izdelka« (mentor: doc. dr. Marjan Leber, somentor: doc. dr. Iztok Palčič); Andraž KRANJC z naslovom »Konstruiranje vzmetnega mehanizma za dirkalnik GPE14« (mentor: prof. dr. Srečko Glodež); Gregor TUMPEJ z naslovom »Zasnova dvižnega reduktorja mostnega dvigala« (mentor: doc. dr. Aleš Belšak, somentor: izr. prof. dr. Miran Ulbin); dne 11. septembra 2014: Luka FEKONJA z naslovom »Vpliv nezadostne in prevelike toplotne obdelave TVP na udarno žilavost na jekla za povišane temperature« (mentor: doc. dr. Tomaž Vuherer, somentor: prof. dr. Franc Zupanič); Rok PAHIČ z naslovom »Vpenjalni mehanizem« (mentor: izr. prof. dr. Karl Gotlih, somentor: doc. dr. Janez Kramberger); Klemen SENOVRŠNIK z naslovom »Vpliv neprimerne toplotne obdelave po varjenju na trdoto TVP jekla odpornega na lezenje« (mentor: doc. dr. Tomaž Vuherer, somentor: doc. dr. Gorazd Lojen); dne 16. septembra 2014: Jure BREČKO z naslovom »Konstruiranje naprave za polnjenje cisterne« (mentor: prof. dr. Srečko Glodež); Peter KLOBASA z naslovom »Preračun transportne naprave« (mentor: prof. dr. Iztok Potrč, somentor: izr. prof. dr. Tone Lerher); Luka LAH z naslovom »Naprava za pozicioniranje palet« (mentor: prof. dr. Iztok Potrč, somentor: izr. prof. dr. Tone Lerher); Anže MOLKA z naslovom »Razvoj in optimiranje statično in dinamično obremenjenega zvarnega spoja cevi in pločevine« (mentor: prof. dr. Zoran Ren, somentor: dr. Matej Borovinšek); dne 18. septembra 2014:

Klemen ARLIČ z naslovom »Statična analiza in izboljšava konzolno vpete nadstrešne konstrukcije« (mentor: doc. dr. Boštjan Harl , somentor: izr. prof. dr. Marko Kegl); Petar BELČIĆ z naslovom »Zasnova statično nedoločene konstrukcije strešnika« (mentor: prof. dr. Nenad Gubeljak, somentor: izr. prof. dr. Marko Kegl); Dario ČIVIĆ z naslovom »Analiza izbranih postopkov masivnega preoblikovanja« (mentor: doc. dr. Leo Gusel, somentor: prof. dr. Borut Buchmeister); Jernej FALNOGA z naslovom »Razvoj polnilne naprave orodja za sintranje polietilenskih plošč« (mentor: izr. prof. dr. Stanislav Pehan); Jakob KOSTREVC z naslovom »Analitični in numerični preračun vodnega hladilnika GPE 2014 in njuna primerjava« (mentor: izr. prof. dr. Jure Marn, somentor: dr. Jurij Iljaž); Vid OSRAJNIK z naslovom »Sistem za sledenje vetrne turbine spremembam smeri vetra« (mentor: doc. dr. Ignacijo Biluš, somentor: prof. dr. Brane Širok); Danijel SLIPČEVIĆ z naslovom »Smernice za oblikovanje orodij za brizganje plastičnih mas« (mentor: izr. prof. dr. Igor Drstvenšek, somentor: dr. Tomaž Brajlih); Robi ZEBEC z naslovom »Ergonomsko oblikovano delovno mesto v šivalnici« (mentor: doc. dr. Nataša Vujica Herzog); dne 25. septembra 2014: Dejan ČAS z naslovom »Računalniško integrirana proizvodnja« (mentor: prof. dr. Jože Balič); Matej SIMONIČ z naslovom »Računalniško podprto konstruiranje in MKE analiza plužne deske« (mentor: izr. prof. dr. Miran Ulbin); dne 29. septembra 2014: David PIRNAT z naslovom »Analiza procesa vodenja proizvodnje v Orodjarna & inženiring Alba d.o.o.« (mentor: prof. dr. Borut Buchmeister, somentor: doc. dr. Iztok Palčič). * Na Fakulteti za strojništvo Univerze v Mariboru je pridobil naziv diplomirani gospodarski inženir (UN): dne 8. septembra 2014: Rok VIDOVIČ z naslovom »Projekt nadgradnje pakirne linije v farmacevtskem podjetju« (mentor: doc. dr. Palčič, somentor: doc. dr. Igor Vrečko); dne 16. septembra 2014: Monika LEBENIČNIK z naslovom »Racionalnost uporabe robota v proizvodnji« (mentorja: izr. prof. dr. Karl Gotlih, doc. dr. Igor Vrečko); dne 18. septembra 2014: SI 127

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, SI 124-130

Janž KROFEL z naslovom »Upravljanje zalog kluba “eMCe plac”« (mentorja: doc. dr. Palčič Iztok, doc. dr. Lutar Skerbinjek Andreja); Drago KUSTER z naslovom »Izdelek BrefNEO na kitajskem trgu leta 2050« (mentorja: doc. dr. Marjan Leber, izr. prof. dr. Ženko Zdenka); Katja KOREN z naslovom »Uporaba vrednostne analize pri razvoju ščitnika okvirja kolesa« (mentorja: doc. dr. Iztok Palčič, doc. dr. Vrečko Igor); Davorin MESARIČ z naslovom »Proces vodenja razvoja izdelka s poudarkom na upoštevanju kupčevih želja« (mentorja: doc. dr. Marjan Leber, doc. dr. Igor Vrečko); dne 25. septembra 2014: Boštjan REBERNIK z naslovom »Vrednotenje karakterističnih veličin hrapavosti na izdelku« (mentorja: prof. dr. Bojan Ačko, izr. prof. dr. Zdenka Ženko); dne 29. septembra 2014: Jaka LUŠANC z naslovom »Upravičenost modernizacije tehnološke opreme polnilne linije« (mentor: doc. dr. Marjan Leber, somentor: doc. dr. Barbara Bradač Hojnik). * Na Fakulteti za strojništvo Univerze v Mariboru je pridobil naziv diplomirani inženir tehniškega varstva okolja (UN): dne 18. septembra 2014: Miha OREŠNIK z naslovom »Ravnanje z odpadki v centru za ravnanje z odpadki Ormož« (mentor: prof. dr. Niko Samec). * Na Fakulteti za strojništvo Univerze v Mariboru je pridobil naziv diplomirani inženir mehatronike (UN): dne 5. septembra 2014: Janja MIKOLIČ z naslovom »Model robota za reševanje v Gazebo simulatorju« (mentor: izr. prof. dr. Karl Gotlih, somentor: doc. dr. Suzana Uran); dne 18. septembra 2014: Tomas NOVAK z naslovom »Delo robotizirane celice skladno s predpisi ISO 10218/1-2 in direktivo 2006/42/ES (Pravilnik o varnosti strojev)« (mentor: izr. prof. dr. Karl Gotlih, somentor: izr. prof. dr. Aleš Hace). * Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv diplomirani inženir strojništva: dne 1. septembra 2014: SI 128

Boštjan STRMLJAN z naslovom: »Postopek priprave popravila podpornega rebra glavnih nog podvozja« (mentor: izr. prof. dr. Tadej Kosel, somentor: izr. prof. dr. Janez Kušar); dne 10. septembra 2014: Miran LIKAR z naslovom: »Posledice širjenja cvetnega prahu in metoda za njegovo identifikacijo« (mentor: prof. dr. Vincenc Butala); dne 12. septembra 2014: Tomaž LEVIČAR z naslovom: »Tehnološke meritve v procesu izdelave ležajnih ohišij« (mentor: prof. dr. Janez Kopač); dne 15. septembra 2014: Boštjan DRAGOŠ z naslovom: »Naprava za merjenje čistoče hidravlične kapljevine« (mentor: doc. dr. Franc Majdič); Uroš TINTOR z naslovom: »Primerjava karakteristik hidravličnih in pnevmatičnih sistemov« (mentor: doc. dr. Franc Majdič); Jan KOBLAR z naslovom: »Spajanje kontaktne plošče in bakrenih žic z uporovnim varjenjem« (mentor: prof. dr. Janez Tušek). * Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv diplomirani inženir strojništva (VS): dne 4. septembra 2014: Matej FIDERIH z naslovom: »Dimenzioniranje nosilnega ogrodja polietilenskega silosa« (mentor: doc. dr. Janez Kramberger); dne 8. septembra 2014: Primož PEČOLER z naslovom: »Postopki kovanja jekla« (mentor: izr. prof. dr. Ivan Pahole); dne 18. septembra 2014: Damir LUČIĆ z naslovom: »Kurjenje odpadnih plinov tovarne formalina Nafte Lendava v obstoječem parnem kotlu« (mentor: viš. pred. dr. Filip Kokalj). * Na Fakulteti za strojništvo Univerze v Ljubljani so pridobili naziv diplomirani inženir strojništva (VS): dne 1. septembra 2014: Gašper BERGANT z naslovom: »Zveza gredi in roke hidravličnega prijemala« (mentor: doc. dr. Boris Jerman); David BOJANEC z naslovom: »Trajektorija leta izstrelka na velike razdalje« (mentor: izr. prof. dr. Tadej Kosel); Domen GORJUP z naslovom: »Določitev lastnosti aeroprofila iz fotografije prereza krila« (mentor: izr. prof. dr. Tadej Kosel, somentor: doc. dr. Alojz Suhadolnik);

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, SI 124-130

Jonas TROJER z naslovom: »Naprava za natančno vodenje brezpilotnih helikopterjev« (mentor: doc. dr. Viktor Šajn); Žiga VREČAR z naslovom: »Vgradnja avtopilota v brezpilotno letalo« (mentor: izr. prof. dr. Tadej Kosel); Mirjan MARKEŽIČ z naslovom: »Michelsonov interferometer za merjenje pomikov z visoko ločljivostjo« (mentor: prof. dr. Janez Možina); Tine RENER z naslovom: »Vpliv parametrov na proces spajkanja po postopku CMT« (mentor: prof. dr. Janez Tušek); Aleksander STOJKOVIĆ z naslovom: »Razvoj laserskega sistema za varjenje pločevinastih dodatkov na galovo verigo« (mentor: doc. dr. Matija Jezeršek, somentor: prof. dr. Janez Tušek); Martin POVŠE z naslovom: »Kontrola dostopa kot element razvoja pametnega laboratorija« (mentor: prof. dr. Peter Butala); dne 10. septembra 2014: Špela KOKALJ z naslovom: »Izboljšanje čiščenja odpadne vode iz galvanizacij« (mentor: prof. dr. Iztok Golobič); Klemen KRAVANJA z naslovom: »Eksperimentalno ovrednotenje vplivnih parametrov na nabiranje vodnega kamna na grelni površini« (mentor: prof. dr. Iztok Golobič); Miha MURNC z naslovom: »Izboljšana metoda določitve/ocenitve rabe energije na nivoju primarne energije energijsko učinkovite stavbe« (mentor: prof. dr. Vincenc Butala); Jože SEVČNIKAR z naslovom: »Lokalno temperaturno polje ploščnega prenosnika toplote« (mentor: prof. dr. Iztok Golobič); dne 12. septembra 2014: Monika HRIBAR z naslovom: »Postavitev tehnologije robotske obdelave lesene vetrne lopatice« (mentor: doc. dr. Franci Pušavec, somentor: prof. dr. Janez Kopač); Jože REBOLJ z naslovom: »Opazovanje kavitacijske erozije s hitro kamero« (mentor: izr. prof. dr. Matevž Dular, somentor: prof. dr. Branko Širok); Jernej VOVK z naslovom: »Meritve hitrosti toka v vetrovniku in primerjava polj hitrosti z vizualizacijskimi metodami« (mentor: izr. prof. dr. Marko Hočevar, somentor: prof. dr. Branko Širok); dne 15. septembra 2014: Primož BERNJAK z naslovom: »Primerjava procesov razvlaževanja zraka v farmaciji« (mentor: prof. dr. Iztok Golobič); Blaž CERK z naslovom: »Eksperimentalna proga za merjenje prenosnika toplote s toplotnimi cevmi« (mentor: prof. dr. Iztok Golobič);

Rok DROBIČ z naslovom: »Krmiljenje avtomatske pralne steze za osebna vozila« (mentor: prof. dr. Janez Diaci); Anja JURIŠEVIČ z naslovom: »Daljinsko voden model mobilnega portalnega žerjava« (mentor: prof. dr. Janez Diaci, somentor: doc. dr. Boris Jerman); Matjaž MEGLEN z naslovom: »Izboljšano razvlaževanje zraka z lamelnim prenosnikom toplote s toplotnimi cevmi« (mentor: prof. dr. Iztok Golobič); Boštjan PAJK z naslovom: »Eksperimentalna proga za testiranje fizikalne priprave procesne vode« (mentor: prof. dr. Iztok Golobič); Ožbej DRMOTA z naslovom: »Planiranje leta enomotornega letala med Evropo in Kanado« (mentor: izr. prof. dr. Tadej Kosel); Matic LENARŠIČ z naslovom: »Aktivna kontrola mejnega sloja zraka na krilu« (mentor: izr. prof. dr. Tadej Kosel); Matej LUHENI z naslovom: »Zasnova lahkega proporcionalnega potnega ventila za mobilno hidravliko« (mentor: doc. dr. Franc Majdič); Žiga ŠTEKOVIĆ z naslovom: »Primerjalna analiza helikopterjev za operacije helikopterske nujne medicinske pomoči« (mentor: izr. prof. dr. Tadej Kosel); Matej JEVŠČEK z naslovom: »Povezava robota SCARA z virtualnim programskim okoljem V-REP« (mentor: prof. dr. Peter Butala); Tilen ROJKO z naslovom: »Tehnologija izdelave jeklenega stola« (mentor: doc. dr. Joško Valentinčič, somentor: prof. dr. Janez Tušek). * Na Fakulteti za strojništvo Univerze v Mariboru so pridobili naziv diplomirani inženir strojništva (VS): dne 4. septembra 2014: Blaž BALAŽIC z naslovom: »Določevanje dinamične trdnosti poliamida PA6E« (mentor: prof. dr. Srečko Glodež, somentor: doc. dr. Tomaž Vuherer); Mitja KOZEL z naslovom: »Načrtovanje in numerična analiza stojala za shranjevanje poviškov« (mentor: izr. prof. dr. Miran Ulbin, somentor: dr. Matej Borovinšek); dne 8. septembra 2014: Primož FILIPIČ z naslovom: »Uporaba koncepta vitke proizvodnje v področju vzdrževanja« (mentor: doc. dr. Marjan Leber, somentor: izr. prof. dr. Igor Drstvenšek); Primož VINKOVIČ z naslovom: »Konstruiranje polosi iz ogljikovega kompozita za Formulo S« (mentor: prof. dr. Srečko Glodež, somentor: izr. prof. dr. Jožef Predan); dne 16. septembra 2014: SI 129

Strojniški vestnik - Journal of Mechanical Engineering 60(2014)10, SI 124-130

Danijel BUČEK z naslovom: »Oblikovanje akumulacijskega transporterja v polnilni liniji brezalkoholnih pijač« (mentor: prof. dr. Iztok Potrč, somentor: izr. prof. dr. Tone Lerher); dne 18. septembra 2014: Rok BRATUŠA z naslovom: »Modifikacija krpljev, da iz njih nastanejo smuči« (mentor: izr. prof. dr. Stanislav Pehan); Eva KOCBEK z naslovom: »Numerična analiza obratovalnih karakteristik mešala za nevtralizacijo odpadnih vod iz pralnice perila« (mentor: doc. dr. Ignacijo Biluš, somentor: dr. Matej Zadravec); Domen KRAJNC z naslovom: »Računalniško podprta proizvodnja« (mentor: prof. dr. Jože Balič, somentor: doc. dr. Mirko Fick); Rastko LORGER z naslovom: »Izbira in dimenzioniranje mikro hidroelektrarne na potoku Sapočnica« (mentor: doc. dr. Ignacijo Biluš, somentor: prof. dr. Brane Širok); Amadej OPARENOVIĆ z naslovom: »Točke vpetja šesttočkovnega varnostnega pasu« (mentor: izr. prof. dr. Jožef Predan, somentor: prof. dr. Nenad Gubeljak); Primož SMOGAVC z naslovom: »Izgradnja sistema preventivnega vzdrževanja nove talilne peči v družbi Impol LLT d.o.o.« (mentor: doc. dr. Marjan Leber); Kristjan ŠPINDLER z naslovom: »Konstruiranje prehodne komore« (mentor: doc. dr. Aleš Belšak, somentor: izr. prof. dr. Miran Ulbin); Primož VERONIK z naslovom: »Konstruiranje manipulatorja za dviganje plošč« (mentor: prof. dr. Srečko Glodež, somentorja: izr. prof. dr. Stanislav Pehan, doc. dr. Marjan Leber); dne 25. septembra 2014: Sebastjan ČREŠNAR z naslovom: »Uvedba optičnega sistema na kontrolni liniji za sortiranje ojnic« (mentor: prof. dr. Bojan Ačko);

SI 130

Branko JAVERNIK z naslovom: »Finalizacija preoblikovalnih orodij za istosmerno iztiskavanje Alzlitin« (mentor: izr. prof. dr. Ivan Pahole, somentor: dr. Tomaž Brajlih); Branko LESNIK z naslovom: »Uporaba kompaktnih hidravličnih agregatov v strojegradnji« (mentor: izr. prof. dr. Darko Lovrec, somentor: dr. Vito Tič); Matic POLAK z naslovom: »Optimiranje parametrov regulatorja podajalnega pogona obdelovalnega stroja v Simulink okolju« (mentor: doc. dr. Uroš Župerl); Miha SRT z naslovom: »Analiza obstoječega stanja in postavitev nove linije za izdelavo aparatov« (mentor: doc. dr. Nataša Vujica Herzog); Matjaž ŠKORJANC z naslovom: »Analiza tehnologije izdelave orodja za steklenico« (mentor: doc. dr. Ficko Mirko); Tomaž VOGEL z naslovom: »Analiza gonila prečnega transporterja gredic« (mentor: doc. dr. Aleš Belšak, somentor: izr. prof. dr. Ulbin Miran). * Na Fakulteti za strojništvo Univerze v Mariboru je pridobil naziv diplomirani inženir mehatronike (VS): dne 4. septembra 2014: Tadej PERŠAK z naslovom: »Visokotlačni vbrizgalni sistem za polnjenje pijač z vsebnostjo CO2« (mentor: izr. prof. dr. Karl Gotlih, somentor doc. dr. Suzana Uran); dne 5. septembra 2014: Tadej STENIČNIK z naslovom: »Merilnik upogibne trdnosti brusnih etalonov« (mentor: izr. prof. dr. Karl Gotlih, somentor doc. dr. Andreja Rojko); dne 25. septembra 2014: Boštjan NAMAR z naslovom: »Gradnja 3D mikrostruktur« (mentor: izr. prof. dr. Karl Gotlih, somentor prof. dr. Riko Šafarič).

Technical Editor Pika Škraba University of Ljubljana, Faculty of Mechanical Engineering, Slovenia

Founding Editor Bojan Kraut

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 photos have been taken in the framework of the European project RES-e Regions (Renewable Energy Sources – Electricity) in which the Faculty of Mechanical Engineering participated. The task was a secondary school competition for the best photography on the subject of renewable energy sources in Slovenia. The co-ordinator of the project was the Upper-Austrian Energy Agency, from Linz. Courtesy: University of Ljubljana, Faculty of Mechanical Engineering, Slovenia.

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. We would like to thank the reviewers who have taken part in the peerreview process.

http://www.sv-jme.eu

60 (2014) 10

Since 1955

Papers

619

Uroš Stritih, Andrej Bombač: Description and Analysis of Adsorption Heat Storage Device

629

Rita Ambu, Andrea Manuello Bertetto, Costantino Falchi: Design of a Prototype System Operant in Lunar Environment

638

Xin Jin, Hua Liu, Wenbin Ju: Wind Turbine Seismic Load Analysis Based on Numerical Calculation

649

Srečko Glodež, Marko Šori, Tomaž Verlak: A Computational Model for Bending Fatigue Analyses of Sintered Gears

656

665

Liu Lei, Zhang Desheng, Zhao Jiyun: Design and Research for the Water Low-pressure Large-flow Pilot-operated Solenoid Valve

675

Oğuz Çolak: Optimization of Machining Performance in High-Pressure Assisted Turning of Ti6Al4V Alloy

Journal of Mechanical Engineering - Strojniški vestnik

Contents

10 year 2014 volume 60 no.

Strojniški vestnik Journal of Mechanical Engineering