machine design Vol.5(2013) No.1 ISSN

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1 UNIVERSITY OF NOVI SAD FACULTY OF TECHNICAL SCIENCES ADEKO ASSOCIATION FOR DESIGN, ELEMENTS AND CONSTRUCTIONS machine design Vol.5(2013) No.1 ISSN editor IN CHIEF: prof. phd. siniša kuzmanović novi sad, 2013

2 Publication Machine Design editor IN CHIEF Prof. Siniša KUZMANOVIĆ, Ph.D., University of Novi Sad, Faculty of Technical Sciences Publisher University of Novi Sad, Faculty of Technical Sciences, Trg Dositeja Obradovića 6, Novi Sad, Serbia Supported by ADEKO, Association for Design, Elements and Constructions CEEPUS III RS0304; CEEPUS III PL0033; CEEPUS III BG0703 Printed by Faculty of Technical Sciences, Graphic Center GRID, Novi Sad, Serbia technical preparation and cover design Eng. Milan RACKOV, MSc., University of Novi Sad, Faculty of Technical Sciences Electronic version of journal available on journal Frequency Four issues per year machine design is covered by the following indexes INDEX COPERNICUS JOURNAL MASTER LIST ( DOAJ Directory of Open Access Journals ( CIP Каталогизација у публикацији Библиотека Матице српске, Нови Сад 62-11: MACHINE Design / editor in chief Siniša Kuzmanović Novi Sad : University of Novi Sad, Faculty of Technical Sciences, cm Тромесечно. ISSN COBISS.SR-ID

3 SCIENTIFIC ADVISORY board Prof. Carmen ALIC, Ph.D. Prof. Miodrag JANKOVIĆ, Ph.D. Prof. Peter NENOV, Ph.D. University Politehnica Timisoara, Faculty of Engineering Hunedoara, Hunedoara, Romania University of Belgrade, Faculty of Mechanical Engineering, Belgrade, Serbia "Angel Kanchev" University of Rousse, Faculty of Transport Engineering Rousse, Bulgaria Prof. Kyrill ARNAUDOW, Ph.D. Prof. Dragoslav JANOŠEVIĆ, Ph.D. Prof. Milosav OGNJANOVIĆ, Ph.D. Bulgarian Academy of Sciences, Sofia, Bulgaria University of Niš, Faculty of Mechanical Engineering, Niš, Serbia University of Belgrade, Faculty of Mechanical Engineering, Belgrade, Serbia Prof. Livia Dana BEJU, Ph.D. Prof. Juliana JAVOROVA, Ph.D. Prof. Zoran PANDILOV, Ph.D. "Lucian Blaga" University of Sibiu, Engineering Faculty, Sibiu, Romania University of Chemical Technology and Metallurgy, Deptartment of Applied Mechanics, Sofia, Bulgaria Ss. Cyril and Methodius University, Faculty of Mechanical Engineering, Skopje, Macedonia Prof. Jaroslav BERAN, Ph.D. Prof. Miomir JOVANOVIĆ, Ph.D. Prof. Jose I. PEDRERO, Ph.D. Technical University of Liberec, Faculty of Mechanical Engineering, Liberec, Czech Republic University of Niš, Faculty of Mechanical Engineering, Niš, Serbia UNED, Departamento de Mecanica, Madrid, Spain Prof. Ilare BORDEAŞU, Ph.D. Prof. Svetislav JOVIČIĆ, Ph.D. Prof. Victor E. STARZHINSKY, Ph.D. Politehnica University of Timisoara, Faculty of Mechanical Engineering, Timisoara, Romania University of Kragujevac, Faculty of Mechanical Engineering, Kragujevac, Serbia V.A. Belyi Metal-Polymer Research Institute of National Academy of Sciences of Belarus, Gomel, Belarus Prof. Juraj BUKOVECZKY, Ph.D. Prof. Imre KISS, Ph.D. Prof. Slobodan TANASIJEVIĆ, Ph.D. Slovak University of Technology, Faculty of Mechanical Engineering, Bratislava, Slovakia University Politehnica Timisoara, Faculty of Engineering Hunedoara, Hunedoara, Romania University of Kragujevac, Faculty of Mechanical Engineering, Kragujevac, Serbia Prof. Radoš BULATOVIĆ, Ph.D. Prof. Kosta KRSMANOVIĆ, Ph.D. Prof. Wiktor TARANENKO, Ph.D. University of Montenegro, Faculty of Mechanical Engineering, Podgorica, Montenegro University of Arts in Belgrade, Faculty of Applied Arts, Belgrade, Serbia Lublin University of Technology, Institute of Technological Systems of Information, Lublin, Poland Prof. Ilija ĆOSIĆ, Ph.D. Prof. Sergey A. LAGUTIN, Ph.D. Prof. Marián TOLNAY, Ph.D. University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia Chief Expert on Gears, Design and Technology, JS Co EZTM, Electrostal, Moscow, Russia Slovak University of Technology, Faculty of Mechanical Engineering, Bratislava, Slovakia Prof. Lubomir DIMITROV, Ph.D. Prof. Tihomir LATINOVIĆ, Ph.D. Prof. Radoslav TOMOVIĆ, Ph.D. Technical University of Sofia, Faculty of Mechanical Engineering, Sofia, Bulgaria University of Banja Luka, Faculty of Mechanical Engineering, Banja Luka, Bosnia and Herzegovina University of Montenegro, Faculty of Mechanical Engineering, Podgorica, Montenegro Prof. Mircea-Viorel DRAGOI, Ph.D. Prof. Stanislaw LEGUTKO, Ph.D. Prof. Radivoje TOPIĆ, Ph.D. "Transilvania" University of Brasov, Faculty of Technological Engineering and Industrial Management, Brasov, Romania Poznan University of Technology, Institute of Mechanical Technology, Poznan, Poland University of Belgrade, Faculty of Mechanical Engineering, Belgrade, Serbia Prof. Vlastimir ĐOKIĆ, Ph.D. Prof. Imrich LUKOVICS, Ph.D. Prof. Andrei TUDOR, Ph.D. University of Niš, Faculty of Mechanical Engineering, Niš, Serbia Tomas Bata University in Zlin, Faculty of Technology, Zlin, Czech Republic University POLITEHNICA of Bucharest, Faculty of Mechanical Engineering and Mechatronic, Bucharest, Romania Prof. Milosav ĐURĐEVIĆ, Ph.D. Prof. Nenad MARJANOVIĆ, Ph.D. Prof. Lucian TUDOSE, Ph.D. University of Banja Luka, Faculty of Mechanical Engineering, Banja Luka, Bosnia and Herzegovina University of Kragujevac, Faculty of Mechanical Engineering, Kragujevac, Serbia Technical University of Cluj-Napoca, Faculty of Machine Building, Cluj-Napoca, Romania Prof. Dezso GERGELY, Ph.D. Prof. Zoran MARINKOVIĆ, Ph.D. Prof. Krasimir TUJAROV, Ph.D. University College of Nyíregyháza Faculty of Engineering and Agriculture Nyíregyháza, Hungary University of Niš, Faculty of Mechanical Engineering, Niš, Serbia Angel Kunchev University of Rousse, Faculty of Agricultural Mechanisation, Department of Thermotehnics, Hydro- and Pneumotechnics, Rousse, Bulgaria Prof. Veniamin GOLDFARB, Ph.D. Prof. Štefan MEDVECKY, Ph.D. Prof. Karol VELISEK, Ph.D. Izhevsk State Technical University, Institute of Mechanics, Izhevsk, Russia University of Žilina Faculty of Mechanical Engineering, Žilina, Slovakia Slovak University of Technology in Bratislava, Faculty of Materials Science and Technology STU, Trnava, Slovakia Prof. Ladislav GULAN, Ph.D. Prof. Athanassios MIHAILIDIS, Ph.D. Prof. Miroslav VEREŠ, Ph.D. Slovak University of Technology, Faculty of Mechanical Engineering, Bratislava, Slovakia Aristotle University of Thessaloniki, Faculty of Engineering, Lab. of Machine Elements & Machine Design, Thessaloniki, Greece Slovak University of Technology, Faculty of Mechanical Engineering, Bratislava, Slovakia Prof. Csaba GYENGE, Ph.D. Prof. Vojislav MILTENOVIĆ, Ph.D. Prof. Simon VILMOS, Ph.D. Technical University of Cluj-Napoca, Faculty of Machine Building, Cluj-Napoca, Romania University of Niš, Faculty of Mechanical Engineering, Niš, Serbia Budapest University of Technology and Economics, Department of Machine and Product Design, Budapest, Hungary Prof. Sava IANICI, Ph.D. Prof. Radivoje MITROVIĆ, Ph.D. Prof. Jovan VLADIĆ, Ph.D. Eftemie Murgu University of Resita, Faculty of Engineering, Resita, Romania University of Belgrade, Faculty of Mechanical Engineering, Belgrade, Serbia University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia Prof. Milan IKONIĆ, Ph.D. Prof. Slobodan NAVALUŠIĆ, Ph.D. Prof. Miodrag ZLOKOLICA, Ph.D. University of Rijeka, Faculty of Engineering, Rijeka, Croatia University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia

4 editorial review board Prof. Milosav ĐURĐEVIĆ, Ph.D. Prof. Dragoslav JANOŠEVIĆ, Ph.D. Prof. Dragoljub NOVAKOVIĆ, Ph.D. University of Banja Luka, Faculty of Mechanical Engineering, Banja Luka, Bosnia and Herzegovina University of Niš, Faculty of Mechanical Engineering, Niš, Serbia University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia Prof. Slobodan GAJIN, Ph.D. Prof. Juliana JAVOROVA, Ph.D. Prof. Milosav OGNJANOVIĆ, Ph.D. University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia University of Chemical Technology and Metallurgy, Deptartment of Applied Mechanics, Sofia, Bulgaria University of Belgrade, Faculty of Mechanical Engineering, Belgrade, Serbia Prof. Milosav GEORGIJEVIĆ, Ph.D. Prof. Imre KISS, Ph.D. Prof. Rastislav ŠOSTAKOV, Ph.D. University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia University Politehnica Timisoara, Faculty of Engineering Hunedoara, Hunedoara, Romania University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia Prof. Katarina GERIĆ, Ph.D. Prof. Kosta KRSMANOVIĆ, Ph.D. Prof. Radivoje TOPIĆ, Ph.D. University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia University of Arts in Belgrade, Faculty of Applied Arts, Belgrade, Serbia University of Belgrade, Faculty of Mechanical Engineering, Belgrade, Serbia Prof. Valentin GLAVARDANOV, Ph.D. Prof. Tihomir LATINOVIĆ, Ph.D. Prof. Andrei TUDOR, Ph.D. University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia University of Banja Luka, Faculty of Mechanical Engineering, Banja Luka, Bosnia and Herzegovina University POLITEHNICA of Bucharest, Faculty of Mechanical Engineering and Mechatronic, Bucharest, Romania Prof. Miroslav GOJO, Ph.D. Prof. Stanislaw LEGUTKO, Ph.D. Prof. Krasimir TUJAROV, Ph.D. University of Zagreb, Faculty of Graphic Arts, Zagreb, Croatia Poznan University of Technology, Institute of Mechanical Technology, Poznan, Poland Angel Kunchev University of Rousse, Faculty of Agricultural Mechanisation, Department of Thermotehnics, Hydro- and Pneumotechnics, Rousse, Bulgaria Prof. Ladislav GULAN, Ph.D. Prof. Ratko MARETIĆ, Ph.D. Prof. Karol VELISEK, Ph.D. Slovak University of Technology, Faculty of Mechanical Engineering, Bratislava, Slovakia University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia Slovak University of Technology in Bratislava, Faculty of Materials Science and Technology STU, Trnava, Slovakia Prof. Gorazd HLEBANJA, Ph.D. Prof. Athanassios MIHAILIDIS, Ph.D. Prof. Miroslav VEREŠ, Ph.D. University of Ljubljana, Faculty of Mechanical Engineering, Ljubljana, Slovenia Aristotle University of Thessaloniki, Faculty of Engineering, Lab. of Machine Elements & Machine Design, Thessaloniki, Greece Slovak University of Technology, Faculty of Mechanical Engineering, Bratislava, Slovakia Prof. Sava IANICI, Ph.D. Prof. Vojislav MILTENOVIĆ, Ph.D. Prof. Simon VILMOS, Ph.D. Eftemie Murgu University of Resita, Faculty of Engineering, Resita, Romania University of Niš, Faculty of Mechanical Engineering, Niš, Serbia Budapest University of Technology and Economics, Department of Machine and Product Design, Budapest, Hungary Prof. Milan IKONIĆ, Ph.D. Prof. Slobodan NAVALUŠIĆ, Ph.D. Prof. Jovan VLADIĆ, Ph.D. University of Rijeka, Faculty of Engineering, Rijeka, Croatia University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia University of Novi Sad, Faculty of Technical Sciences, Novi Sad, Serbia All the publications in this journal have the authorship, whereas the authors of the papers carry entire responsibility for originality and content. The use of some items or complete papers is permitted only if the source is given.

5 from the editor The journal Machine Design comes up to the fifth volume for 2013 and continues to be published four times a year. This kind of formatting and editing certainly bring a high level of journal quality which will satisfy most of readers. The journal receives more and more interesting papers for review. Machine Design publishes fundamental research about mechanical engineering and design including machine elements, design fundamentals, computer aided design, product forms, shapes and performances, manufacturing processes and technologies, theory of materials, its structures and capabilities, product design management, technology management, communication and cognitive science. The journal is a good opportunity to show and present the results of our recent work and researching. Also, it is a chance for leader researchers and scientists in the field of machine design from abroad to represent their researching results. In such way, we would like to obtain insight in the present situation of mechanical engineering in the region, to know and learn about researching in other institutions, to compare results and find out new solutions, as well as to make new contacts and find out mutual interests for international cooperation and researching on a project or some topic. Machine Design is on the Index Copernicus international journals master list and on DOAJ Directory of Open Access Journals. Its editorial board will try further to develop this publication in order to achieve and maintain a high quality of publications, so we can receive an Impact factor. Our goals are to be referred in international publication databases, to provide an international medium for scientific contribution and participation to mechanical engineers and to create a platform for the communication between science and industry in the field of technical sciences. Also, we would like to promote and to encourage international cooperation, mutual researching, projects and publishing papers between foreign partners institutions. Thus, we want to help better understanding and knowing about work and researching of colleagues from all over the world. I hope You will recognize the interest to publish Your paper in the journal Machine Design; so, with a great pleasure, I call You to send further Your papers for this journal. At the end of the journal we gave the instructions for formatting and preparing the paper. For additional information, please visit our website: Editor in Chief, Prof. Ph.D. Siniša Kuzmanović

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7 CONTENTS: Original scientific papers 1. Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel By GMAW P. SREERAJ, T. KANNAN, SUBHASISMAJI Proposed Model for Cavitation Erosion Test Results Presentation Mircea Octavian POPOVICIU, Ilare BORDEASU Studies and Experimental Research on the Behavior in Exploitation of the Beam Ensemble of the Ingot Handling Cranes in Siderurgical Metallic Workshop Cristina Carmen MIKLOS, Imre Zsolt MIKLOS, Carmen Inge ALIC The Impact of Bearing Deformation in the Field of Pressure and Its Hydrodinamic Characteristics Koço BODE, Odhisea KOÇA, Ilirian KONOMI Preliminary notes 5. Design Analysis Algorithm of a Petrol Engine-Powered Air Compressor Emmanuel SIMOLOWO, Taiwo OLUMIDE The Influence of the Injection Processing Temperature on the Thermal Stability of Polymers Used in the Manufacture of Items in the Automotive and Sports Industry Gheorghe Radu Emil MĂRIEŞ Research papers 7. Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach Bhaveshkumar P. PATEL, Jagdish M. PRAJAPATI Dimensional Experimental and Finite Element Stress Analysis Of C.I. Wedge of Sluice Valve Narayan DHARASHIVAKAR, Prashant PATIL, Krishnakumar JOSHI Simple Clutches with Multiple Functions Elena EFTIMIE MANUSCRIPT FORMAT

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9 machine design, Vol.5(2013) No.1, ISSN pp Original scientific paper SENSITIVITY ANALYSIS OF PROCESS PARAMETERS IN CLADDING OF STAINLESS STEEL BY GMAW P. SREERAJ 1, * - T. KANNAN 2 - SUBHASISMAJI 3 1 Department of Mechanical Engineering, Valia Koonambaikulathamma College of Engineering and Technology, Kerala, India 2 SVS College of Engineering,Coimbatore,Tamilnadu, India 3 Department of Mechanical Engineering IGNOU, Delhi,110068, India Received ( ); Revised ( ); Accepted ( ) Abstract: To improve the corrosion resistant properties of carbon steel usually cladding process is used. It is a process of depositing a thick layer of corrosion resistant material over carbon steel plate. Most of the engineering applications require high strength and corrosion resistant materials for long term reliability and performance. By cladding these properties can be achieved with minimum cost. The main problem faced on cladding is the selection of optimum combinations of process parameters for achieving quality clad and hence good clad bead geometry. In this study Sensitivity analysis was performed to identify various input process parameters (welding current, welding speed, gun angle, contact tip to work distance and pinch) exerting most influence in stainless steel cladding of low carbon structural steel plates using Gas Metal Arc Welding (GMAW) and the bead geometry. Experiments were conducted based on central composite rotatable design with full replication technique and mathematical models were developed using multiple regression method. The developed models have been checked for adequacy and significance. Studies reveal that a change in process parameters affects bead geometry. Key words: Mathematical model, cladding, GMAW, Sensitivity Analysis, Clad bead geometry, corrosion. 1. INTRODUCTION Prevention of corrosion is a major problem in Industries. Even though it cannot be eliminated completely it can be reduced to some extent. A corrosion resistant protective layer is made over the less corrosion resistant substrate by a process called cladding. This technique is used to improve life of engineering components but also reduce their cost. This process is mainly now a day s used in industries such as chemical, textiles, nuclear, steam power plants, food processing and petro chemical industries [1]. Most accepted method of employed in weld cladding is GMAW. It has got the following advantages [2]. High reliability All position capability Ease to use Low cost High Productivity Suitable for both ferrous and non-ferrous metals High deposition rate Absences of fluxes Cleanliness and ease of mechanization The mechanical strength of clad metal is highly influenced by the composition of metal but also by clad bead shape. This is an indication of bead geometry. Fig 1 shows the clad bead geometry. It mainly depends on wire feed rate, welding speed, arc voltage etc. Therefore it is necessary to study the relationship between the process parameters and bead parameters to study clad bead geometry. Using mathematical models it can be achieved. This paper highlights the study carried out to develop mathematical and Sensitivity models to optimize clad bead geometry, in stainless steel cladding deposited by GMAW. The experiments were conducted based on four factor five level central composite rotatable designs with full replication technique [3]. The developed models have been checked for their adequacy and significance. Again sensitivity analysis was performed to identify various input parameters exerting influence on the bead parameters. Percentage dilution (D) = [B/ (A+B)] X 100 Fig 1. Clad bead geometry 2. EXPERIMENTAL PROCEDURE The following machines and consumables were used for the purpose of conducting experiments. 1) A constant current gas metal arc welding machine (Invrtee V 350 PRO advanced processor with amps output range) 2) Welding manipulator. 3) Wire feeder (LF 74 Model). *Correspondence Author s Address: Department of Mechanical Engineering, Valia Koonambaikulathamma College of Engineering and Technology, Kerala, India

10 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp ) Filler material Stainless Steel wire of 1.2mm diameter (ER 308 L). 5) Gas cylinder containing a mixture of 98% argon and 2% of oxygen. 6) Mild steel plate (grade IS 2062). Test plates of size 300 x 200 x 20mm were cut from mild steel plate of grade IS 2062 and one of the surfaces is cleaned to remove oxide and dirt before cladding. ER-308 L stainless steel wire of 1.2mm diameter was used for depositing the clad beads through the feeder. Argon gas at a constant flow rate of 16 litres per minute was used for shielding. The properties of base metal and filler wire are shown in Table 1. The important and most difficult parameter found from trial run is wire feed rate. The wire feed rate is proportional to current. Wire feed rate must be greater than critical wire feed rate to achieve pulsed metal transfer. The relationship found from trial run is shown in equation (1). The formula derived is shown in Fig 2. Wire feed rate = *current (1) The selection of the welding electrode wire based on the matching the mechanical properties and physical characteristics of the base metal and weld size [4]. A candidate material for cladding which has excellent corrosion resistance and weldability is stainless steel. These have chloride stress corrosion cracking resistance and strength significantly greater than other materials. These have good surface appearance, good radiographic standard quality and minimum electrode wastage. Experimental design procedure used for this study is shown in Fig 3 and importance steps are briefly explained. Table 1. Chemical Composition of Base Metal and Filler Wire Elements, Weight % Material C SI Mn P S Al Cr Mo Ni s IS ER308L Fig.2. Relationship between Current and Wire Feed Rate 3. PLAN OF INVESTIGATION The research work was planned to be carried out in the following steps [5]: 1) Identification of factors and responses. 2) Finding limits of process variables. 3) Development of design matrix. 4) Conducting experiments as per design matrix. 5) Recording the responses. 6) Development of mathematical models. 7) Checking the adequacy of developed models. 8) Conducting conformity tests. Identification of factors and responses Finding the limits of process variables Development of design matrix Conducting experiments as per design matrix Recording the responses Development of mathematical models Checking adequacy of developed models Conducting conformity Fig.3. Experimental design procedure 3.1. Identification of factors and responses The following independently controllable process parameters were found to be affecting output parameters. These are wire feed rate (W), welding speed (S), welding gun angle (T), contact tip to work to distance (N) and pinch (Ac), The responses chosen were clad bead width (W), height of reinforcement (R), Depth of Penetration. (P) and percentage of dilution (D). The responses were chosen based on the impact of parameters on final composite model. The basic difference between welding and cladding is the percentage of dilution. The properties of the cladding are significantly influenced by dilution obtained. Hence control of dilution is important in cladding where a low dilution is highly desirable. When dilution is quite low, the final deposit composition will be closer to that of filler material and hence corrosion resistant properties of cladding will be greatly improved. The chosen factors have been selected on the basis to get minimal dilution and optimal clad bead geometry. 2

11 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp Few significant research works have been conducted in these areas using these process parameters and so these parameters were used for experimental study Finding the limits of process variables Working ranges of all selected factors are fixed by conducting trial runs. This was carried out by varying one of factors while keeping the rest of them as constant values. At each run settings for all parameters were disturbed and reset for next deposit. This is very essential to introduce variability caused by errors in experimental set up. Table 2. Welding Parameters and their Levels Parameters Unit Notation Factor Levels Welding Current A Welding Speed mm/min S Contact tip to work distance mm N Welding gun Angle Degree T Pinch - Ac Parameters were decided upon by inspecting the bead for smooth appearance without any visible defects. The upper limit of given factor was coded as -2. The coded value of intermediate values was calculated using the equation (2). = (2) Where X i is the required coded value of parameter X is any value of parameter from X min X max. X min is the lower limit of parameters and X max is the upper limit parameters [4]. The chosen level of the parameters with their units and notation are given in Table Development of design matrix Design matrix chosen to conduct the experiments was central composite rotatable design. The design matrix comprises of full replication of 2 5 (= 32), Factorial designs. All welding parameters in the intermediate levels (o) Constitute the central points and combination of each welding parameters at either is highest value (+2) or lowest value (-2) with other parameters of intermediate levels (0) constitute star points. 32 experimental trails were conducted that make the estimation of linear quadratic and two way interactive effects of process parameters on clad geometry [5] Conducting experiments as per design matrix The experiments were conducted at SVS College of Engineering, Coimbatore,India. In this work, thirty two experimental runs were allowed for the estimation of linear quadratic and two-way interactive effects of 3.5. Recording of Responses In order to measure clad bead geometry of transverse section of each weld overlays were cut using band saw from mid length. Position of the weld and end faces were machined and grinded. The specimen and faces were polished and etched using a 5% nital solution to display bead dimensions. The clad bead profiles were traced using a reflective type optical profile projector at a magnification of X10, in M/s Roots Industries Ltd. Coimbatore. Then the bead dimension such as depth of penetration height of reinforcement and clad bead width were measured [6]. The traced bead profiles were scanned in order to find various clad parameters and the percentage of dilution with help of AUTO CAD software. This is shown in Fig 4. Fig.4. Traced Profiles (Specimen No.2) 02A 02B 02A represents profile of the specimen (front side) and 02B represents profile of the specimen (rear side). The measured clad bead geometry is shown in Table 3. 3

12 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp Table 3. Design Matrix and Observed Values of Clad Bead Geometry Trial No. Design Matrix Bead Parameters I S N T Ac W (mm) P (mm) R (mm) D (%) W - Width; P - Penetration; R - Reinforcement; D - Dilution % 4

13 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp Development of Mathematical Models The response function representing any of the clad bead geometry can be expressed as [7, 8, 9], Y = f (A, B, C, D, E) (3) where, Y = Response variable A = Welding current (I) in amps B = Welding speed (S) in mm/min C = Contact tip to Work distance (N) in mm D = Welding gun angle (T) in degrees E = Pinch (Ac) The second order surface response model equals can be expressed as below Y = β 0 + β 1 A + β 2 B + β 3 C + β 4 D + β 5 E + β 11 A 2 + β 22 B 2 + β 33 C 2 + β 44 D 2 + β 55 E 2 + β 12 AB + β 13 AC + β 14 AD + β 15 AE + β 23 BC + β 24 BD + β 25 BE + β 34 CD + β 35 CE+ β 45 DE (4) Where, β 0 is the free term of the regression equation, the coefficient β 1,β 2,β 3,β 4 andβ 5 is are linear terms, the coefficients β 11,β 22, β 33,β 44 andß 55 quadratic terms, and the coefficients β 12,β 13,β 14,β 15, etc are the interaction terms. The coefficients were calculated using Quality America six sigma software (DOE PC IV). After determining the coefficients, the mathematical models were developed. The developed mathematical models are given as follows. Clad Bead Width (W), mm = A B C D E 0.423A B C D E AB AC AD AE 0.134BC BD BE CD CE DE. (5) Depth of Penetration (P), mm = A 0.032B C 0.032D 0.008E 0.124A B C D E AB AC AD AE 0.018BC BD BE CD CE 0.036DE. (6) Height of Reinforcement (R), mm = A 0.151B 0.060C D 0.002E A B C D E AB AC 0.008AD 0.048AE 0.024BC 0.062BD 0.003BE CD 0.092CE 0.095DE. (7) Percentage Dilution (D), % = A B C 0.039D 0.153E 1.324A B C D E AB AC AD AE BC BD BE CD CE DE. (8) 3.7. Checking the adequacy of the developed models The adequacy of the developed model was tested using the analysis of variance (ANOVA) technique. As per this technique, if the F ratio values of the developed models do not exceed the standard tabulated values for a desired level of confidence (95%) and the calculated R ratio values of the developed model exceed the standard values for a desired level of confidence (95%) then the models are said to be adequate within the confidence limit [10]. These conditions were satisfied for the developed models. The values are shown in Table 4. Table 4. Analysis of variance for Testing Adequacy of the Model Parameter 1 st Order terms 2 nd order terms Lack of fit Error terms SS DF SS DF SS DF SS DF F-ratio R-ratio Whether model is adequate W Adequate P Adequate R Adequate D Adequate SS - Sum of squares; DF - Degree of freedom; F Ratio (6, 5, 0.5) = ; R Ratio (20,5,0.05) = SENSITIVITY ANALYSIS FOR BEAD GEOMETRY From the above developed mathematical equations 5-8 to be used for the estimation of bead geometry, sensitivity equations obtained by differentiating them with respect to the process parameters of interest such as welding current (I), welding speed (S), contact tip to work distance (N), welding gun angle (T) and pinch (Ac).Mathematical models shown in Equations W= *I *S+0.587*N *T *P *I*I 0.291*S*S *N*N *T*T * P*P I*S *I*N *I*P I*P Xp x5*N *T *P N*T N*P *P (9) 5

14 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp P = *I *S *N-0.032*T *P *I*I *S*S *N*N *T*T *P*P- 0.33*I*S *I*N *I*T *I*P *S*N *S*T *S*P *N*T N*P T*P (10) R= *I 0.151*S *N T *P *I*I+0.037*S*S *N*N *T*T-0.006*P*P I*S I*N *I*T *I*P *S*N *S*N *S*T S*P N*T N*P T*P (11) D = *I+ 0.34*S+3.141*N *T *P-1.324*I*I-0.925*S*S-1.012*N*N-1.371*T*T *P*P I*S I*N I*T I*P S*N+ 0.46TS*T *S*P *N*T N*P *T*P (12) From the above developed Mathematical equations 9-12 to be used for the estimating bead geometry. The sensitivity equations are obtained by differentiating the equations with respect to the process parameters of interest, Such as welding current (1), welding speed (S), Contact tip to work distance (N), welding gun angle (T), Pinch (Ac)[11]. The sensitivity equation for welding current obtained by differentiating equations 9-12 with respect to welding current(i) are given below: = *2*I * S *N *T *P (13) = *2*I 0.335*S *N *T *P (14) = *2*I * S *N *T * P (15) = *4*2*I 200*S *N *T * P (16) Sensitivity equation for welding speed were obtained by differentiating equations 9-12 with respect to welding speed(s) and are given below: = *2*S *S * N * T * P (17) = *2*S * I *N *T *P (18) = *2*S T*I *N *N * T *P (19) = *2*S *I * N *T * P (20) 6 The sensitivity equations for contact tip to work distance were obtained by differentiating equation 9-12 with respect to contact tip to work distance (N) are given below. = *2*N *I *S *T *P (21) = *2*N *I * S *T * P (22) = *2*N * I * S * T *P (23) = *N+0.001*I *P (24) The sensitivity equation for Welding gun angle were obtained by differentiating equations 9-12 with respect to Welding gun angle (N) are given below: = *2*T *I * S *N *P (25) = *2*T *I * S *N *P (26) = S *2*T * I *S *N * P (27) = *2*T *I *T* S * N * P (28) The sensitivity equations for Pinch were obtained by differentiating equations 9-12 with respect to pinch (Ac) given below: = *2 P * I *S *N * T (29) = *2*P * I *S * N *T (30) = *2*P * I *S *N * T (31) = *2*P I * S *N *T (32)

15 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp Table 5. Bead width sensitivity of process parameters Welding speed (S) =166, Welding gun angle (T) = 90, Pinch (Ac)= 0, contact tip to work distance (N) =18. Welding current (-2) (-1) (0) (1) Table 8. Dilution sensitivities of process parameters Welding speed(s) =166, Contact to tip distance (N) =18, Welding gun angle (T) =90, Pinch (Ac) =0 Welding current dd ds dd di dd dn dd dt dd dv (2) Table 6. Depth of penetration sensitivities of process parameters Welding speed (S) = 166, Contact tip to work distance (N)= 18, Welding gun angle(t) = 90, Pinch(Ac) = 0 Welding current sensitivity of welding current Welding current Bead Width Penetration Reinforcement Dilution Fig.5. Sensitivity Analysis of Welding current on Bead width, Penetration, Reinforcement, Dilution S=166, N=18, T=90, Ac=0 Table 7. Reinforcement sensitivities of process parameters Welding speed (S)= 166, Contact tip to work distance (N) =18,Welding gun angle(t)= 90, Pinch(Ac) = 0 Welding current Sensitivity of welding gun angle 1,5 1 0,5 0 0,5 1 1, Welding current Bead width Penetration Reinforcement Dilution Fig.6. Sensitivity Analysis of Welding gun angle on Bead width,penetration,reinforcement,dilution S=166,N=18,T=90,Ac=0 7

16 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp Sensitivity of welding speed 0,8 0,6 0,4 0,2 0 0,2 0,4 0,6 0, Welding current Bead width Fig.7. Sensitivity Analysis of Welding speed on Bead width, Penetration, Reinforcement, Dilution S=166, N=18, T=90, Ac=0 Sensitivity of bead width 3 2,5 2 1,5 1 0,5 0 0,5 1 1, Welding current Welding current Welding speed Contact to plate distance Welding gun angle Pinch Fig.10. Sensitivity of Bead width Sensitivity of contact to plate distance Welding current Bead width Penetration Reinforcement Dilution Fig.8. Sensitivity Analysis of contact to plate distance on Bead width, Penetration, Reinforcement, Dilution S=166, N=18, T=90, Ac=0 Sensitivity of penetration 0,8 0,6 0,4 0,2 0 0,2 0,4 0,6 0, Welding current Fig.11. Sensitivity of penetration Welding current Welding speed Contact to plate distance Welding gun angle Pinch Sensitivity of pinch 0,3 0,2 0,1 0 0,1 0,2 0, Welding current Bead width Penetration Reinforcement Dilution Sensitivity of reinforcement 3 2,5 2 1,5 1 0,5 0 0, Welding current Welding current Welding speed Nozzle to plate distance Welding gun angle Pinch Fig.9. Sensitivity Analysis of pinch on Bead width,penetration,reinforcement,dilutions=166, N=18, T=90,Ac=0 Fig.12. Sensitivity of Reinforcement 8

17 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp Sensitivity of dilution Welding current Welding speed Nozzle to plate distance Welding gun angle Pinch penetration becomes positive. With increase in welding angle the dilution is increasing gradually. The dilution is more sensitive than other parameters Sensitivity of pinch on bead geometry Fig 9 shows the sensitivity of pinch on. With the bead geometry increase in pinch bead width is increasing in a constant manner and it is more sensitive than other parameters. With the decreasing of pinch the dilution becomes negative. The penetration first decreases and then becomes constant throughout CONCLUSION 6 Welding current Fig.13. Sensitivity of dilution 5. RESULTS AND DISCUSSION 5.1. Sensitivities of welding current on bead geometry. They are presented in Table 5-8 and are shown in Fig 5-9.Fig 5 shows sensitivity of welding current on weld bead geometry viz bead width, penetration, reinforcement, dilution. Of all welding parameters dilution is more sensitive to welding current than others. When welding current is constantly increased beyond 200 amps; the dilution changes to negative value. When welding current increased; penetration goes on decreasing. It is interesting to note that bead width decreasing in regular manner with the increasing in welding current. With the increasing in welding current reinforcement becomes positive Sensitivity of welding speed on bead geometry Fig 7 depicts the sensitivity of welding speed on weld bead geometry. It is evident from the figure that dilution remains less sensitive to welding speed and it is almost constant at slight changes of welding speed. Reinforcement is gradually increasing to positive value when welding speed is increased Sensitivity of contact to plate distance on weld geometry Fig 8 shows the sensitivity of contact to plate distance on weld bead geometry; the percentage of dilution is more sensitive to contact to plate distance. Variation in contact to plate distance shows a considerable change in dilution. When contact to plate distance is increased bead width gradually increased. The penetration remains unchanged and reinforcement has no considerable effect on contact to plate distance Sensitivity of welding gun angle on weld bead geometry Fig 6 shows with increasing in welding gun angle the bead width becomes positive. The same happens with penetration also. With increase in welding gun angle Experiments were conducted using GMAW to produce cladding on austenitic stainless steel material. From the experimental results a mathematical model was developed using regression models, which were checked for their accuracy and found to be satisfactory. Sensitivity analysis was performed to identify process parameters exerting the most influence on bead geometry. The change in welding current has more significant effect on bead geometry than welding speed. Percentage of dilution can be easily controlled with minimal change in the value of N, while the other parameters are kept at desired level. Height of reinforcement is more sensitive to changes in welding speed than other process parameters.hence it is reasonable to control the welding speed to get desired reinforcement. Figures shows some typical sensitivity plots. From these figures it can be observed that sensitivity of penetration is more prominent for changes in welding current, whereas changes in sensitivity of dilution and bead width are prominent changes in contact to plate distances compared to changes in welding current and welding speed. ACKNOWLEDGEMENT The authors sincerely acknowledge the help and facilities extended to them by the department of mechanical engineering SVS college of Engineering, Coimbatore, Tamil Nadu, India. REFERENCES [1] Kazakcci Palani P K, Murugan N.(2006) Prediction of Delta ferrite Content and Effect of Welding Process Parameters in Claddings by FCAW Journal of Materials and Manufacturing Process ( 21:5pp [2] Kannan.T and Murugan.N.(2006) Prediction of ferrite number of duplex stainless steel clad metals using RSM - Welding Journal pp 91-s to 99-s [3] Gunaraj V. and Murugan N.(2005) Prediction and control of weld bead geometry and shape relationships in submerged arc welding of pipes - 9

18 P. Sreeraj, T. Kannan, Subhasismaji: Sensitivity Analysis of Process Parameters in Cladding of Stainless Steel by GMAW; Machine Design, Vol.5(2013) No.1, ISSN ; pp Journal of Material Processing Technology (2005) Vol. 168, pp [4] KimI.S, SonK.J, YangY.S,Yaragada P. K. D.V(2003), Sensitivity analysis for process parameters in GMA welding process using factorial design method - International Journal of Machine tools and Manufacture Vol. 43, pp. 763 to 769 [5] Cochran W.G and CoxzG.M. Experimental Design(1987) pp.370, New York, John Wiley & sons [6] SerdarKaraoglu,AbdullahSecgin(2008) Sensitivity analysis of submerged arc welding process parameters Journal of Material Processing Technology Vol-202, pp [7] Ghosh P.K, Gupta P.C. and Goyal V.K(1998). Stainless steel cladding of structural steel plate using the pulsed current GMAW processes- Welding Journal 77(7) pp.307-s-314-s. [8] Gunaraj V. and Murugan N.(1999) Prediction and comparison of the area of the heat effected zone for the bead on plate and bead on joint in SAW of pipes Journal of Material processing Technology. Vol. 95, pp. 246 to 261. [9] Montgomery DC Design and analysis of Experiments (2003) John Wiley & Sons (ASIA) Pvt. Ltd. [10] Palani P.K., Murugan.N.(2010) Sensitivity analysis of process parameters in cladding of stainless steel by Flux Cored Gas Metal Arc Welding. Journal of Manufacturing process.vol-8/no.2, pp [11] Praikshit Dutta, Dilip Kumar Pratihar(2007) Modelling TIG welding process using conventional regression analysis and neural network based approaches - Journal of material processing technology, 184,

19 machine design, Vol.5(2013) No.1, ISSN pp Original scientific paper PROPOSED MODEL FOR CAVITATION EROSION TEST RESULTS PRESENTATION Mircea Octavian POPOVICIU 1 - Ilare BORDEASU 2, * 1 Academy of Romanian Scientists, Timisoara, Romania 2 Polytechnic University of Timisoara, Timisoara, Romania Received ( ); Revised ( ); Accepted ( ) Abstract: Recently in Timisoara Hydraulic Machinery Laboratory was realized a cavitation erosion test facility respecting the new ASTM G32-10 Standard recommendations. An exhaustive comparison between the ASTM recommendations and the realized facility is given. It was also decided to respect the standard for conducting and presentation of tests results. The main purpose of this work is to obtain a model for the presentation of future test results. The selected test material was the stainless steel 12/8 with very good cavitation erosion behaviour. There were tested three specimens designated 1, 2 and 3. The arithmetic mean of the specimens 1-3, for each testing time, gives the value considered representative for the material. The results are compared with the stainless steel OH12NDL used in the past for manufacturing numerous hydraulic turbines. For plotting the results, different types of regression equations have been chosen such as polynomials and an exponentials. In addition to the characteristic curves, the ASTM G32-10 Standard recommends the presentation of some other important parameters, such as: total cumulative mass loss, cumulative mean depth of erosion, maximum rate of erosion nominal incubation time, erosion threshold time and the loss for extended exposure. For the tested material all these parameters are presented and their importance is discussed. Key words: cavitation erosion characteristic curves, vibratory cavitation apparatus, stainless steel, regression equations 1. INTRODUCTION For the new T2 cavitation erosion facility, with piezoelectric crystals, realized recently in the Timisoara Hydraulic Machinery Laboratory (THML) it was decided to change radically the results presentation, taking into account both the recommendations of the ASTM G32-10 Standard and the experience achieved in the past by using T1, the old magnetostrictive device, with nickel tube [2], [3], [6]. For T1 the obtained results were given as time dependence (with the time expressed in seconds) of mass or volume losses (mg or mm 3 ). In G-32 Standard is recommended the use of mean depth erosion (m) against time (expressed in hours). Evidently, the general aspect of the characteristic curves remains unchanged but the numerical data present great differences. For beginning it was chosen a material with very good but not excellent cavitation erosion behaviour, namely a stainless steel with 12% chromium and 8% Ni. To obtain the characteristic curves there were used polynomial and exponential equations. 2. TEST FACILITIES The test facility T2 can be seen in Fig. 1. The comparison of the T2 main parameters with those of the G-32 Standard are given in Table 1 [1]. The test specimen important data are compared in Table 2. Test facility 1-Horn, 2-Electronic system, 3-Temperature control system, liquid vessel and cooling coil, 5- Ventilation system Diagram of test facility 1-Piezoelectric transducer, 2-Ultrasonic generator, 2- Cooling system, 3-Liquid temperature control Fig.1. T2 vibratory device * Correspondence Author s Address: Polytechnica University of Timisoara, Faculty of Mechanical, Bvd. Mihai Viteazul no.1, Timisoara, Romania, ilarica59@gmail.com

20 Mircea Octavian Popoviciu, Ilare Bordeasu: Proposed Model for Cavitation Erosion Test Results Presentation; Machine Design, Vol.5(2013) No.1, ISSN ; pp Table 1. Test Facility Important Parameters No. M* Parameter MU G32 THML 1 Type - Piezoelectric Piezoelectric Magnetostrictive - Catenoidal - 2 Horn profile - Exponential Conical - Stepped - 3 Acoustic power W M Frequency khz M Peak to peak ampl. μm 50±5% 50±5% 6 Cavitation liquid - Distilled water Distilled water 7 M Depth of water mm 100± M Specimen immersion mm 12± M Liquid temperature ºC 25±2 23± M Air pressure mmhg 760±6% Amplitude control - Digital control 12 Liquid temp. control - Digital control 13 Cooling system Immersed External Immersed *Mandatory Table 2. Test Specimen Important Data No. M* Data MU G32 THML 1 M Diameter (D) mm 15.9± Thickness (H) mm M Chamfer or radius (E) mm M Working face perpendicularity on transducer axis mm M Radial run-out (r) mm Thread nom. diam. (T) mm M M Thread length (L) mm 10.0±1.0 10±1.0 8 Specimen mass (steel) g *M- mandatory 3. TESTED MATERIAL The tested material was obtained as 300 g test samples in a furnace with electron beams. The resulted chemical composition was determined with a Foundry Master spectrometer. The steel symbolized 12/8 has as principal composition: C= 0,036%, Cr=12,206%, Ni=7,847%, Fe=78,365, Si=0,696%, Mn=0,427% and W=0,146%. During the chemical composition analyze, other elements have been found, with the concentration under 0.1%: P, As, V, Nb, S, Ti, Co, Mo, Al and Cu. The equivalent content of chromium (Cr)e=13,548 and nickel (Ni)e=9,158 was determined using the well known relations [2]. After casting, the samples were subjected to a hardening heat treatment. From each sample there were manufactured three specimens designated 1, 2 and 3. The arithmetic mean of the specimens 1-3, for each testing time give the value m, considered representative for the tested material. Mechanical properties: R m = 1002 MPa, R p0.2 = 701 MPa, hardness = 30 HRC, elongation A 5 = 8.8 %, rupture constriction Z= 31.3%. In conformity with Schäfller diagram, the microstructure is formed by 90 % martensite and 10 % austenite G32-10 RECOMMENDATIONS FOR REPORTING THE TEST RESULTS ASME G32-10 [1] recommends reporting the information listed below, if applicable, for each material tested: 1. identification, specification, composition, heat treatment, and mechanical properties including hardness, measured on the specimen, sample or the stock from which it came; 2. the method of preparing test specimens and test surface (preferably including surface roughness measurement, before tests beginning); 3. number of specimens tested; 4. a tabulation giving the following information on each specimen tested: total cumulative length of exposure in hours (h), total cumulative mass loss in milligrams (mg) and total cumulative mean depth of erosion (MDE) in micrometers (m); 5. maximum rate of erosion; 6. nominal incubation time: the intercept on the time axis of the straight-line extension of the maximum-slope

21 Mircea Octavian Popoviciu, Ilare Bordeasu: Proposed Model for Cavitation Erosion Test Results Presentation; Machine Design, Vol.5(2013) No.1, ISSN ; pp portion of the cumulative erosion-time curve; while this is not a true measure of the incubation stage, it serves to locate the maximum erosion rate line on the cumulative erosion versus time coordinates; 7. erosion threshold time: the exposure time required to reach a mean depth of erosion of 1.0 m; 8. the cumulative exposure times to reach a mean depths of 50, 100, and possibly 200 m; designated t 50, t 100 and t 200 respectively; 9. a tabulation giving the normalized erosion resistance and normalized incubation resistance for each material tested, relative to one of the reference materials, included in the test. Calculate these values from averaged data of replicate tests of the same material; 10. tabulation of cumulative mass losses and corresponding cumulative exposure time for each specimen; 11. plot of cumulative mean depth of erosion versus cumulative exposure time for each specimen; 12. report any special or unusual occurrences or observations; Table 3. Tests Results 5. PRECAUTION MEASURES TAKEN DURING TESTS The total exposure time was 2.75 hours, divided in 12 periods, the first of hours, the second h and the rest of 0.25 h. The total exposure duration was chosen empirically, taking into account the data disposition in the plots. In the MDE=f(t) plot, the experimental points present reduced scatter and frequently the last four points respect a linear regression. Consequently it was considered that after 1.5 h of testing, the erosion progresses in a constant manner and the maximum exposure time must not overcame 2.75 hours. Before beginning the tests but also after each period, the specimens were washed successively with tap water, distilled water, alcohol and acetone, carefully dried and finally weighted. From each sample were manufactured five probes but only three were tested until the final exposure time. The spare pieces were used in exceptional situations when one of the specimens suffered damages during the tests. 6. EXPERIMENTAL RESULTS The principal results required by G32 Standard are presented in Table 3. Specimen MU* S1 S2 S3 m Total length of exposure hours Total cumulative mass loss mg Tot. cum. mean depth of erosion μm *Measurement unit The comparison is made with a steel, used for the manufacturing numerous hydraulic turbines. The chosen steel was OH12NDL. Specimens from this steel were tested in the same facility and presented m mean depth erosion after 2.75 hours. It results that 12/8 has better cavitation erosion qualities Cumulative mean depth of erosion The characteristic curve mean depth of erosion against time for the mean of tested specimens is presented in Figure 2. From the points disposition results that the simplest regression line is that obtained with a second degree polynomial. It also results with clarity that for short exposure time the points have the trend to follow a curved line but for longer exposure time a straight line represent better the point s disposure. For comparisons in the same figure there are presented also the data for the reference steel OH12NDL. Figure 3 present the normalized MDE relative to the reference material. It can be seen that the erosion of the tested steel is only 45% from that of OH12NDL stainless steel. MDE h <µm> /8 OH T2. OH12NDL and 12/8 0 0,25 0,5 0,75 1 1,25 1,5 1,75 2 2,25 2,5 2,75 3 Time x <hours> Fig.2. Comparison between OH12NDL and 12/8 steels Rate 0,50 0,45 0,40 0,35 0,30 0,25 0,20 0,15 0,10 0,05 0,00 T2. Normalized MDE (12-8/OH12NDL) 0 0,25 0,5 0,75 1 1,25 1,5 1,75 2 2,25 2,5 2,75 3 Exposure Time x <hours> Fig.3. Normalized Mean Depth Erosion (12-8/OH12NDL) 13

22 Mircea Octavian Popoviciu, Ilare Bordeasu: Proposed Model for Cavitation Erosion Test Results Presentation; Machine Design, Vol.5(2013) No.1, ISSN ; pp Regression Equations In a previous paper [3] a second degree polynomial regression was applied. Such an equation is useful only when the results can be presented exclusively as cumulative loss against time. In the present paper there were applied both fourth degree and exponential equations. The exponential equation is [2]: H Bx A x 1 e (1) In Table 4 are given the coefficients of the exponential equation, as well as the standard estimation error and the tolerance interval for a 0.99 probability. Table 5 gives the same values for the fourth degree polynomial equations. Table 4. Exponential Equation Coefficients and Values Specimen ser ±IT99% A B UM μm μm m Sp. UM Table 5. Fourth Degree Regression Equations Coefficients and Values Equation R 2 R ser μm m x x x x ,9998 0, x x x x ,9960 0, x x x x ,9991 0, x x x x ,9992 0, ±IT99% μm The great value of R, for all specimens, confirms that the fourth degree polynomial approximate excellent the measured points. Besides the correlation coefficient, it was computed also the standard estimation error : equations the 0.99 probability tolerance interval is very restraint but encompasses all the measured points. The fourth degree polynomial is a little bit better because it gives a smaller tolerance interval. ( h H ) n 2 2 ser (2) which represents the scatter in the vertical direction of the observed points about the regression line. Assuming a normal distribution of the scatter, and a 99% probability, the 99% tolerance interval is obtained as: TI 3 ser (3) Fig.5. Measured data and tolerance interval for polynomial equations Fig.4. Measured data and tolerance interval for exponential equations Figures 4 and 5 present the measured results for all three specimens, the characteristic curve for the mean value and the 99% tolerance intervals. The superior limit is presented for the weakest specimen (H1 in our case) whiles the inferior limit is computed for the strongest specimen (H2 in our case). The tolerance interval is not symmetric because the equation for H1 and H2 are not identical. For both regression Incubation Time Analyzing the problem of the incubation time the G Standard states that this period is usually thought to represent the accumulation of plastic deformation and internal stresses under the surface that precedes significant material loss. Exact measures for the duration of the incubation period can not be obtained, so two approximations erosion threshold time and nominal incubation period are recommended to be reported [1]. Considering that 1.0 m mean depth of erosion is an accurately measurable value, it was defined as erosion threshold time. On the other hand the nominal incubation time was defined as the intercept with the time axis by the extension of the straight line of the maximum-slope portion. Examining Fig. 1, it can be seen that for the 12/8 steel the maximum slope is placed in the final stage of 14

23 Mircea Octavian Popoviciu, Ilare Bordeasu: Proposed Model for Cavitation Erosion Test Results Presentation; Machine Design, Vol.5(2013) No.1, ISSN ; pp tests. To eliminate the influence of the regression curve, the used procedure was: obtaining the equations for the regression of straight lines passing through the last three points and computing the nominal incubation time as the rate between the regression coefficients. The values obtained for these two approximations are presented in table 6 together with other important figures. To obtain the erosion threshold time we use the three types of regression equation exponential, fourth degree and second degree polynomials, the last one obtained in a previous work [4]. The results are presented in Table 6. The value of nominal incubation time must be smaller than the erosion threshold time. As it can be seen from table 6 this do not happen every time. Upon our opinion the nominal incubation time is not a very reliable indicator of the erosion process. The values of the erosion threshold time depend on the regression equation adopted. Table 6. Nominal Incubation Time and Erosion Threshold Time Specime n NIT hours ETT S hours ETT F hours ETT E hours m As a consequence, the researches expenses increased exponential with the total exposure time. In many cases the researchers operating with high quality steel renounce to specify the test time for reaching the 50 or 100 m mean depth. As an example in the presented researches by Yeng-Min Chen [5] the total exposure time was taken 180 minutes (3 hours). The use of regression curves offers the possibility to obtain by extrapolation the time values needed for great depth of penetration. On the other hand, the correct mathematical correlation between x and h, established through physical considerations is not known and consequently the extrapolation of the regression curves across the tested domain is not recommended. Neglecting those recommendation, in Figure 7 is presented the extrapolation of the regression curves till x=20 hours for three regressions (second and fourth degree polynomials and an exponential one). Details for the second degree polynomial were presented in a previous paper [3]. It can be seen that the values for the fourth degree polynomial and the exponential regression are enough close. The second degree polynomial gives greater differences. 7. EXTENDED EROSION Fig.7. Great Extension of Regression Curves The G32 Standard recommends determining the time at which the mean erosion depth reaches 50, 100, and if possible 200 m. Because the tested stainless steel has very good resistance, the extended test time is expected to be very long. On the other hand, a component of the test facility (the horn) is subjected to high fatigue stresses. Although the materials used for manufacturing the horn possesses high fatigue resistance, after some testing cycles occur fissures and the horn must be replaced. In Figure 7 is presented one cracked horn. Fig.8. Small Extension of Regression Curves Fig.6 Cracked horn It is easy to observe that for MDE=200 μm enormous differences occur between those three curves, even if for small extensions of the curves (until 3.3 hours) all the used regression equations approximates well the experimental data, as can be seen in Figure 8. To determine the most favorable regression equation, at least for one specimen, the tests must be continued till 10 hours, perhaps even more. In Table 7 there are presented the numerical values obtained for different equations. There are very great differences even for 50 μm MDE. 15

24 Mircea Octavian Popoviciu, Ilare Bordeasu: Proposed Model for Cavitation Erosion Test Results Presentation; Machine Design, Vol.5(2013) No.1, ISSN ; pp Table 7. Numerical values if the curves are extended till 50, 100 and 200 μm EE S 50 EE F 50 EE E 50 EE S 100 EE F 100 EE E 100 MU h h h h h h h h h m MU-measuring units; 1,2,3,m specimens, EE E 100- extended erosion time necessary for a mean depth erosion of 100 m and exponential regression (S to second degree and F to fourth degree polynomial regressions). EE S 200 EE F 200 EE E RECOMMENDATION FOR CAVITATION EROSION TEST PRESENTATION Table containing the differences between the employed test facility and the G32-10 Standard. Information about the tested material (chemical composition and principal mechanical properties) and the conditions of the specimen exposed surface. Table containing the final test results (exposure time, mass loss and mean depth erosion). Plotted results for cumulative mean depth erosion (micrometers) against exposure time (hours) as well as plotted results for normalized MDE. Proposed regression equation and the resulted standard estimation error. Table containing erosion threshold time computed with the chosen regression equation and the nominal incubation time. The extended erosion experimentally obtained at least for a MDE of 50 micrometers. 9. CONCLUSIONS 1. Comparing the tested stainless steel with the OH12NDL steel it results that the 12/8 steel has very good cavitation erosion qualities. 2. Both the values of the correlation coefficients and the values of the standard error of deviation confirm that the fourth degree polynomial and the exponential equation are suited for estimating the measured data. 3. For a small extrapolation (under maximum hours) the obtained MDE are enough close for the used regression equations. 4. For a greater extrapolation (over 3.3 hours) the MDE differences are very great. At least for one specimen the tests must be continued till 10 hours. In this way we can obtain information upon the best regression equation. 5. The erosion threshold time and the nominal incubation time have relatively close values. For the tested steel, those parameters do not give important additional information. REFERENCES [1] *** ASTM G32 10 (2010). Standard Test Method for Cavitation Erosion Using Vibratory Apparatus, [2] Bordeasu, I.; Popoviciu, M.; Patrascoiu, C-tin.; Bălăsoiu. V. (2004). An Analytical Model for the Cavitation Erosion Characteristic Curves, Scientific Bulletin Politehnica University of Timisoara, Transaction of Mechanics, Tom 49 (63), p , ISSN: [3] Bordeasu, I. (2006) Eroziunea cavitaţională a materialelor, Editura Politehnica, ISBN: (10) ; (13) , Timişoara [4] Popoviciu, M. O.; Bordeasu, I. (2011), Reappraisal of cavitation Erosion Test Results. Second Degree Polynomial Regression, Autumn Conference of the Academy of Romanian Scientists, September 8-10, Mioveni, Romania [5] Chen Y-M.; Tessier, J-J.; Caze, D. (2008). Cavitation Erosion Test for ASTM G32 Method Revision, Cavitation Work Shop, Grenoble, March 2011 [6] Bordeasu, I; Popoviciu, M.O.; Novac, D. (2012). Machine Design, Monograph University of Novi Sad, Faculty of Technical Sciences, vol. 4, Nr. 2, pp , ISSN

25 machine design, Vol.5(2013) No.1, ISSN pp Original scientific paper STUDIES AND EXPERIMENTAL RESEARCH ON THE BEHAVIOR IN EXPLOITATION OF THE BEAM ENSEMBLE OF THE INGOT HANDLING CRANES IN SIDERURGICAL METALLIC WORKSHOP Cristina Carmen MIKLOS 1, * - Imre Zsolt MIKLOS 1 - Carmen Inge ALIC 1 1 POLITEHNICA University of Timisoara, Faculty of Engineering Hunedoara, Dept. Engineering and Management, Romania Received ( ); Revised ( ); Accepted ( ) Abstract: The paper includes in first part the experimental investigations proper, carried out according to the steps of the research program given in the final part of the precedent paper. The research has been done on the geometry of the beam ensemble of the ingot-handling crane from the Pit Furnace Workshop under analysis, and the results have been interpreted both qualitatively and quantitatively, with respect to the influence of the specific exploitation conditions in siderurgy upon the safety of these structures. The critical analysis and the interpretation of the results of the measurements was further extended over the investigation of the causes and factors that made for the evolution of geometrical imperfections and deviations that have been noticed. In the final part of the paper we suggested possibilities of turning into account the data and information obtained as a result of the investigation and experimental research programs. Key words: experimental investigation, geometrical imperfections, structural imperfections, deviations. 1. INTRODUCTION The results of the studies and experimental investigations, [1], meat to analyze the behavior in time of the beam ensemble of the ingot-handling cranes in siderurgical metal workshop proved the existence of certain factors that generate both geometrical and structural imperfections, due to which the resistance structures in exploitation are different from the theoretical ones, as to their behavior when charged. An estimate on the safety in exploitation of the beam ensemble of the ingot-handling cranes in such a workshop implies therefore the approach of the imperfections, deviations and flaws that can appear during manufacturing, mounting or exploitation, respectively those that arise in the course of time, due to the specific actions determined by the technological process. 2. EXPERIMENTAL RESEARCHES CARRIED OUT. THE OBTAINED RESULTS The experimental investigations proper were carried out according to the steps in the Research Program given in the final part of the preceding paper. The estimation of the magnitude of deformations and movement that appeared in the course of over 30 years of exploitation has been done according to the measurements made inside and outside the workshop. The succession of the operations can be seen in the Plan of Topographical Operation and Adjacent Measurements given in Table 1, where, according to each step, we mentioned the topographical method we used and the measurements that we have been done on site. Table 1. The Plan of Topographical Operation and Adjacent Measurements THE STEP and the Target Fundamental magnitudes of the objective Means/ procedures used and the necessary calculations A. ALTIMETRIC MEASUREMENTS A1. The level determinati on of detail points The height of the detail points situated on level 2t and 2s - The method of the mean geometric leveling - The calculation and compensation of level Measurements carried out on site Level differences with respect to the level guide marks RNE1 and RNE2 B. DIRECT MEASUREMENTS OF THE ENSEMBLE OF THE INGOT HANDLING CRANE GEOMETRY The distance B1. The between the Direct measurements of the effective axes of the respective distances, setting rolling tracks followed by an analytical apart of the lying on and graphical processing of rolling different the results tracks lines B2. Horizontal deviations from the straight of the lines Horizontal distances between the theoretic view axis and the line of track. Views and direct readings of the horizontal deviations with respect to the end points of the rolling tracks. *Correspondence Author s Address: Politehnica University of Timisoara, Faculty of Engineering Hunedoara, 5 Revolutiei Str, Hunedoara, Romania, cristina.miklos@fih.upt.ro

26 Cristina Carmen Miklos, Imre Zsolt Miklos, Carmen Inge Alic: Studies and Experimental Research on the Behavior in Exploitation of the Beam Ensemble of the Ingot Handling Cranes in Siderurgical Metallic Workshop; Machine Design, Vol.5(2013) No.1, ISSN ; pp THE STEP and the Target B3. The geometry of supporting areas and clamping the rolling beams to the pillars of crossframes. Fundamental magnitudes of objective Means/ procedures used and the necessary calculations - The distances from the face of the pillars to the axis of the upper foot of the rolling beams; - The dimensions of the upper foot of the rolling beams; - The dimensions of the crosssections of the pillars Measurements carried out on site idem to point B1 For the altimetric leveling, taking into account the provisions of [2], we used as a topographic instrument the Ni007 Automatic Level, recommended for precision measurements, for which, in the case of using the invar leveling rod, leads to square error of mm/km of double leveling. After the measurements on site have been done, the provisional elevation calculated and the leveling traversing compensated, we obtained the final elevation marks of the detail points positioned on the resistance structure. As a result of direct measuring of the elements of the beam ensemble of the rolling tracks, we determined: a. The actual distance between the rolling tracks, respectively the distance between the rolling lines whose axes had been marked on this purpose, on the upper surface of the rails, in the area of the axis of each cross - frame. The results of the measurements after having carried out the necessary corrections, as well as the deviations calculated with respect to the nominal distance are given in Table 2. b. Deviation from the straightness of the rolling tracks, given in Fig.1, and estimated according to [3], by direct reading (on a horizontal ruler, attached to a vertical rod) of the rail deviation with respect to the axis of the viewpoint from station S toward point V (also marked in Fig.5 of the preceding paper), both points being placed at the end bay of the workshop. Table 2. Measurements of the effective distance between the rolling tracks Axis No. Cross frame Effec. opening [m] Opening I - II Nom. opening [m] Deviation [mm] Effect. opening [m] Opening II - III Nom. opening [m] Deviation [mm] 1 30, , , , , , , , , , ,044 30, ,002 11, , , , , , , , , , , , , , ,001 0 c. The geometry of the areas where the rolling beam are clamped to the pillars The direct measuring of the dimensions and distances exemplified in Fig.2 for the 1 st row of pillars have been carried out during the repairs on the Pit Furnace Workshop, after the rolling rails were removed in order to be replaced and the system of their clamping to the upper foot of the beams be modified. 18 Fig.1. Leveling of horizontal deviation from the straightness of the rolling track Fig.2. Dimensions and distances characteristic to the geometry of the rolling beams The values, Table 3, were obtained after having measured the dimensions marked b 2 and a, with a steel ruler and the dimension b 1 with a sliding calipers.

27 Cristina Carmen Miklos, Imre Zsolt Miklos, Carmen Inge Alic: Studies and Experimental Research on the Behavior in Exploitation of the Beam Ensemble of the Ingot Handling Cranes in Siderurgical Metallic Workshop; Machine Design, Vol.5(2013) No.1, ISSN ; pp Table 3. Measurements of characteristics to the geometry of the rolling beams Pillars of: row I II [mm] row II I [mm] Zone a b1 b2 d a b1 b2 Axis , opening 1_ Pillars of: row I II [mm] row II I [mm] Zone a b1 b2 d a b1 b2 Axis , opening 2_ Axis opening 3_ Axis , opening 4_ Axis , opening 5_ Axis , opening 6_ Axis , opening 7_ Axis , opening 8_ Axis , opening 9_ Axis , opening 10_ Axis , opening 11_ Axis , opening 12_ Axis , THE CRITICAL ANALYSIS AND INTERPRETATION OF THE RESULTS OF THE EXPERIMENTAL PROGRAM The values obtained through the measurement program give the possibility of analyzing the geometrical behavior of the beam ensemble of the rolling tracks in the workshop under consideration, which has been in exploitation for the last 30 years. This analysis has been further carried on with respect to the following aspects: quantitatively, by a comparison of the values of the experimentally determined deviation with the admissible ones, according to the actual designing norms and provisions; qualitatively, by an estimation of the causes and factors that lead to the respective geometrical imperfections, as well as considerations upon their influence on the work conditions when the resistance elements or the entire structure are charged Aspects of the quantitative analysis According to the aspects mentioned above, the analysis and interpretation of the results of the study program as well as the technical and topographic measuring of rolling track ensemble looks as follows: a. The level marks of the rolling beams from the Pit Furnace Workshop according to the longitudinal section drawn in Fig.3 shows obvious differences both along the line and cross - sectionally. The highest level difference between the maximum elevation mark and the minimum one, measured at the face of the upper foot was 62 mm for the rolling beams on the row II I and 27 mm for those on the row I II, values which do not rank within the maximum 20 mm difference admitted by norm [4] along the same line. Fig.3.The longitudinal execution shape of the rolling beams; Length scale 1:100 ; Height scale 1:1 It has also been noticed that the difference of level have also been surpassed cross-sectionally (norm [4] admits a maximum of 10 mm) starting with axis 5 along the entire workshop, with a maximum value of 47 mm at axis 9. The value of these deviations as well as the important slope variations shown by the rolling tracks have caused problems in the good functioning of the ingot handling crane, wearing the rails and causing extra strains both for their own resistance structure and for the one of rolling beams. That is why, on the occasion of repair of the workshop, after about ten years of exploitation, the solution envisaged was to weld on the upper foot, in the area marked Z* in figure 4, some intermittent fixing plates of variable thickness. But this solution solved the problems mentioned only partially, as one can notice from the situation resulting from the measurements and given in Fig.3. 19

28 Cristina Carmen Miklos, Imre Zsolt Miklos, Carmen Inge Alic: Studies and Experimental Research on the Behavior in Exploitation of the Beam Ensemble of the Ingot Handling Cranes in Siderurgical Metallic Workshop; Machine Design, Vol.5(2013) No.1, ISSN ; pp b. The results of the measurements pointed out to the fact that the actual deviation from the distance between the beams is above the limit value of the deviation, (which, according to norm [5] is in the case of the workshop under consideration mm), all along the zone between axes 4 and 9 of the cross frames in the Pit Furnace Workshop; a maximum deviation of 44 mm has been noticed at axis 6. c. Horizontal straightness deviations of the rolling track lines, gives above in Fig.1, surpass the limit deviation accepted according to [5], at axes 4 and 10 on line I II and along the zone between axes 5 and 11 on the line II I in the Pit Furnace Workshop. d. The level differences between the upper sides of the rolling tracks, calculated according to the values pf the elevation marks, have been given in Fig.4, by drawing the longitudinal shape of each line of the rolling track. One can notice that the maximum value of the admissible difference, i.e. 20 mm, according to [5] between the maximum elevation mark and the minimum one on the same line of the rolling track are exceeded on both lines on row II of pillars, the deviations having a 31 mm maximum value on line II I and 52 mm on line II III of the rolling tracks; one can also notice an excess in the 1/1500 maximum value of the admissible slope. 20 Fig.4. The execution longitudinal shape of the rolling tracks Length scale 1:100 ; Height scale 1:1 As to the cross-sectional level differences on the upper surface of the tracks, the admissible value according to [5], is 10 mm on the support and 15 mm on the field, but they are exceeded in the workshop from axis 5. At the same time, in the Regenerative Chamber Workshop the level differences at axes 4, 7, 8, 9, 10, 11 and 12 of the cross section frames fail to rank within the same admissible deviation values Aspects of the quantitative analysis Qualitatively, the causes and factors that made for the geometrical imperfections we pointed out to, have been considered as follows The overall conception of building the resistance of the workshop The solution of inserting some incomplete cross sectional frames within the resistance structure, the rolling beams seated on the upper part of the pillars, as well as the constructive structure of these supporting points (given in Fig.2 and Fig.4 of the preceding paper) can be considered as potential factors in the modification of the geometrical structure under exploitation charges. We came to this conclusion after having analyzed the degree of freedom in the above-mentioned joints, which, although rigid to anything but the movements of positioning marks, when the resistance elements in the joint, may yet allow both horizontal and vertical movements. We can thus explain the longitudinal and cross sectional slides, which, at axes 5, 8 and 10, where the three incomplete frames are placed, acquire significant values. The similar deviations, ranging within the same order of magnitude or, in some cases, even above, which were noticed with the resistance of elements from the bays next to the incomplete frames, may be caused by the transmission of the unfavorable behavior from the cross sectional frames to the neighboring ones, which would represent an experimental confirmation of the fact that the phenomenon of spatial interaction of the resistance structures, specific to workshops with traveling cranes, appears as an interaction in movement. The favorable effects of the phenomenon of spatial interaction, considered in terms of a redistribution of the charge from the more loaded frames to the less loaded ones, or to those not loaded at all with the charges deriving from the action of the traveling cranes, may also be accompanied by unfavorable effects, i.e. deformations, which, aggravated resistance structure Imperfections and appearing in the stage of mounting the resistance structure The high values of the deviations noticed with the elements of the rolling beam ensemble, both horizontally and vertically, may be caused either by deviations from the pillar positioning marks, when mounting them, or a later sinking of the foundations, accompanied by twisting. The second hypothesis would be partially justified by the movements noticed on the 2 nd level of the measurements, i.e. on the pillars of the cross-sectional frames in the area of supporting the rolling beams. This sinking, which according to the elevation marks obtained by topographic measurements, should have had maximum values in the central pillar area, between axes 1 5, would have modified similarly the elevation mark RNE1, placed in the moment of construction. The checking of this hypothesis was done by measuring the elevation mark mentioned on RNE1, with respect to

29 Cristina Carmen Miklos, Imre Zsolt Miklos, Carmen Inge Alic: Studies and Experimental Research on the Behavior in Exploitation of the Beam Ensemble of the Ingot Handling Cranes in Siderurgical Metallic Workshop; Machine Design, Vol.5(2013) No.1, ISSN ; pp the known elevation of a topographic mark in the area. The elevation mark we obtained in this way, i.e m, compared to that of m, found on the mark, proves that the foundations did not sink. That is why, we considered as valid the first hypothesis i.e. that of incorrect positioning of the pillars for resistance structure, the deviations being not only longitudinal or cross sectional, but also with respect to the positioning of their setting upon the foundation Imperfections due to exploitation The intense exploitation of the three travelling cranes with rigid support, operative in the Pit Furnace Workshop has caused in time a loosing of the screws that fix the rails, which caused both deviations from the straightness of the rolling tracks, and premature wear of the rails. The supplementary strains upon the resistance structure of the travelling cranes due to the existence of important uneven areas, and of deviations from the admissible tolerance with respect to the nominal distance between the rolling tracks, had unfavorable effects upon the geometry of both the crane and the ensemble of rolling beams. Moreover, the systematic use of the travelling cranes in supplementary technological operations such as the cleaning of the pit furnace hearth, made for the accentuation of the geometrical imperfections and deviations that mentioned. At the same time, we mention modifications in time of the technological process in the Pit Furnace Workshop, which imposed changes in the configuration of the train of travelling cranes. Thus, contrary to the provisions in the project, which implied three travelling cranes, two 50 kn ones and a 50/300 kn one, placed at one end of the train, the workshop is furnished with two 50/300 kn travelling cranes, plus a third one, in between, with a capacity of 100/400 kn. Taking into account the important contribution of the charges upon the travelling cranes as compared to that of other categories of actions, one can consider that is a similar manifestation of the negative influence of intensity modifications and of the alternation of the configuration of charge trains. This influence, completing the effects of the other deviations and geometrical inexactitudes analyzed in this paper, lead to the alteration of the initial geometry of the ensemble of rolling tracks and, at the same time, to an increase in the strains that generated these supplementary deformations and movements of the resistance structure of the workshop in exploitation. 4. SUPPLEMENTARY INVESTIGATIONS UPON THE EXPLOITATION BEHAVIOR The visual analysis of the rolling beams lead to the observance of other flaws, which may be the result of the inadequate physical and mechanical characteristics of the material, as follows: the existence of longitudinal fissures in the upper foot of the 11m wide rolling beam from bays 2 3 and of the 2 nd row of pillars, whose position is given in Fig.5a and b. the existence of cross sectional fissures in the upper foot of the 11m wide rolling beams from bays 2 3 on the 1 st row, respectively bays 3 4, 4 5, 5 6, 10-11, and on the 2 nd row (Fig. 5a, b) and of fissures in the core of the rolling beams on the 1 st row of pillars (Fig.5b). a. Mapping of longitudinal fissures and sectional fissures in the upper foot of the rolling track beams b. Details of the fissures from the feet and core foot of the rolling track beams Fig.5. Results of the visual analysis of the rolling beams from the Pit Furnace Workshop The appearance and development of such fissure-type flaws, as the ones noticed particularly in the smaller rolling beams could also be caused by the mechanical aging of the steel, a behavior that can be associated to the phenomenon of flowing, where the resistance of the material increases, but its tenacity and deformability characteristics decrease. 5. CONCLUSIONS The studies and experimental research envisaged by the steps of the project conceived with this aim prove the possibility of estimating the real geometry of the resistance structures in exploitation. The investigation of the causes and factors which made for the evolution of geometrical imperfections and deviations up to values which go way beyond the admissible ones, as established by the norms in force, pointed out to the important role of the general conception and the detail approach in the design of resistance 21

30 Cristina Carmen Miklos, Imre Zsolt Miklos, Carmen Inge Alic: Studies and Experimental Research on the Behavior in Exploitation of the Beam Ensemble of the Ingot Handling Cranes in Siderurgical Metallic Workshop; Machine Design, Vol.5(2013) No.1, ISSN ; pp structures, elements that can influence in a significant way the later behavior of the construction in exploitation. The research carried out has been done the geometry of the beam ensemble of the ingot-handling crane from the Pit Furnace Workshop under analysis, and the results have been interpreted both qualitatively and quantitatively, with respect to the influence of the specific exploitation conditions in siderurgy upon the safety of these structures. The critical analysis and the interpretation of the results of the measurements was further extended over the investigation of the causes and factors that made for the evolution of geometrical imperfections and deviations that have been noticed. The valorization of these results can be done by their being taken as a basis for the monitoring programs of the constructions, both in order to obtain a closer physical modeling of the real resistance structures in view of evaluating their safety in exploitation. REFERENCES [1] Alic, C. (1998). Contributii la evaluarea gradului de siguranta al halelor metalice din siderurgie, tinand seama de conditiile reale de exploatare Teza de doctorat, Universitatea Politehnica din Timisoara. [2] Oprescu, N. (1973). Manualul inginerului geodez, vol. I, II, III. Editura Tehnica, Bucuresti,. [3] Cristescu, N. (1978). Topografie inginereasca, Editura Didactica si Pedagogica, Bucuresti. [4] *** STAS Constructii pentru sustinerea cailor de rulare ale macaralelor si podurilor rulante. Prescriptii generale. [5] *** STAS Cai de rulare si opritoare pentru poduri rulante si macarale. Prescriptii generale. [6] Coşarcă, C. (2003). Topografie inginereasca, Editura Matrix Rom, ISBN , Bucureşti. 22

31 machine design, Vol.5(2013) No.1, ISSN pp Original scientific paper THE IMPACT OF BEARING DEFORMATION IN THE FIELD OF PRESSURE AND ITS HYDRODINAMIC CHARACTERISTICS Koço BODE 1, * - Odhisea KOÇA 1 - Ilirian KONOMI 2 1 Polytechnic University of Tirana, Faculty Mechanical Engineering, Dep. of Mechanics, Tirana, Albania 2 Polytechnic University of Tirana, Faculty Civil Engineering, Dep. Of Hydrotechnics, Tirana, Albania Received ( ); Revised ( ); Accepted ( ) Abstract: This article aims to make use of a model developed in MATHCAD [12] for studying the impact of the deformation field of bearing material on the field of pressure and hydrodynamic characteristics of hydrodynamic bearings. The study and analysis of the problem shows that the deformation of the bearing as a phenomenon can not be neglected in the study of static and dynamic behavior of a rotating system as a whole. The built algorithm combines simultaneous solution of elastic and hydrodynamic problem and allows the study and analysis of a number of parameters that characterize the stiffness and the pressure field of hydrodynamic bearing. Adoption of the solution regarding plane elasticity theory with complex numbers allows efficient use of this solution in the study of the influence of deformation of the coating layer of the bearing. A model built on this basis has greater efficiency than analog models calculation (FEM methods) in terms of simplicity and speed of getting the results. Key words: hydrodynamic and elastic problem for circular ring 1. INTRODUCTION The hydrodynamic theory equations [1] allow us to write the characteristics of a contact in cases when the hydrodynamic loading operations are practically insensitive to the elements in contact. When work pressures are intense, thus a characteristic for high speed rotating systems there happen contact geometry changes that can not be ignored. The latter significantly influence the hydrodynamic lubrication because the lubricating layer thickness and contact surface deformations are roughly of the same order. In this case additional equations are needed that allow the calculation of the elastic deformations on the contact surface. The solution that reflects this effect and provides mutual connection fluid - structure, is obtained only based on the simultaneous usage of the hydrodynamic lubrication equations and linear elasticity equations through an iterative process. The generalized results are presented by variables without units. 2. SOLUTION OF ELASTO - HYDRODYNAMIC PROBLEM 2.1. Geometry of bearing The geometric parameters of bearing are shown in figure 1. R b e R c represent respectively the radius of shaft and of the non-deformed bearing, and t is its elastic layer thickness. (Fig 1) Fig.1. Geometry of couple Solution of this problem is made accepting that fluid viscosity is independent of temperature. Equations necessary to solve this problem are: The hydrodynamic equation for fluid flow. Linear elasticity equations. Law of fluid pressure - viscosity dependence The solution to the elastic-hydrodynamic problem for a random couple is obtained numerically by simulating the first two equations by means of an iterative process. Let s consider a fixed coordinative system xoy. Relative to it, the shaft is rotated with an angular velocity. Outer loading W is constant in size and direction. The direction: the center of the ring-bearing axis is determined W,r. by the angle, in such a way that 2.2. Calculation of pressure field The pressure field is calculated integrating the Reynolds equation [1]. For a bearing with ratio L 4, accepting D *Correspondence Author s Address: Polytechnics University of Tirana, Faculty of Mechanical Engineering, Sheshi Mather Teresa - 4, Tirana, Albania, kbode@fim.edu.al 0

32 Koço Bode, Odhisea Koça, Ilirian Konomi: The Impact of Bearing Deformation in the Field of Pressure And Its Hydrodinamic Characteristics; Machine Design, Vol.5(2013) No.1, ISSN ; pp x 0, y 0 and U rc 0, the Reynolds equation in variables without units is written: 24 2 H U P R H U 0 k x rc 1 H0 U 2 rc L z 0 K z 3 rc P z For the case when the bearing length is greater in comparison to its diameter L 4,, the above equation D is written as follows: 3 H U P H U 0 k x rc 2 where: H 1 cos dhe U U rc C rc (1) rc (2) For solving the equation (1) and (2), the method of finite difference method is used. In the circular direction it is applied the division by 6 while in the axial direction is applied the network with 10 divisions. The above solution ensures accuracy in the calculation of the pressure. The requirements for a more dense network are linked with the convergence problems during the simulation calculations. For this purpose, in the literature there are recommendations on the density of separating network [4] 2.3. Calculation of displacement field caused by hydrodynamic action Calculation of the displacement field in bearing surface is based on the following assumptions: Hypothesis of small deformations. Hypothesis of plan deformations. The first hypothesis can be verified with the help of the results obtained. The displacements in any case remain much smaller in relation to the thickness of the elastic layer. The method used to calculate the displacement depends on the problem studied. In the case of a long bearing, L 4, the problem is one-dimensional and for calculating D the displacement, two models are used: Analytical model. Thin layer model. For bearings with finite length, L 1 D the pressure field varies both in the axial and circular direction. To calculate the displacement field, the thin layer model is applied for the two dimensional case. This is treated in the following paragraph. For that purpose we make use of the relations that allow the calculation of the displacement field of bearing interior surface according to the solution of the plan case of the elasticity theory using complex numbers introduced by Solomon [8] and then developed further more by Villechaise [9] Application of the principle of superposing Tangential displacements have a negligible impact on changing the geometry of the contact of the couple bearing-shaft. Therefore the terms representing their effect are neglected and only the simplified model of equation for calculating the radial displacement are taken into consideration. These equations can be written [8]: U rc Nt k C H G ' ' 2 H G G ' cos k sin k k c k d k k k (3) where N t represents the number of terms of the Fourier series and k, k, are Fourier coefficients of pressure distribution. They are presented (3) 2 k 1 cos k P d (4) k sin k o The sign (-) represents the sign of the radial stress and pressure. The integrals (4) allow to calculate (2 N t +1) coefficients. To integrate the trapezium method is used. The terms number N t depends on the number M used for the divisions in circular direction of the bearing. Thus, for an odd number of divisions N t is M /2 and expression (4) is written: 1 k 2 M k M i0 cos k i Pi sin k i ku, k = 0,1,..., Nt, and 2 i. (5) M Since the geometry of the bearing is axi-symmetrical, the displacement calculations are made using the principle of superposing. In this case, we make use of the recognition of the displacement field on the bearing surface for a pressure unit, which operates at a given point. Let s present in a more detail manner this model: Let i j be the radial displacement of the point i, caused by the unit pressure, applied in the joint j, Fig 2. If pressure P j is applied at this point, the radial j displacement is given: Urc i P j 1. Considering the j 0 symmetry it is accepted: -i j 0 and j 0 i M i j for i j0. i i1 The radial displacement relation is generalized [8]: M 1 rc i P j j0 U j i

33 Koço Bode, Odhisea Koça, Ilirian Konomi: The Impact of Bearing Deformation in the Field of Pressure And Its Hydrodinamic Characteristics; Machine Design, Vol.5(2013) No.1, ISSN ; pp convergence of the iterative process the calculated displacements must be pondered before usage at the hydrodynamic part. The Algorithm called the over-relaxing algorithm is written [2]:. U H k 1 rc k 1 U H o k 1 rc U 1 U k 1 rc k rc (7) Fig.2. Network nodes for interior surface Thin layer model This model of calculation of radial displacement field is applied more accurately to the case of long bearing compared with the short ones. Thickness of the elastic layer should be very small compared to the diameter of the bearing. For the case of a long bearing the expression (3) is written [4], [5], [8], [9] U 1 c 1 2 c 1 t, z Cd P z (6) R rc, c c R where, C 0 c d Ec C Application of this model in the case of short bearing needs to consider the fact that pressure in bearing free z L, is zero. In this case it is assumed that boundaries, 2 in each segment (i, j+1) we have U rc UmN N 1. Axial distribution of pressure for the accepted division net is written: Pi, j1 Pi, j P m. 2 This pressure causes the displacement U jm (fig 3).Thus the displacement of node j is: k where, U 1 rc and U k rc, are the displacements of iterations k+1 and k; ; is coefficient of over-relaxation. Its values according to literature recommendations vary within the limits 0-1. In this case the problem deals with the definition of an * optimum to ensure the rapid convergence of calculation. The necessity of the application of such an algorithm in joining the two parts of the EHD program stems from the fact that the problem under study is not linear and the displacement caused by the hydrodynamic pressure field are in some cases larger than the thickness of the lubricating layer which has a third order effect to equations (1) and (2) and therefore a small change of the thickness of fluid layer leads to significant influence to the hydrodynamic pressure. The steps of the calculation are as follows U rc j = U mj U 2 1 mj Mutual action structure -fluid. Calculation method. In Figure 3, it is presented the algorithm of the elastohydrodynamic problem solution based on the analytical approximation. In the case of thin layer model the calculation process is simpler because the diffraction in Fourier series is not used [2]. For a given initial geometry it is calculated the pressure field and then the field of displacements. The latter deviates the cylindrical shape of the couple. Further it is calculated the new distribution of pressure. The calculation process continues up to the stabilization of the deformation process. This is linked with the getting of two displacement fields or of the two pressure ones that will meet the convergence requirements. To ensure the Fig.3. Schematic blocks of EHD solution with analytical method The test completion of this iterative process is related to the fulfillment of the condition: k 1 k 1 P,, 5 i j P i j 10 k 1 p i, j P i, j p (8) N Where, p is the admitted relative error, and N p, the number of points where the pressure value is positive. 25

34 Koço Bode, Odhisea Koça, Ilirian Konomi: The Impact of Bearing Deformation in the Field of Pressure And Its Hydrodinamic Characteristics; Machine Design, Vol.5(2013) No.1, ISSN ; pp The validity of hydrodynamic-elastic approximations. In this section the methods and assumptions used to solve the hydrodynamic-elastic problem are verified. The obtained results are compared with those given in the literature [1], [2], [3], [6], [7] in the case of long bearings L and those with finite length for 0. 5 D Long Couple L D L. D The hydrodynamic pressure values obtained using the analytical method and that of the thin layer are compared with the results of the literature [6], [7]. As it can be seen two values for the elasticity module are taken into consideration Short Couple L 4 D Let s compare the results obtained from the three elastohydrodynamic approaches for the short couple. The first couple material is stainless steel, while the second pair, bronze. Couple 1 The graphics of Fig.5a provide the pressure in the symmetry section for first couple for a relative eccentricity 0.6. These results show that in the case of this couple the pressure field, which is calculated by means of the two-dimensional thin layer model, is very similar to the results obtained by the finite elements method. In the same introduction we notice that the analytical model and that of one dimensional thin layer give smaller results in comparison with the finite element method. 26 Fig.4 a,b. The distribution of pressure on the average plane (long couple) The graphics of Fig.4 provide respectively the pressure distribution for isoviscous and piezo-viscous fluid. It is noticed a satisfactory compatibility between the results obtained with the two used methods and those published by Canway and Lee [1]. Couple 2 Fig.5 a,b The distribution of pressure on the average plane (short couple) The graphics of Fig. 5b give the pressure field in the symmetry section for the couple 2. The results reveal evidence of a significant difference of the pressure value calculated with the analytical method in comparison with that of the hybrid method [4], which represents a combination of finite elements method and that of finite differences. Thin layer model gives results close to the hybrid method.

35 Koço Bode, Odhisea Koça, Ilirian Konomi: The Impact of Bearing Deformation in the Field of Pressure And Its Hydrodinamic Characteristics; Machine Design, Vol.5(2013) No.1, ISSN ; pp CASE STUDY After we have presented the validity range of the above hydrodynamic-elastic approximation we consider a case study. First, let s compare the results obtained for the case of a rigid and deformable pair when it is calculated with a presumptive eccentricity. Secondly, we study the influence of the elasticity module in the hydrodynamic characteristics when calculations refer to an assumed eccentricity and load. In both cases, the results obtained with the analytical method are compared with those obtained using the thin layer method Field of the pressure and the layer geometry in the hydrodynamic and elastohydrodynamic regime In this section the hydrodynamic characteristics of a rigid pair are compared with those of a deformable pair. Then and it is presented the influence of elastic deformation in the pressure field and the lubricating layer geometry. Fig.6. The distribution of pressure on the average plan The geometry of the fluid layer The graphics of Fig. 7 represent the dependence of the fluid layer in the case of deformable and non-deformable pair. For a solid pair the layer thickness in the circular direction has the form of a sinusoid and its minimum 0 value is at the edges of the angle180. For a deformable pair there happens an increase of the fluid layer thickness. In this case it is accepted a great value for the eccentricity and a small value for the elasticity module to better highlight the deforming effect in the bearing behavior. Pair characteristics. Geometric characteristics: length L m beam ratio R a m 5 radial clearance C 5*10 m layer thickness t m Conditions of operation: rotational speed of the beam N 3000 rrot min relative eccentricity 0. 3 free border pressure p patmosferik As it can be seen L 3 1, C 2*10 dhe t 0. 4 Fluid properties: dynamic viscosity capacity measures 885 dan m piezoviscosity coefficient 0 D R 3 Elastic characteristics of the material elasticity module E c 942 MPa coefficient of Poisson Field of pressure c Graphics of Fig. 6 represent pressure curves for all the bearings in the symmetry plane. It is noticed that for the deformable pair the pressure area stretches in an arc While for the rigid pair it has an 0 approximately stretch of 45. R a Fig.7. Circular variations of thickness H in the average plan 3.2. Study of the influence of deformation coefficient C d in hydrodynamic characteristics In this section we present and compare the results obtained from the analytical method and that of the two dimensional thin layer one. The coefficient values C d are accepted in the range (0-1). The value 0 corresponds to the case of solid bearing for which the value of the elasticity module E c whereas the value 1 corresponds to the material with elasticity module E c 1178 MPa. The operation conditions of the pair correspond to the accepted relative eccentricity in the values 0. 6 and 0. 85, while the no unit load is accepted in values W 0 2 and W 0 4 and correspond to the values W N and W N. 27

36 Koço Bode, Odhisea Koça, Ilirian Konomi: The Impact of Bearing Deformation in the Field of Pressure And Its Hydrodinamic Characteristics; Machine Design, Vol.5(2013) No.1, ISSN ; pp Assuming eccentricity The graphics of Fig. 8 and Fig. 9 provide the circular variations for the pressure and the fluid layer at the symmetry level for eccentricity for different values C d. It can be seen that for the increase in the deformation coefficient C d there is a maximum pressure reduction, a reduction of the cavitations area and an increase of the fluid layer thickness in the average section. used method. In case of the usage of thin layer, H min is calculated for the lateral borders of the bearing while for the case of using the analytical method the calculations belong to the average plan. Fig.10. Variations of minimal thickness of fluid from C d for different values of H min Fig.8. Pressure variations for different values of C d in average plan For the value C d 1 the geometry of lubricating layer changes significantly in the effect of the elastic deformation presence. The minimum thickness layer H min for deformable pair is about 4 times greater than the one calculated for the non deformable pair. The graphics of Fig. 11 represent the hydrodynamic forces F variation. The bearing deformation decreases significantly the hydrodynamic force. It is noticed that the results obtained by the two methods are increased with the growth of eccentricity. Fig.11. Variations of hydrodynamic forces from for different values of C d Fig.9. Circular variations of thickness H for different values of C d in average plan. The graphics of Fig. 10 represent the minimum thickness variation H min obtained by means of the two methods. It is noticed that the curves are different from each other and that the minimum thickness calculated with analytical method is greater in comparison with those calculated by the method of thin layer. This is directly linked with the Assuming Load The impact study of the deforming coefficient C d in static deformation features of the pair in case of the alleged foreign load coincides with the real case of the functioning of the hydrodynamic pair. To balance the external applied load with the module and direction it is necessary to determine the equilibrium position of the axis in relation to the bearing. The search 28

37 Koço Bode, Odhisea Koça, Ilirian Konomi: The Impact of Bearing Deformation in the Field of Pressure And Its Hydrodinamic Characteristics; Machine Design, Vol.5(2013) No.1, ISSN ; pp of relative eccentricity corresponds to the definition of this position. The equilibrium condition is written W0 F. The relative eccentricity value which satisfies this condition is required by an interpolation method known as Regula-Falsi. For the case of an assumed load the reduction P max is not so important as in the case of the alleged eccentricity (Fig. 12) increase of the relative eccentricity with the increase of pair flexibility. For the hydrodynamic regime, the value 1 of eccentricity corresponds to the contact shaft bearing. Also it is found that the calculated values by the analytical method are greater than those calculated by the thin layer method. This clearly proves the hypothesis of plane deformations used for calculating the displacements which do not consider the axial variation of the fluid flow. Fig.12. Variations of maximal pressure from C d for different values of applied load When the load is great the minimum thickness H min calculated by two dimensional thin layer method is reduced very little under the dependence of the deformation coefficient C d. This does not happen with the results obtained from the analytical method. Fig 13 Also for the thin layer method results is noticed a reduction H min in the case of an assumed load and an increase in case of supposed eccentricity. Fig.14. Variations of maximal pressure from C d for different values of applied load 4. CONCLUSIONS In case of a long pair the analytical model represents a good approximation for elastic materials. In case of a short couple the pressure field differs in the axial and circular direction. In this case, the 2D thin layer model provides an efficient approximation for the elastic deformable materials, while for the materials with Poisson coefficient 0, 4 the model does not have the same efficiency. The analytical method, ignoring the axial pressure variation, does not provide satisfactory results. However it can be used as an approximate method for materials with small Poisson coefficient. The shear displacements are always much smaller than the radial ones. When calculating the static and dynamic characteristics of the rotary systems their influence is not taken into consideration. Fig.13. Variations of minimal thickness H min from C d for different values of applied load The graphics of Fig 14 show the influence of the elastic deformations on the relative eccentricity. It is observed an For a presumptive eccentricity the analytical approximation and that of 2D thin layer give approximate values for the maximum pressure p max and hydrodynamic force F but different for the thickness of fluid H min because the approximation neglects the axial variations For the assumed load the results change. The influence of the model is visible especially to chart Fig. 13 and Fig

38 Koço Bode, Odhisea Koça, Ilirian Konomi: The Impact of Bearing Deformation in the Field of Pressure And Its Hydrodinamic Characteristics; Machine Design, Vol.5(2013) No.1, ISSN ; pp REFERENCES [1] Conway H.D, Lee H.C. The analysis of lubrication journal bearings, Journal of lubrication Technology, [2] Jahn S.C, Sinhasan R, Singh D.V. The study of elastohydrodynamic lubrication in journal bearings with piozoviscous lubricants. 1990, ASLE Transactions 1999 [3] Braun M, Dougherty J., Hydrodynamic analysis and fluid solid iteration effect on behavior of a complaint thick journal bearing (part 2), Results, ASME Transactions, [4] Grugin A.N. Fundamentals of hydrodynamic theory of heavily loaded cylindrical surfaces, Mashinostroene, Moscow, No 30, 1949 [5] Reynolds O. On the theory of lubrication and application to Beuchamp experiments, Trans Roy London [6] Carl T, E. An experimental investigation of a cylindrical journal bearing under constant sinusoidal loading, Lubrication and Wear [7] Carl T, An experimental investigation of cylindrical journal bearings under contact and sinooidal loading, Lubrication and Wear, [8] Solomon. S,, The mathematical basic of the theory of elasticity, Bucharest, [9] Villechaise B, Etyde bidimensionale des contact largest entre domains elastiques finis, Journal of tribology [10] England A Complex variable in elasticity, Wiley Interscience [11] Bode K, Koçi O, Konomi I. An analytical model for solving the elastohydrodynamic problem of the bearing, BTS-UPT-Tirane

39 machine design, Vol.5(2013) No.1, ISSN pp Preliminary note DESIGN ANALYSIS ALGORITHM OF A PETROL ENGINE-POWERED AIR COMPRESSOR Emmanuel SIMOLOWO 1, * -_ Taiwo OLUMIDE 2 1 Department of Mechanical Engineering, University of Ibadan, Ibadan, Nigeria 2 Department of Mechanical Engineering, Lagos State University, Lagos, Nigeria Received ( ); Revised ( ); Accepted ( ) Abstract: The complete design of a petrol engine-powered air compressor comprises five critical aspects namely, force analyses, materials selection, power transmission, gearing and bolting. The learning algorithm developed in this work encompasses the first three stages of the design for the compressor drive train. The algorithm structure presented herein is characterized by design- input entry, design- path selection, and results computation and display. Unique features of the program include; a multi-purpose structure due to the various sessions embedded in the program and the multiple-scenario design ability which enables up to four simultaneous designs on different windows. The validation of the software algorithm was done by comparing the results generated for a case study in literature with those obtained following numerical procedures. The results were in the same range with no differences in values. In conclusion, interactive learning software that will capture the complete design of a powered air compressor has been initiated in this work by the development of software algorithm for kinematics, force analysis and power transmission design of the air compressor assembly. Key words: Machine Design, Interactive Learning, Petrol Engine, Air Compressor 1. INTRODUCTION This work is on the development of an Interactive learning computer algorithm for the design of a petrol engine-powered air Compressor (Figure 1). It presents the concept behind the air compressor and the functions of compressor components with the assertion of originisation as an important factor in engineering education [1, 2, 3]. The design concepts presented in this work can be synthesized and analyzed in the following stages (i) Preliminary design of a compressor drive train (ii) Preliminary design of shafts for a compressor drive train (iii) Design of spur gears for a compress drive train (iv) Design of the head bolts for an air compressor. [14]. An air compressor is a machine that decreases the volume and increases the pressure of a quantity of air by mechanical means. Air compressors have helped many engineering equipment in performing various process such as breaking of ground by jack hammer tighten or loosen lug nuts on automobile wheel by the use of pneumatic wrenches and spraying of paint through an atomizing nozzle [6, 7]. Computer algorithms have helped engineers to efficiently develop correct, reliable, and robust programs that assist in design and eradication of errors during such processes. Various problems are encountered when performing numerical design. These include: (i) Tedious, Long and cumbersome design procedures:- In designing a product various equations are required to be used and each of which have their own branches of formulas and considerations. (ii) Limited illustrations while learning: - Numerical approach hinders student from adequately visualising concepts that are difficult or impossible to view. The objective of this work is to aid the effective teaching and learning of the design of a compressor drive train by developing a computer program for its design. Specifically, this work is to perform; (i) The numerical design of compressor drive train (ii) The design of the shafts of the drive train which basically involves making decisions that help in providing information for the complete design. Such decisions with regards to the shaft of drive train include maximum and minimum forces, maximum and minimum resultant forces, maximum and minimum moment on the shaft, mean and alternating components of both moment and torque and many more which determines the design. (iii) The development of algorithm which constitutes the bulk of the work to be done. 2. NUMERICAL DESIGN The basic parts of the air compressor-assembly considered in this work as shown in Figure 2 are; (i) The driving unit (petrol engine) (ii) The transmission unit (gear Unit) (iii) The compressor s components which include the cylinder, piston, metal ring, connecting rod, bearing, cylinder crankshaft, value and spring value. Procedures discussed in earlier works of machine design as well as other design principles [4, 6, 7, 11, 16] are also of relevance to this work. Conceptualized and detailed designs mentioned in other works [9, 10] were done numerically prior to development of software algorithm. The aspects of the component design that make up the complete power transmission design for the air compressor drive train include force, moment and torque requirements, shaft dimensions, material selection, fatigue and stress effects. The maximum (T max ) and the minimum (T min ) torque generated by the crank within the 77 * Correspondence Author s Address: University of Ibadan, Department of Mechanical Engineering, Faculty of Technology, Nigeria, esimmar@yahoo.com

40 Emmanuel Simolowo, Taiwo Olumide: Design Analysis Algorithm of a Petrol Engine-Powered Air Compressor; Machine Design, Vol.5(2013) No.1, ISSN ; pp Fig.1. Petrol engine-powered air compressor compressor are obtained from the analysis of Torque time relation available in literature [14]. To commence the design, certain parameters have to be assumed ad later adjusted to meet up with the design requirements. They are gear pinion diameter (d p ), Gear output diameter (d g(o) ), pressure angle (θ), ball bearings of standard diameters are to be used Force Requirements bending and torsion at that critical location where both moment and torque components are largest. After the gears are designed and the bearings are selected, the design can be redefined to include stepped shoulders and use more accurate stress concentration factors. The mean (M m, T m ) and alternating (M a, T a ) components of both moment and torque are obtained using equations (7) and (8). The loading is combination of fluctuating moment and fluctuating torque. The maximum (F tmax ) and minimum forces (F tmin ) generated by the crank mechanism of the compressor (Figure 3) and transmitted to the output shaft is obtained using equation (1) where (r g ) is gear radius (7) (8) The gear radius is used in equation (1) and (2) because it is attached to compressor which produces the torque. The maximum (F max ) and minimum (F min ) resultant forces are calculated using equations (3) and (4). The maximum (M max ) and minimum (M min ) moments on the shaft based on the design criteria that the gears are centered between the simple supported bearings L mm apart, are determined by equations (5) and (6) 32 (1) (2) (3) (4) (5) (6) Since a keyway will probably be needed at the gear, a stress concentration factor of 3[14] is applied for both 2.2. Material Selection (9) (10) Considering a design based on Low carbon inexpensive cold rolled steel, the following reference equations (11) and (12) are used. Where, S ut is ultimate tensile strength, S y is yield strength ,400 (11) 689 1,400 (12) From eq. (11) Se is the endurance strength and Sf is the fatigue strength. This must be reduced by various factors to account for difference between the part and the test specimen and to obtain the corrected endurance strength (Se)in equation (13). (13) (14)

41 Emmanuel Simolowo, Taiwo Olumide: Design Analysis Algorithm of a Petrol Engine-Powered Air Compressor; Machine Design, Vol.5(2013) No.1, ISSN ; pp Loading is bending and torsion where C load, C size, C surf, C relial, C temp are all stress concentration factors based on load type, part size, type of surface finish, reliability and temperature of operation respectively. The coefficient A and exponential b corresponding to S ut are available in literature [17]. The notch sensitivity is obtained based on the neuler s constant (a) as expressed in equation (15) for SI unit. The value of (a) used in equation (15) is also obtained from standard tables for various values of S ut and r is the assumed notch radius.. (15) The fatigue stress concentration factor (k f ) is given in equations (16) and (17) where K t is geometric stress concentration factor for the case to be designed. The coefficient A and exponential b corresponding to dimension ratios are available in literature [15]. 1 1 (16) K t b r A (17) d 2.3. Power Transmission The output shaft diameter (ds (o) ) is given by equation (19), where K f is fatigue stress factor, K fsm is stress concentration factor for mean shear stress, K fs is stress concentration factor for shear stress, K fm is stress concentration factor for mean, N sf is the factor of safety. The values of the concentration factors are obtained based on design situation. For this design the same factors are used on the mean and shear stress components (eq. 18). The input diameter (ds (i) ) is obtained using equation (20) K fm = K f ; K fsm = K fs (18) (19) (20) accommodate some angular and parallel misalignment due to tolerance in the mounting of these subassemblies of engine gearbox and compressor. 3. DEVELOPMENT OF LEARNING ALGORITHM The development of the learning algorithm follows steps similar to other works [18, 19, and 20]. Other relevant works are on tools and methodology for software design of tribo-elements [5,8,12] The main stages involved in the software development of this work are (i) selection of appropriate programming language (ii) creation of user friendly and interactive interfaces (iii) development of program codes (iv) packaging and deployment. The ideal choice of programming language for the development of the software algorithm used in this work is the C ++. The choice was made based on the following (i) high program execution speed (ii) adaptability (iii) the relative ease with which it allows a user to create interfaces. Once the algorithm was fully developed a complete evaluation was conducted by validating the software results with those obtained in literature Algorithm Structure The first step in building program algorithm is to decide on the program structure. Shown in Figure 2 is the program flow chart. The program structure consists of two modules; (i) study session (ii) practice session. The study session involves a systematic and sequential display of files to the user. The session starts with the display of the first interface and enables proper navigation through the study material. The practice session is the largest component of the software algorithm and consists of iterative operations such as the collection of data, computation execution and return of computed values for further operations. The overall program flow of the design algorithm is characterized by the following stages (i) design- input entry (ii) design- path selection by user (iii) results display. Some of the unique features of the program include (i) Multi-purpose structure: This feature is as result of the various sessions embedded in the program. (ii) Multiple design ability: The program enables a maximum of four different design scenarios to be performed using different interfaces as shown in Figure 8. The coupling suggested for use in this design is of complaint type (jar type coupling). This is to 33

42 Emmanuel Simolowo, Taiwo Olumide: Design Analysis Algorithm of a Petrol Engine-Powered Air Compressor; Machine Design, Vol.5(2013) No.1, ISSN ; pp , 6 and 7 are the results of the validation of the software using the case study considered. Fig.3. Assembly of Case Study Compressor train Fig.2. Flow chart of software algorithm 4. SOFTWARE VALIDATION AND RESULTS The software algorithm developed in this work was validated using a design case study presented in literature [14] The design case focuses on the need by a contractor for a small gasoline-engine powered air compressor to drive air hammers. The design process begins with conceptualization and creativity where the following criteria are taken into consideration. (i) Capacity of gasoline engine depending on the type of air hammer to be used. (ii) Type of compressor needed. (iii) General layout for the design. (iv) Machine elements involved. (v) Compressor air flow rate and mean effective pressure. After these design criteria have been properly analyzed, the following components and specifications were used in the design [14]. (i) 2.5 hp single-cylinder, two stroke engine with flywheel at 3,800rpm. (ii) Gear set (ratio 1,500rpm : 3,800rpm or 0.39 : 1). (iii) Compressor (4x10-4 stroke volume) rev = 1,500rpm. (iv) Shafts, couplings, bearings and gears. Shown in Figures 3 and 4 are the layout of the compressor assembly and stages of design. The results obtained by applying the procedures discussed in section 2 are presented in Tables 1. Shown in Figures Fig.4. Design Stages for Compressor Assembly 34

43 Emmanuel Simolowo, Taiwo Olumide: Design Analysis Algorithm of a Petrol Engine-Powered Air Compressor; Machine Design, Vol.5(2013) No.1, ISSN ; pp Table 1. Case Study Design for Software s Validation S/N Design Parameters Software Results Ftmax (N) Ftmin (N) specifications 1. T max 66.5 Nm 2. T min -20 Nm 3. dp (i) mm 4. dp (o) 254 mm 5. t g 50 mm 6. Θ 20 o 7. N sf 3 Force Requirement 8 F tmax N 9 F tmin -157 N 10 F max N 11 F min -167 N 12 L 100 mm 13 M max Nm 14 M min -418 Nm 15 M m Nm 16 M a 9.05 Nm 17 T m Nm 18 T a Nm Materials Specifications and Design Fmax (N) Status of Parameter Assumed values based on conceptualizat ion stage Results obtained 19 Selection Material Steel based on selected formability 20 S ut MPa Standard 21 S y MPa values 22 Surface Finish Machined Fabrication process 23 a Based on S ut from tables 24 r 0.254m Assumed 25 q Result Obtained Power Transmission Shafting 26 K t 3 27 K f 2 28 d s(o) m 29 d s(i) 0.02 m Result obtained Numerical Results Fmin (N) L (mm) Fig.5. Comparison of results for Force Requirements q ( (x 10 1 ) Software Results Fig.6. Results for Torque Analyses Numerical Results ds(i) (m) ds(o) (m) Figure 7: Results for Power Transmission Shafting 5. CONCLUSION Numerical Results Software Results 0 0,02 0,04 0,06 The objective of this work which is to commence the development of comprehensive learning software for the design of petrol engine powered air compressors has been achieved. This work shows that the use of computation and design interactive software constitutes an improvement on conventional teaching methods. The interactive software with the aid of appropriate features takes the user through critical design concepts and offers a proper understanding of the design principles of a petrol-engine-powered compressor. The validation of the software using a design case study in literature proved the reliability of the software. This work opens further tasks on various types of engine-powered air compressor systems by adapting and extending the algorithm potentialities of the learning software developed herein. 35

44 Emmanuel Simolowo, Taiwo Olumide: Design Analysis Algorithm of a Petrol Engine-Powered Air Compressor; Machine Design, Vol.5(2013) No.1, ISSN ; pp Fig.8. Software Interface showing four different design sessions REFERENCES [1] BILJANA, G., DESNICA, E., SUBIC, N., New Trends in Technology And Higher Education, Engineering Technical Professions (E-Learning) Machine Design, 2011, Vol.3, No.4, ISSN , pp [2] CHAPLIN, C., Creativity in Engineering Design Fellowship of Engineering Publishers, London, [3] DHILLON, B. S., Engineering Design: A modern Approach, Richard Irwin, Homewood, Illinois [3] DHILLON, B. S., Engineering Design: A modern Approach, Richard Irwin, Homewood, Illinois [4] FRENCH, J., Invention and evolution: design in nature and engineering, Cambridge University Press, United Kingdom, [5] GEORGIA N., SOTIRIOS S., IOANNIS K., A software tool for parametric design of turbo machinery blades. Advances in Engineering Software, 2009, Vol. 40, No. 1, pp [6] HEINZ, P., A Practical Guide to Compressor Technology, John Wiley, New Jersey, [7] HEINZ, P., and ARVIND G., Compressors and Modern Process Application, John Wiley, New Jersey, [8] HONG Z., Software Design Methodology, Butterworth Heinemann, [9] JON, B., MCCRACKEN, B., The detail design phase, SEED, Sharing Experience in Engineering Design, [10] JON, B., SMITH, D., The conceptual design phase, SEED, Sharing Experience in Engineering Design, [11] KHURMI, R., GUPTA, K., Theory of Machines, S. Chand publishers, Delhi 2002 [12] MAÍRA, M., OLIVIER, B., JAN, S., WIM, D., HENDRIK, V., Computer-aided integrated design for machines with varying dynamics, Mechanism and Machine Theory, 2009, Vol. 44, No. 9, pp [13] NORTON, L., Machine design, an integrated approach, Prentice Hall, USA, 2006 [14] NORTON, L., Machine design, an integrated approach, Prentice Hall, USA, [15] PETERSON, R., Stress Concentration factors, John Wiley & Sons, New Jersey,1975. [16] ROLLASON, E., Metallurgy for Engineer, Edward Arnold, London, [17] SHIGLEY, J.E. and MISCHKE, C.R., Standard handbook of machine design, McGraw Hill, New York, [18] SIMOLOWO, E., BAMIRO, O., Roller-Cams Systems Design: Development of a profile Analysis Software, Pacific Journal of Science and Technology, 2009, Vol. 10 No. 1 pp [19] SIMOLOWO, E., BAMIRO O., Software Developed for Polynomial Cams, The NSE. Technical Transaction. 2010, Vol. 45, No. 3 pp [20] SIMOLOWO E., BAMIRO O., The Development of An Analysis Intensive Software for Improved Cam Systems Design, Journal of Science and Technology, 2008,Vol. 28. No 1 pp

45 machine design, Vol.5(2013) No.1, ISSN pp Preliminary note THE INFLUENCE OF THE INJECTION PROCESSING TEMPERATURE ON THE THERMAL STABILITY OF POLYMERS USED IN THE MANUFACTURE OF ITEMS IN THE AUTOMOTIVE AND SPORTS INDUSTRY Gheorghe Radu Emil MĂRIEŞ 1, * 1 University of Oradea, Faculty of Fine Arts, Oradea, Romania Received ( ); Revised ( ); Accepted ( ) Abstract: The object of the present study is the analysis of the variation in thermal stability (the vitrification temperature, the melting point, the flowing temperature) depending on the processing temperature for polymers used in the manufacture of various technical items in the automotive and sports industry. The studied polymers were polyamide (PA), thermoplastic polyurethane (TPU), polyoxymethylene (POM) and polypropylene (PP). Key words: polymers, thermal stability, differential scanning calorimetry (DSC) 1. INTRODUCTION 1.1 Generalization Polyamides, thermoplastic polyurethanes, polyoxymethylenes and polypropylene are macromolecular composites frequently used in the manufacture of various technical items in the automotive and sports industry. In the automotive industry, polyamides are used for casings, ventilators, complicated construction parts, fuel tanks, bushes, pivots, flexible wiring, floats and brake fluid reservoirs. Polyoxymethylenes are used for cogwheels, slideways, active parts of oil or diesel pumps, etc. Thermoplastic polyurethanes are used in the automotive industry for air grilles, sealing rings, side skirts, connecting bellows, flexible and resistant air cooling systems, oil- and vaseline-resistant protective rings for anti-roll bar ends, protective bellows; polypropylene is used in the manufacture of bumpers and car door guards. In the sports industry, polyamides are used in the manufacture of inline skates, tents, ropes used in climbing and speleology, crash helmets, etc. [1]. POM is used in the manufacture of rear derailleur jockey wheels, inline skates parts, etc. Thermoplastic polyurethanes are used in the manufacture of sports equipment (ski boots, ski tip protectors, sledges, crossbows, etc.)[1,2,3]. Polypropylene is used in the manufacture of sports shoes innersoles [4], football boots crampons, bicycle pedals, height adjustment elements in push scooters, etc. All these polymers are used both as standalone materials and reinforced with various reinforcing agents: glass fibre, carbon fibre, molybdenum disulfide, silicon dioxide, metallic fibres, mica, etc. [5-8]. Polyamides, thermoplastic polyurethanes, polyoxymethylenes and polypropylene are often used in the automotive and sports industry due to the following properties [1,9-23]: good dimensional stability; good mechanical resistance; good rigidity; excellent fatigue resistance; great impact strength; abrasion and wear resistance; good balance between rigidity and elasticity; easy to process; their fine surface Thermal analysis methods The terminology recommended by the International Confederation for Thermal Analysis and Calorimetry (ICTAC) [24] is used further on. According to this terminology, thermal analysis refers to a group of techniques by means of which, dependent on time or temperature, a physicochemical property of the investigated sample is monitored, when the temperature of the sample, in a certain atmosphere, is programmed. The programming may imply the increase or decrease of the temperature according to a variation function during a period of time (temperature programme) or maintaining the temperature constant. An extensive enumeration of the thermal analysis techniques is given in Table 1 [25]. The most frequently used techniques are TGA, DTA, DSC and DMA. These techniques are used for characterizing technical polymers [26,27,28]. The differential thermal analysis (DTA) consists of measuring and recording the difference T between the temperature of the sample to be investigated and the temperature of an inert reference material, given that the oven temperature observes a certain programme. The DSC technique is similar to the DTA technique, with the exception that this technique measures the difference in heat flux between the sample to be investigated and the reference sample. The main physical and chemical processes detected by the DTA and DSC techniques are summarised in Table 2. Moreover, the DTA and DSC techniques can also signalize processes of thermal and thermo-oxidative degradation of organic and inorganic materials. Thermal *Correspondence Author s Address: University of Oradea, Faculty of Fine Arts, Str. P-ta Independentei nr. 39, , Oradea, Bihor, Romania, maries.radu@rdslink.ro

46 Gheorghe Radu Emil Mărieş: The Influence of the Injection Processing Temperature on the Thermal Stability of Polymers Used in the Manufacture of Items in the Automotive and Sports Industry; Machine Design, Vol.5(2013) No.1, ISSN ; pp degradation is the decomposition of a material in composites with a smaller molecular mass and it is characterized by endothermic peaks. Thermo-oxidative degradation refers to the material s interaction with oxygen and it is characterized by exothermic peaks. Through DTA and DSC, one can evaluate the domain of chemical stability of a material, i.e. the domain of temperature in which, by means of a rapid analytical technique (such as DTA and DSC), a process of thermal or thermo-oxidative degradation is not outlined. Each material has a specific DTA and DSC curve. That is why each of these materials may be regarded as a thermal spectroscopy used to identify a certain material. Applying this method of material identification requires a standardization of the thermal analysis methods [25]; this task proves to be quite difficult because of the influence of operational factors on the results of the DTA and the DSC. The present study aims to analyse, by means of the DSC thermal analysis method, the variation in thermal stability (the vitrification temperature, the melting point, the flowing temperature), depending on the injection processing temperature, for polymers which are frequently used in the manufacture of various technical items in the automotive and sports industry. 2. THE EXPERIMENTAL PART 2.1. The polymeric materials used The following types of polymers were used in the injection of the samples necessary to the experimental part: Polyamide 6.6, type TECHNYL A 221 (Dry), Thermoplastic polyurethane, type Desmopan KA 8377, Polyoxymethylene, type Tenac 2013A, Polypropylene, type Homopolymer Resin 100-GBO6. All four types of technopolymers are polymers without fillers Equipment The equipment used in the processing of the polymers was a horizontal injection machine (fig.1). Name of the machine: ENGEL Version: G/11/10/116/3 Manufacture year: 2002 Type: VC 500/110 TECH Manufacturer: Engel Maschinenbau Gesellschaft m.b.h A-4311 Schwertberg [29]. The temperature measurement of the flowing material was recorded by means of a DYNISCO thermocouple, model Ti422J, installed in the nozzle of the injection cylinder so as to record the real temperature of the melted polymer flux. The equipment used for the differential scanning calorimetry (DSC) analysis of the four technopolymers (PA 6.6, TPU, POM and PP) was: a type NETZSCH DSC 204 device, manufacturing company: NETZSCH, manufacture year: 2000 The differential scanning calorimetry (DSC) showed, for each polymer, the influence of the processing temperature variation on the vitrification temperature, the melting point and the flowing temperature of the polymer. Fig.1. Injection machine ENGEL VC 500/110 TECH 3. RESULTS AND DISCUSSION 3.1. The differential scanning calorimetry (DSC) of polyamide 6.6, type TECHNYL A 221 (Dry) The following real injection temperatures were used in order to obtain the specimens: 270 C, 285 C, 300 C, 315 C, 330 C and 345 C. All other parametres which influence the injection cycle remained constant throughout the injection of the specimens. The specimens obtained this way were subjected to the DSC thermal analysis. In some cases, unprocessed granules were used for the measurements, in order to trace the changes in the material during the injection. The DSC measurements were recorded in a nitrogen atmosphere, using the following temperature scheme: temperature increase from 0 to 270 C by 10 K/min, temperature decrease to -100 C by 10 K/min, isothermal mode at -100 C for 5 minutes, temperature increase to 400 C by 5 k/min. Table 3 shows the values of the vitrification temperature T v (inflection point on the DSC curve) and the melting point T m (the endothermic peak on the DSC curve) depending on the six processing temperatures. It can be observed that T v varies between 59.6 C and 60.6 C, and T t varies between C and 262 C, the modifications are therefore non-essential, having the same value as those of the unprocessed granules. In conclusion, the processing temperature does not influence the transition temperatures of polyamide 6.6, type Technyl A The differential scanning calorimetry (DSC) analysis of thermoplastic polyurethane, type Desmopan KA 8377 The following real injection temperatures were used in order to obtain the specimens: 200 C, 210 C, 220 C, 230 C and 240 C.. All other parametres which influence the injection cycle remained constant throughout the injection of the specimens. The specimens obtained this way were subjected to the DSC thermal analysis. In some cases, unprocessed granules were used for the measurements, in order to trace the changes in the 38

47 Gheorghe Radu Emil Mărieş: The Influence of the Injection Processing Temperature on the Thermal Stability of Polymers Used in the Manufacture of Items in the Automotive and Sports Industry; Machine Design, Vol.5(2013) No.1, ISSN ; pp material during the injection. The DSC measurements were recorded in a nitrogen atmosphere, using the following temperature scheme: temperature increase from 20 to 200 C by 10 K/min, temperature decrease to -100 C by 10 K/min, isothermal mode at -100 C for 5 minutes, temperature increase to 400 C by 5 k/min. Table 4 shows the values of the vitrification temperature T v and the flowing temperature T f (the endothermic peak on the DSC curve) depending on processing temperature. It can be observed that T v increases very little with the increase in processing temperature, and T f is insignificantly influenced by the processing temperature The differential scanning calorimetry (DSC) analysis of polyoxymethylene, type Tenac 2013A The following real injection temperatures were used in order to obtain the specimens: 180 C, 190 C, 200 C, 210 C, 220 C, 230 C şi 240 C. All other parametres which influence the injection cycle remained constant throughout the injection of the specimens. The specimens obtained this way were subjected to the DSC thermal analysis. In some cases, unprocessed granules were used for the measurements, in order to trace the changes in the material during the injection. The DSC measurements were recorded in a nitrogen atmosphere, using the following temperature scheme: temperature increase from 0 to 180 C by 10 K/min, temperature decrease to 0 C by 10 K/min, isothermal mode at 0 C for 5 minutes, temperature increase to 400 C by 5 k/min. Table 5 shows the values of the vitrification temperature T v (inflection point on the DSC curve) and the melting point T m (the endothermic peak on the DSC curve) depending on the processing temperature. It can be observed that T v varies between C and C, and T t varies between C and C, the modifications are therefore non-essential. In conclusion, the processing temperature has very little influence on the transition temperatures of the polyoxymethylene, type Tenac 2013 A The differential scanning calorimetry (DSC) analysis of the polypropylene homopolymer, type INNOVENE 100-GBO6. The following real injection temperatures were used in order to obtain the specimens: : 220 C, 240 C, 260 C, 280 C şi 300 C. All other parametres which influence the injection cycle remained constant throughout the injection of the specimens. The specimens obtained this way were subjected to the DSC thermal analysis. In some cases, unprocessed granules were used for the measurements, in order to trace the changes in the material during the injection. The DSC measurements were recorded in a nitrogen atmosphere, using the following temperature scheme: temperature increase from 20 to 220 C by 10 K/min, temperature decrease to -100 C by 10 K/min, isothermal mode at -100 C for 5 minutes, temperature increase from -100 to 400 C by 5 k/min. Table 6 shows the values of the melting point T m (the endothermic peak on the DSC curve) for PP homopolymer, type INNOVENE 100-GBO6, depending on the processing temperature. It can be observed that T t varies between C and C, the modifications are therefore non-essential. In conclusion, the processing temperature has very little influence on the melting point of polypropylene homopolymer, type INNOVENE 100-GBO6. 4. TABLES Table 1. Thermal analysis techniques [25] Crt. no. Technique Abbreviation Property Use 1. Thermogravimetric analysis TGA 2. Differential thermal analysis DTA 3. Differential scanning calorimetry DSC Mass Difference in temperature Heat flux 4. Thermomechanical analysis TMA Deformation 5. Dynamic mechanical analysis DMA Modulus 6. Dielectric thermal analysis DETA Permittivity 7. Exhaust gas analysis EGA Gas 8. Thermophotometry Optical properties Chemical decompositions Oxidations Dehydrations Phase transitions Chemical reactions Heat capacity Phase transitions Chemical reactions Calorimetry Mechanical changes Expansion, contraction Phase transitions Polymer reticulation Phase transitions Polymer transformations Decompositions Catalytic reactions or reactions on surfaces Phase transitions Reactions on surfaces Colour changes 39

48 Gheorghe Radu Emil Mărieş: The Influence of the Injection Processing Temperature on the Thermal Stability of Polymers Used in the Manufacture of Items in the Automotive and Sports Industry; Machine Design, Vol.5(2013) No.1, ISSN ; pp Crt. no. Technique Abbreviation Property Use 9. Thermosonimetry TS Sound Mechanical and chemical processes 10. Thermo-magnetometry TM Magnetic Variation of the magnetic properties Curie points 11. Thermoluminescence TL Light emission Material flaws 12. Emanation thermal analysis ETA Gas emissions Structural changes 13. Simultaneous thermal analysis STA One or several simultaneous techniques 14. Controlled rate thermal analysis CRTA Table 2. Physical and chemical processes detected by DTA and DSC [25] Constant variation rate of the properties The kinetics of the chemical processes Phenomenon Endothermic heat effect Exothermic heat effect Physical Melting + Crystallisation + Vaporisation + Sublimation + Glass transition There is no heat effect, only a There is no heat effect, only a change in baseline change in baseline Crystalline transitions + + Chemical Dehydration + Decomposition + (+) Oxidative degradation + Solid phase reactions + + Combustion + Polimerization + Reticulation + Table 3. The vitrification temperature and the melting point of PA processed at various temperatures The processing temperature of PA [ C] T v [ C] T m [ C] Granules Table 4. The vitrification temperature and the melting point of TPU processed at various temperatures Processing temperature of TPU [ C] T v [ C] T f [ C] Granules

49 Gheorghe Radu Emil Mărieş: The Influence of the Injection Processing Temperature on the Thermal Stability of Polymers Used in the Manufacture of Items in the Automotive and Sports Industry; Machine Design, Vol.5(2013) No.1, ISSN ; pp Table 5. The vitrification temperature and the melting point of POM processed at various temperatures Processing temperature of POM [ C] T v [ C] T m [ C] Granules Table 6. Melting points of PP processed at various temperatures Processing temperature of PP [ C] T m [ C] Granules CONCLUSION The following polymers were analysed: Polyamide 6.6, type TECHNYL A 221 (Dry) Thermoplastic polyurethane, type Desmopan KA 8377, Polyoxymethylene, type Tenac 2013A, Polypropylene, type Homopolymer Resin 100-GBO6. Following the analysis, it can be stated that, for these four polymers, the injection processing temperatures do not influence the vitrification temperature, the melting point and the flowing temperature in a significant way. The largest variation in vitrification temperature and melting point (from 2 C to 3 C) is recorded for the thermoplastic polyurethane, type Desmopan KA REFERENCES [1] Mărieş, Gh., R., E., Mărieş, I., Study on the properties of some thermoplastic polymers recommended in manufacturing power transmission components for engineering and automotive industries, Proceedings - The 3 rd International Conference Power Transmissions `09, Pallini Beach Hotel, Kallithea, Greece, 1-2 october 2009, Editor: Athannassios Mihailidis, pp [2] Salm, W., Nur Fliegen ist Schoner, TPU in neuen Kombinationen fur Sport und Spiel, Kunststoffe, 7, 1999, p.118. [3] Scholl, A., Piroddi, A., Sicher und Gerauschdampfend, Kunststoffe, 4, 1998, p.556. [4] Guichard, A., RocTool, l`induction pour chaussure de sport, Plastiques & Caoutchoucs Magazine, Nr.831, 05, 2005, p.12. [5] Trotignon, J., P., Verdu, J., Dobracginsky, A., Piperaud, M., Matieres Plastiques. Structuresproprietes, Mise en oeuvre, Normalisation, Editions Nathan, Paris, 1996, p [6] Manoviciu, V., Mărieş, Gh., R., E., Materiale compozite cu matrice organică, Editura Universităţii Oradea, Oradea, [7] Jacksch, E., Chetaru, D., Manca, Gh., Materiale plastice poliamidice, Editura Tehnică, Bucureşti, [8] Manoviciu, I., Chimia compuşilor macromoleculari, Institutul Politehnic Traian Vuia, Timişoara, 1979, p [9] DuPont De Nemours, Product & Market Literature, DuPont Engineering Polymers, From Concept to Commercialization, 11, [10] DuPont De Nemours, Product & Market Literature, DuPont Engineering Polymers, DuPont Minlon and Zytel, nylon resins, Design Information-Module II, 05,

50 Gheorghe Radu Emil Mărieş: The Influence of the Injection Processing Temperature on the Thermal Stability of Polymers Used in the Manufacture of Items in the Automotive and Sports Industry; Machine Design, Vol.5(2013) No.1, ISSN ; pp [11] ***, BASF, Ultramid/Capron, Polyamide (PA), BASF Plastics key to your success, BASF The Chemical Company, 09, 2004, p [12] ***, BASF, Ultramid, Polyamide (PA), BASF Plastics key to your success, BASF The Chemical Company, 09, 2004, p [13] DuPont De Nemours, Product & Market Literature, DuPont Engineering Polymers, Delrin, acetal resin, Product guide and properties, 04, 2003, prin /Mediator?common=2,477,478. [14] DuPont De Nemours, Product & Market Literature, DuPont Engineering Polymers, Delrin, acetal resin, Design Information, 02, 2003, prin /Mediator?common=2,477,478. [15] DuPont De Nemours, Product & Market Literature, DuPont Engineering Polymers, Delrin, acetal resin, Desig whith Delrin, 2003, prin /Mediator?common=2,477,478. [16] ***, BASF, Ultraform, Polyoxymethylene (POM), BASF Plastics key to your success, BASF The Chemical Company, 09, 2004, p [17] ***, BASF, Elastollan Material Properties, Elastogran BASF Group, 02, 1990, p.9. [18] Brenner, E., Polypropylene an Alternative?, Kunststoffe, 4, 2000, p.35. [19] Mărieş, Gh., R., E., Contribuţii la studiul unor caracteristici fizice ale polimerilor, utilizabili în articole sportive de performanţă, prin metode termice, Editura Politehnica, Timişoara, [20] Mărieş, Gh., R., E., Manoviciu, I., Bandur, G., Rusu, G., Pode, V., Study by Thermal Methods of Some Physico-mechanical Properties of Polyamides Used for High Performance Sport Products, Materiale Plastice, Vol.46, nr.1, 2009, pp [21] Mărieş, Gh., R., E., Thermal analysis of some mechanical and physical properties of thermoplastic polyurethanes used in manufacturing of performance sport products, Materiale Plastice, Vol.46, nr.2, iunie 2009, pp [22] Mărieş, Gh., R., E., Thermal Analysis of Some Mechanical-Physical Properties of Polyoxymethylenes (POM) used for Manufacturing of Performance Sport Products, Materiale Plastice, Vol.47, nr.2, iunie 2010, pp [23] Mărieş, Gh., R., E., Thermal Analysis of some Physico-Mechanical Properties of Polypropylene (PP) Used for Manufacturing of Performance Sport Items, Materiale Plastice, Vol.47, no.4, December 2010, pp [24] Hess, W., M., Klamp, W., K., Rubber Chem. Technol. 556, 1983, p.390. [25] Sestak, J., Berggren, G., Thermochimica Acta, 3, 1971, p.495. [26] Knappe, St., Mayo, C., Thermische Analyse, Kunststoffe, 12, 1995, p [27] Utschick, H., Nitschke, P., Thermische Analyse Neue TTechniken-neue Anwendungen, Kunststoffe, 11, 1998, p [28] Frick, A., Qualitatssicherung mittels Thermoanalyse, Kunststoffe, 08, 2001, p.149 [29] Urbanek, O., Neumann, P., Schonberger, H., ENGEL, Machine a injecter, Manuel d`instructions, Version:G/11/10/116/3, Cartea tehnica a masinii de injectie ENGEL, Versiune:G/11/10/116/3, Schwertberg, Austria, 2001, p

51 machine design, Vol.5(2013) No.1, ISSN pp STRUCTURAL OPTIMIZATION OF MINI HYDRAULIC BACKHOE EXCAVATOR ATTACHMENT USING FEA APPROACH Research paper Bhaveshkumar P. PATEL 1, * - Jagdish M. PRAJAPATI 2 1 Mechanical Engineering Department, U. V. Patel College of Engineering, Ganpat University, Ganpat Vidynagar , Dist. Mehsana, Gujarat, India. 2 M. S. University of Baroda, Associate Professor, Faculty of Technology and Engineering, Vadodara , Gujarat, India. Received ( ); Revised ( ); Accepted ( ) Abstract: Excavators are heavy duty earthmoving machines and normally used for excavation task. During the excavation operation unknown resistive forces offered by the terrain to the bucket teeth. Excessive amount of these forces adversely affected on the machine parts and may be failed during excavation operation. Design engineers have great challenge to provide the better robust design of excavator parts which can work against unpredicted forces and under worst working condition. Thus, it is very much necessary for the designers to provide not only a better design of parts having maximum reliability but also of minimum weight and cost, keeping design safe under all loading conditions. Finite Element Analysis (FEA) is the most powerful technique for strength calculations of the structures working under known load and boundary conditions. FEA approach can be applied for the structural weight optimization. This paper focuses on structural weight optimization of backhoe excavator attachment using FEA approach by trial and error method. Shape optimization also performed for weight optimization and results are compared with trial and error method which shows identical results. The FEA of the optimized model also performed and their results are verified by applying classical theory. Key words: Digging Forces, Autonomous Excavation, Resistive forces, Heaped capacity 1. INTRODUCTION In the era of globalization and tough competition the use of machines is increasing for the earth moving works, considerable attention has been focused on designing of the earth moving equipments [10]. Today hydraulic excavators are widely used in construction, mining, excavation, and forestry applications [1]. The excavator mechanism must work reliably under unpredictable working conditions. Poor strength properties of the excavator parts like boom, arm and bucket limit the life expectancy of the excavator. Therefore, excavator parts must be strong enough to cope with caustic working conditions of the excavator [8]. But in contradictory, now a day weight is major concern while designing the machine components. So for reducing the overall cost as well as for smoothing the performance of machine, optimization is needed. Structural design has always been a very interesting and creative segment in a large variety of engineering projects. Structures, of course, should be designed such that they can resist applied forces (stress constraints), and do not exceed certain deformations (displacement constraints). Moreover, structures should be economical. Theoretically, the best design is the one that satisfies the stress and displacement constraints, and results in the least cost of construction. Although there are many factors that may affect the construction cost, the first and most obvious one is the amount of material used to build the structure. Therefore, minimizing the weight of the structure is usually the goal of structural optimization [7]. There are many methods can be applied for the optimization problems like, Linear Programming (LP), Non-Linear Programming (NLP), Integer Linear Programming (ILP), and Discrete Non-Linear Programming (DNLP) and However, some newly developed techniques, known as heuristic methods, provide means of finding near optimal solutions with a reasonable number of iterations. Included in this group are Simulated Annealing, Genetic Algorithms, and Tabu Search [7]. Finite Element Analysis is the powerful technique for calculation of the strength of structure under known working load and boundary conditions [6]. Finite Element Analysis (FEA) is also one of the best powerful methods which can be applied for structural weight optimization. There are so many works done by other researchers in the field of FEA and optimization of the backhoe excavator machines which are covered in the reference paper of [2]. For our case we have adopted FEA approach for performing structural weight optimization of mini hydraulic backhoe excavator attachment using trial and error method and shape optimization method. 2. BACKGROUND OF WORK Based on the market survey and reverse engineering and authors expertise in the field of design a 3D model of mini hydraulic backhoe excavator attachment is developed using the Autodesk Inventor professional The resistive forces offered by the terrain to the bucket teeth are found by applying the fundamental knowledge of soil mechanics and McKyes and Zeng models utilized to find soil-tool interaction forces [3]. The developed * Correspondence Author s Address: Mechanical Engineering Department, U. V. Patel College of Engineering, Ganpat University, Ganpat Vidyanagar , Kherva, Dist. Mehsana, State-Gujarat, India, bppmech@gmail.com

52 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp resistive forces must be less than that of the digging forces offered by the actuators. Maximum resistive forces offered by the ground for the proposed tool dimensions is Newton, and the breakout force calculated is 7626 Newton which is higher than the forces required to cut the soil ( Newton), thus this calculated breakout force is adequate and accepted for the job to be performed by the proposed mini backhoe excavator i.e. light duty construction work [3]. The digging force calculations carried out based on SAE standards of SAE J1179 and static force analysis performed for maximum breakout force condition considering static equilibrium. The calculated bucket curl or breakout force F B = Newton, and calculated arm crowd force or digging force F S = Newton [4]. Finite Element Analysis also performed on mini hydraulic backhoe excavator attachment for the purpose of verification of part s strength. The results shows that the developed stresses are far less than that of the designed stress limit [5]. Therefore, there is a scope to perform weight optimization for backhoe attachment using FEA approach based on strength criterion. Here, we have consider thickness of plates as a variable and adopted trial and error method to get optimized backhoe excavator model based on standard available limiting value of plate thickness and limiting safe stress criterion. The optimized model also checked for limiting safe stress and the results verified by applying classical theory. The materials used for the different components are made from HARDOX400 [11], SAILMA 450HI [12] and IS 2062 [9]. Structural optimization is performed for the bucket, arm, boom and swing link which are covered one by one in next coming sections using ANSYS software. 3. OPTIMIZATION OF BUCKET optimized model of bucket is analyzed to check that the optimized model is within safe limit or not. Fig.2. Static force analysis of bucket Fig.3. Boundary conditions for bucket Fig.4. Maximum stresses of the optimized bucket Fig.1. Modified bucket for optimization The Fig. 1 shows the bucket with different parts which are modified to get optimum dimensions based on available standard thickness of plates. Table 1 shows the name of the parts of the bucket which are modified to get the weight optimized model. It also shows the dimensions and total weight of the parts before modification and after modifications. The total weight of the bucket is kg and after modification we got the optimized weight of the bucket is kg. Therefore, we achieved kg reduction in the weight of the bucket. Based on the known boundary conditions calculated as in reference [4], the Fig.5. Maximum displacement of optimized bucket 44

53 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Table 1. Optimization data of the bucket Part no. Part name Quantity Thickness before modification Modifications Thickness after modification Total weight (Kg) Weight before optimization Weight of optimized model 1 Base plate Side protector Bucket top plate Side shear plate Bucket mounting lug Bucket mounting lug bush Fig. 2 shows the static force analysis of the bucket for maximum breakout force condition. Fig. 3 shows boundary conditions applied to bucket for analysis purpose. Design stress for ductile materials, condition. Fig. 8 shows the boundary conditions applied to arm for the purpose of analysis. Fig. 9 shows the results of the Von Misses stresses on optimized arm assembly at the arm cylinder mounting lug and it is MPa. σ σ (1) The maximum Von Misses stress is acting at the end of the mounting lugs as shown in Fig. 4, which is made up of Hardox400 material with the yield strength of 1000 MPa, by taking safety factor as 2, equation (1) yields = MPa, [σ y ] = 500 MPa, this clearly indicates σ VM < [σ y ], so the design of the optimized bucket is safe for strength. Fig. 5 shows the maximum displacements on the bucket of mm which is very small compare to minimum thickness of the plate used in the bucket, therefore it is safe for deflection. 4. OPTIMIZATION OF ARM The failure criterion states that the Von Misses stress σ should be less than the yield stress σ of the material by taking appropriate safety factor into consideration. This indicates for the design of a part to be safe, the condition shown in equation (1) must be satisfied [13]. The Fig. 6 shows the arm with different parts which are modified to get optimum dimensions based on available standard thickness of plates. Table 2 shows the name of the parts of the arm which are modified to get the optimized model. It also shows the dimensions and total weight of the parts before modification and after modifications. The total weight of the arm is kg and after modification we got the optimized weight of the arm is kg. So, we achieved kg reduction in the weight of the arm. Based on the known boundary conditions which are calculated as provided with reference [4], the optimized model of arm is analyzed to check that the optimized model is within safe limit or not. Fig. 7 shows the static force analysis of the bucket for maximum breakout force (a) (b) Fig.6. Modified arm for optimization 45

54 Bhaveshkumar P. P. Patel, Jagdish M. Prajapati: Structural Optimization of of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Table 2. Optimization data of the arm Part no. Part name Quantity Modifications (mm) Thickness before modification Thickness after modification Total weight (Kg) Weight before optimization Weight of optimized model 1 Arm side cover Bucket cylinder mounting lug Bucket cylinder mounting lug bush Arm cylinder mounting lug Arm cylinder mounting lug bush 6 Arm collar Arm collar Cylinder-10, Collar Stiffners-5 Cylinder-10, Collar Stiffeners -5 Cylinder-5, Collar Stiffeners -3 Cylinder-5, Collar Stiffeners Arm reinforcement Arm stiffener Fig.7. Static force analysis for arm Fig.9. Maximum stresses of the optimized arm Fig.8. Boundary conditions for arm Fig.10. Maximum displacements of optimized arm 46 46

55 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Table 3. Optimization data of the boom Part no. Part name Quantity Dimensions before modification Modifications Dimensions after modification Total weight (Kg) Weight before optimization Weight of optimized model Modification in thickness (mm) 1 Boom side cover Arm cylinder mounting lug Arm cylinder mounting lug bush Boom cylinder mounting lug Boom cylinder mounting lug bush Boom to arm joint bush Boom reinforcement Arm cylinder mounting plate Boom cylinder mounting plate Boom top cover Boom bottom cover Modification in width thickness (mm) 12 Boom stiffeners Modification in length (mm) 13 Boom collar Now, yield strength of the material of mounting lug made up from HARDOX400 is 1000 MPa, by taking safety factor as 2, equation (1) yields [σ ] = 500 MPa and σ = MPa (Fig. 9), so σ [σ ] and this indicates that the design of the optimized arm is safe for strength. Fig. 10 shows the maximum displacement on arm is mm at bucket-arm joint end which is very small compare to minimum thickness of the plate used in the arm; therefore it is safe for deflection. 5. OPTIMIZATION OF BOOM The Fig. 11 shows the boom with different parts which are modified to get optimum dimensions based on available standard thickness of plates. Table 3 shows the name of the parts of the boom which are modified to get the optimized model. It also shows the dimensions and total weight of the parts before modification and after modifications. The total weight of the boom is kg and after modification we got the optimized weight of the boom is kg. So, we achieved kg reduction in the weight of the boom. Fig. 12 shows the static force analysis of the boom for maximum breakout force condition. Fig. 13 shows the boundary conditions applied to boom for the purpose of analysis. (a) (b) Fig. 11. Modified boom for optimization 47

56 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Misses stresses is acting on the mounting lug and it is MPa. Mounting lug is made from HARDOX400 and its yield strength is 1000 MPa, by taking safety factor as 2, equation (1) yields [σ ] = 500 MPa and σ = MPa, so σ [σ ] and this indicates that the design of the optimized boom is safe for strength. Fig. 15 shows the maximum displacement in the boom reported is mm at the boom cylinder mounting lug which is very less compare to minimum thickness of the plate used in the boom; therefore it is safe for deflection. Fig.12. Static force analysis for arm 6. OPTIMIZATION OF SWING LINK Fig.13. Boundary conditions for boom Fig.16. Modification of swing link for optimization The Fig. 16 shown the thicknesses of swing link which are modified to get optimum dimensions. Table 4 shows thickness before modification and thickness after modification. The total weight of the swing link is kg and after modification we got the optimized weight of the swing link is kg. So, we achieved 50.4 kg reduction in the weight of the swing link. Fig. 17 shows the static force analysis of the swing link for maximum breakout force condition. Fig. 18 shows the boundary conditions applied to swing link for the purpose of analysis. Fig.14. Maximum stresses of the optimized boom Fig.15. Maximum displacements of optimized boom Fig. 14 shows the results of the Von Misses stresses on optimized boom assembly in which the maximum Von Table 4. Optimization data of the swing link Sr. no. Thickness Thickness before optimization (mm) Thickness after optimization (mm) 1 t t t t t t t t Fig. 19 shows the results of the maximum Von Misses stresses acting on the cylinder mounting lug of optimized swing link of MPa. Cylinder mounting lug made from HARDOX400 and its yield strength is of 1000 MPa. The safety factor is taken as 2. 48

57 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Equation (1) yields [σ ] = 500 MPa, and σ = MPa, so σ [σ ] and this indicates that the design of the swing link is safe for strength. Fig. 20 shows the maximum displacement in the swing link reported is mm at the boom to swing link joint which is very less compare to minimum thickness of the plate used in the swing link; therefore it is safe for deflection. 7. OPTIMIZATION OF BACLHOE ASSEMBLY Fig.17. Static force analysis of swing link Fig.18. Boundary conditions for swing link Fig.21. Boundary conditions for backhoe assembly Fig.19. Maximum Von Misses stresses of optimized swing link Fig.20. Maximum displacements of optimized swing link Fig.22. Maximum Von Misses stresses in optimized backhoe assembly Fig. 21 shows the boundary conditions applied to the backhoe assembly for the purpose to carry out FE analysis. Fig. 22 shows the maximum Von Misses stresses produced at the mounting lugs in the backhoe attachment assembly of MPa. Mounting lugs are made from Hardox400 with the yield strength of 1000 MPa, by taking safety factor as 2, equation (1) yields σ = MPa, [σ ] = 500 MPa, this clearly indicates σ [σ ], so the stresses produced in the assembly of the backhoe are within the safe limits and the design is safe for strength. 49

58 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp STRESS ANALYSIS OF OPTIMIZED BACKHOE PARTS WITH CONSIDERATION OF WELDING Fig.23. Maximum Von Misses stresses of bucket with welding Fig.27. Maximum Von Misses stresses of backhoe assembly with welding As seen in Fig. 23 to Fig. 27 maximum Von Mises stresses developed in the mounting lugs and all mounting lugs are made from HARDOX400 having yielding strength of 1000 MPa. The developed stresses are very less compare to the safe stress [σ ] = 500 MPa, with the factor of safety is 2. Therefore the design of all the backhoe parts and assembly is safe for strength. 50 Fig.24. Maximum Von Misses stresses of arm with welding Fig.25. Maximum Von Misses stresses of boom with welding Fig.26. Maximum Von Misses stresses of swing link with welding 9. SHAPE OPTIMIZATION In this section shape optimization of backhoe excavator parts is carried out with the help of shape optimization tool of ANSYS. In the earlier section optimization is carried out by changing variable parameter that is thickness of plates. In this section results of shape optimization shows the area which can be remove from the part by changing the geometry of the part, it also shows that how much weight can be reduced from the particular part, so that the results obtained from shape optimization will be compared with the obtained results by changing parameter (thickness) based on trial and error method performed in previous section. The material of the parts, loading conditions and constraints (i.e. boundary conditions), and meshing of all the parts remain same as covered in previous all sections Shape optimization of bucket Here, results of ANSYS shape optimization tool is shown in the Fig. 28, it shows the area from which we have to remove material to reduce the weight of the bucket but it is not possible to change the geometry of the bucket. Because if we change the geometry then it will lose its basic functionality and will reduced in the capacity of the bucket. So instead of changing the geometry of bucket, we have changed the parameter (i.e. thickness) to reduced the weight same as taken in the earlier section. Here, the optimized weight obtained from the shape optimization is kg and optimized weight achieved by trial and error method (i.e. by changing thickness) is kg, so both results are very close to each other.

59 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Fig. 29 shows the results of shape optimization of arm. From Fig. 29 we can see that the material can be removed from the arm coloured in red. Here the optimized weight of the arm achieved by shape optimization is kg and optimized weight achieved by trial and error method (i.e. by changing thickness) is kg, so both results are very close to each other Shape optimization of boom (a) (a) (b) Fig. 28. Results of shape optimization for bucket 9.2. Shape optimization of arm (b) Fig. 30. Results of shape optimization for boom (a) Fig. 30 shows the result of shape optimization of boom. From Fig. 30 we can see that the material can be remove from the boom coloured in red. Here, the optimized weight of the boom obtained from shape optimization is kg and optimized weight achieved by trial and error method (i.e. by changing thickness) is kg, so both results are very close to each other Shape optimization of swing link (b) Fig.29. Results of shape optimization for arm Here, results of shape optimization are shown in the Fig. 31 for swing link, it shows the area from which we can remove material to reduce the weight of the swing link in red colour. Here, the optimized weight obtained from shape optimization is kg and optimized weight achieved by trial and error method (i.e. by changing thickness) is kg. The result indicates smaller differences in weight of swing link, obtained by both the 51

60 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp methods. Here, we have gone through the weight of kg of swing link because, the swing link carry the entire weight of the all other parts of the backhoe excavator attachment. The bucket having the complex shape and size therefore not considered for the application of classical theory. Here, arm is taken for the verification of the results of optimized model which are obtained from the FE analysis. For arm, a section plane A-A taken at 452 mm from pivot A 3, which is arbitrarily selected and shown in the Fig. 32. The calculations are made based on classical theory. It is the case of bending and twisting together. Since, the corner tooth is in action it will cause twisting. The cross section of the arm taken for study at section A- A is shown in the Fig. 33. Force analysis at section A-A shown in Fig. 34 for arm. (a) Fig.32. Section plane A-A in the front view section at of the arm (b) Fig.31. Results of shape optimization for swing link 10. VERIFICATION OF STRESS ANALYSIS USING CLASSICAL THEORY In this section the stresses produced in the optimized model of backhoe excavator performed using ANSYS software is verified with the stresses produced at the same section in the part of excavator by classical method. The classical theory applied to heavy duty backhoe excavator to verify the developed stresses by Reena Trivedi [14]. Von Mises theory is applicable for ductile material whereas the maximum principle stress theory is normally applicable for brittle material, but in the present case for validation of stress results of Von Mises, the maximum principle stress theory is applied because the shear stresses developed in the backhoe parts are very less compare to its design shear stresses and it is the case of bending and twisting stresses. Let, σ = Bending stress, N/mm 2 σ = Axial stress, N/mm 2 σ = Combined stress, N/mm 2 τ = Shear stress, N/mm 2 J = Polar moment of inertia of arm, mm 3 Fig.33. Details of arm section plane A-A Fig.34. Forces acting at the arm section Angle of force at A 3 with horizontal axis is θ = 8.27 Angle of force at A 12 with horizontal axis is θ =

61 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Force acting at point A 3 is R = Newton Force acting at point A 12 is R = Newton Distance from neutral axis to the outer fiber, y = 92.5 mm Cross section area of arm, CS = 2352 mm 2 Perpendicular distance between A 3 pivot and centre of section, L = 65 mm Perpendicular distance between A 12 pivot and centre of section, L = 319 mm Taking moment about point (centre of cross section of arm), we get Bending moment, BM R L R L BM = N.mm Let, I = moment of inertia about X axis for hollow rectangular cross-section of arm = mm 4 Bending stresses can be calculated using the following formula, σ (2) σ = MPa Force acting in X direction, F = N Force acting in Y direction, F = N Axial stress, σ F CS (3) σ = MPa Shear stress, τ F CS (4) τ = MPa Combined stress, σ σ σ (5) σ = = MPa Twisting moment, TM half width of bucket FD (6) TM = N.mm Where, Width of bucket = 547 mm FD = Maximum digging force in Newton FD = 7626 Newton Polar moment of inertia, J = 2bdt mm 3 (7) J = J = mm 3 Shear stress due to twisting, τ TM J (8) τ N/mm 2 Mean stress, σ σ σ = MPa Maximum principal stress, σ σ σ τ (10) σ = 41.9 MPa As per classical theory the value of maximum principal stress is MPa. For verification of stress results obtained from classical theory applied for arm, a same section plane is taken at a same distance of 452 mm from the pivot A and stresses developed at that section plane A-A are between MPa MPa and its average value is of MPa, which indicates that the results are remains identical with the results of classical theory, as shown in Fig. 36. Fig. 35 shows the section (9) plane (A-A) taken for calculations and Fig. 36 shows the result of stresses produced at that section plane A-A. Now here, boom is taken for the verification of the result of optimized model which is obtained from the FE analysis. For boom, a section plane B-B taken at 440 mm from pivot A 2 which is arbitrarily selected as shown in the Fig. 37. The calculations are made based on classical theory. It is the case of bending and twisting together. Since, the corner tooth is in action it will cause twisting. The cross section of the boom at section plane B-B taken for study is shown in the Fig. 38. Force analysis at section plane B-B shown in Fig. 39 for boom. Fig.35. Section plane A-A in the arm Fig.36. Stresses at section plane A-A from FE analysis Fig.37. Section plane B-B in the front view of the boom Angle of force at A with horizontal axis is θ = Force acting at point A is R = Newton 53

62 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp Distance from neutral axis to the outer fiber, y = 85 mm Cross section area of boom, CS = 2344 mm 2 Perpendicular distance between A pivot and centre of section, L = mm Fig.38. Details of boom section at section plane B-B FD = Maximum digging force in Newton FD = 7626 Newton Polar moment of inertia, J = 2bdt mm 3 (17) = J = N.mm Shear stress due to twisting, τ TM J (18) = MPa Mean stress, σ σ σ (19) = MPa Maximum principal stress, σ σ σ τ (20) σ = MPa Fig.39. Forces acting on the right hand side of the boom section Taking moment about point S (centre of cross section) the bending moment, BM R L (11) BM = N.mm Let, I = moment of inertia about X axis for hollow rectangular cross-section of boom = mm 4 Bending stress, σ (12) σ = MPa Force acting in X direction, F = N Force acting in Y direction, F = N Axial stress, σ (13) σ = MPa Shear stress, τ (14) τ = 7.29 MPa Combined stress, σ σ σ (15) = σ = MPa Twisting moment, TM half width of bucket FD (16) = N.mm Where, Width of bucket = 547 mm Fig.40. Section plane B-B in the boom Fig.41. Stresses at section plane B-B from FE analysis Fig. 40 shows the section plane B-B taken for calculations and Fig. 41 shows the results of stresses developed at section B-B getting from FE analysis performed using ANSYS. So as per classical theory the value of maximum principal stress is MPa. For verification of stress results obtained from classical theory applied for boom, a same section plane is taken at a same distance of 440 mm from the pivot point A and stresses developed at that section plane B-B are between MPa MPa and its average value is of MPa, Which indicates that the results are remains identical with the results of classical theory, as clearly shown in the Fig

63 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp RESULTS AND DISCUSSION Table 5. Summary of weight reduction by trial and error method (i.e. change in thickness) in different parts of excavator Sr. no. Name of the part Weight before optimization (Kg) Weight after optimization (Kg) Reduction in weight 1 Bucket Arm Boom Swing link Total weight Table 5 shows the weight of the all parts before optimization and weight of the all parts after optimization, table shows the total reduction in the weight by trial and error method (i.e. change in thickness) is kg. The FEA of the backhoe parts with the maximum breakout configuration is carried out for optimized model based on boundary conditions as calculated in chapter 7 are presented in this chapter. The maximum Von Mises stresses acting on bucket, arm, boom and swing link are MPa, MPa, MPa and MPa and the yield strength of these parts are 1000 Mpa, 450 Mpa, 1000 Mpa and 450 Mpa respectively, and by taking safety factor = 2 all the parts are found to be safe. The stress analysis of whole assembly is also carriedout and the stress produced are within the safe limit. Table 6 shows the comparison of sresses produced in the model without and with considering welding. Comparison shows that the backhoe model with welding having reduced stresses, so it is clear that the welding improves the strength of the parts. Table 6. Comparision of stresses produced in the optimized model with and without welding Sr. no. Name of parts Maximum Von Mises stress produced (MPa) Optimized model without welding Optimized model with welding 1 Bucket Arm Boom Swing link Backhoe assembly Table 7 shows the comparison of optimized weight achieved by trial and error method (i.e. reduction in thickness) and shape optimization achieved by ANSYS tool, which shows that the results are very close to each other and results from trial and error method are acceptable. The stresses produced in the optimized model by performing FE analysis using ANSYS software is also verified with the stresses obtained by applying classical theory and the results obtained from both the methods are identical. Table 7. Comparision of optimized weight obtained by trial and error method and shape optimization Sr. no. Name of parts Weight before optimization (Kg) Weight after optimization Trial and error method (i.e. by changing thickness) (Kg) % Shape optimiza -tion (i.e. by changing geometr y) Variatio n in results by both methods 1 Bucket Arm Boom Swing 4 Link Total Weight CONCLUSIONS FE analysis of backhoe parts shows that the parts with welding provide higher strength. Structural weight optimization carried out by trial and error method shows the total reduction in weight is of kg (24.96%) and weight reduced by applying shape optimization is of kg (27.91%). Comparison shows that the variations in results of individual parts are very less and total variation in result is of only 3.93% which reflect that the results of structural weight optimization performed by trial and error method are accurate and acceptable. The differences in results of the Von Mises stresses and the classical theory are very less and we can say that the results are identical and acceptable. REFERENCES [1] BHAVESHKUMAR P. PATEL, DR. J. M. PRAJAPATI, Soil-Tool Interaction as a Review for Digging Operation of Mini Hydraulic Excavator, International Journal of Engineering Science and Technology, Vol. 3 No. 2, February 2011, pp [2] BHAVESHKUMAR P. PATEL AND J. M. PRAJAPATI, A Review on FEA and Optimization of Backhoe Attachment in Hydraulic Excavator, IACSIT International Journal of Engineering and Technology, Vol. 3, No. 5, October 2011, pp [3] BHAVESHKUMAR P. PATEL, DR. J. M. PRAJAPATI AND BHARGAV J. GADHVI, An Excavation Force Calculations and Applications: An Analytical Approach, International Journal of Engineering Science and Technology, Vol. 3, No. 5, May 2011, pp [4] BHAVESHKUMAR P. PATEL AND J. M. PRAJAPATI, Evaluation of Bucket Capacity, Digging Force Calculations and Static Force Analysis of Mini Hydraulic Backhoe Excavator, MACHINE DESIGN The Journal of Faculty of Technical Sciences, Vol.4, No.1, 2012, pp [5] BHAVESHKUMAR P. PATEL AND J. M. 55

64 Bhaveshkumar P. Patel, Jagdish M. Prajapati: Structural Optimization of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach; Machine Design, Vol. 5(2013) No. 1, ISSN ; pp PRAJAPATI, Static Analysis of Mini Hydraulic Backhoe Excavator Attachment Using FEA Approach, International Journal of Mechanical Engineering and Robotics Research, Submitted, [6] C. S. KRISHNAMURTHY, Finite element analysis theory and programming, Tata McGraw-Hill Publishing Company Limited, 2007, pp [7] MOHSEN KARGAHI, JAMES C. ANDERSON AND MAGED M. DESSOUKY, Structural Optimization with Tabu Search, Senior Engineer, Weidlinger Associates, Inc., 2525 Michigan Avenue, D2-3, Santa Monica, California, 90404, [8] MEHMET YENER, Design of a Computer Interface for Automatic Finite Element Analysis of an Excavator Boom, M.S. Thesis, The Graduate School of Natural and Applied Sciences of Middle East Technical University, May 2005, pp 1-4, [9] NARESHKUMAR N. OZA, Finite Element Analysis and Optimization of an Earthmoving Equipment Attachment - Backhoe, M. Tech. Thesis, Nirma University, Institute of science and Technology, Ahmedabad , May [10] MEHTA GAURAV K., Design and Development of an Excavator Attachment, M. tech. Thesis, Nirma University, Institute of science and Technology, Ahmedabad , May 2008, pp 1. [11] SSAB, HARDOX400 data sheet, [12] STEEL AUTHORITY OF INDIA LIMITED (SAIL), SAIL product panorama product brochure, ecifications%20for%20plates.htm, [13] TIRUPATHI R. CHANDRUPATLA AND ASHOK G. BELEGUNDU, Introduction to Finite Elements in Engineering, Pearson Education Publication, Delhi, India, [14] TRIVEDI REENA, Calculation of Static Forces And Finite Element Analysis of Attachments of An Excavator, M. Tech. Thesis, Nirma University, Institute of Science and Technology, Ahmedabad , May

65 machine design, Vol.5(2013) No.1, ISSN pp Research paper 3-DIMENSIONAL EXPERIMENTAL AND FINITE ELEMENT STRESS ANALYSIS OF C.I. WEDGE OF SLUICE VALVE Narayan DHARASHIVAKAR 1, * - Prashant PATIL 1 - Krishnakumar JOSHI 1 1 Department of Mechanical Engineering, TKIET Warananagar, Shivaji University, Kolhapur Abstract: The sluice valves of various capacities are generally used for controlling flow rate of fluids. During operation of sluice valve, the wedge is moved to & fro manually by means of screw. The construction is such that the wedge is reciprocated by the rotation of the screw in a T-Nut which is fixed inside the gap of wedge. The entire experimental work has been carried out on a wedge to control the flow of water, in M/s. Sarvodaya Sugar Industries, Ashta, Maharashtra. While operating, it has been observed that the CI wedge frequently breaks due to load coming on the wedge. It seems that due to eccentric loading, there is a peculiar tearing of the part of the C.I. wedge. Thus this is a problem of 3 dimensional stress analysis. We can analyze the stress pattern by using 3D photo elasticity and verify the results with FE method. Key words: Sluice Valve, 3d Photoelastic Analysis, FEA 1. INTRODUCTION There is an ever increasing use of flow control valves in industries like Agricultural, Sugar and Chemical. These valves are working under abnormal conditions of temperature, Pressure and other conditions. This has given special importance to analytical and experimental methods for determining their working stresses. In such valves, the knowledge of these stresses will help us to avoid their failure. The sluice valves of various capacities are generally used for controlling flow rate of fluids in various industries. During operation of sluice valve, the wedge is moved to & fro manually by means of screw. The construction is such that the wedge is reciprocated by the rotation of the screw in a T-Nut which is fixed inside the gap of wedge as shown in Figure 1. the wedge, made up of cast iron, frequently breaks due to load coming on the post system of the wedge above the T- Nut, the shape of which is intricate. It seems that due to eccentric loading, there is a peculiar tearing of the part of the C.I. wedge. The crack is initiated just near to high stress-concentration region. Thus this is a problem of 3 dimensional stress analysis. We can analyze the stress pattern by using 3 dimensional stress analysis using 3D photo elasticity and verify the results with FE method. 2. EXPERIMENTAL ANALYSIS Here we have carried out Experimental analysis by using 3 dimensional Photoelastic Technique, which involves the following steps: 2.1. Model Making As the accuracy of the Photoelastic model has got the major effects on the results obtained, the preparation of the model bears its own importance in the whole problem of Photoelastic Stress Analysis. The Photoelastic model for investigation should satisfy the following requirements: The model should be free from all residual stresses. The model exhibits all properties of good Photoelastic material. The material used for the model should be sufficiently transparent. The model should not show any Time-Age Effect. Fig.1. Sluice Gate Valve The entire experimental work has been carried out on a wedge of a sluice valve which is used to control the flow of water at M/s. Sarvodaya Sugar Industries, Ashta, Maharashtra. While operating, it has been observed that Considering all requirements of a good Photoelastic model, one would come to know that the preparation of a model is a sort of technique which needs not only great care but a lot of experience and practice. During casting, first we prepared the wooden pattern of the wedge which was then used to manufacture the mould of the synthetic rubber. Owing to the complexity of the *Correspondence Author s Address: Department of Mechanical Engineering, TKIET, Shivaji University, Kolhapur. nsdharashivkar@tkietwarana.org

66 Narayan Dharashivakar, Prashant Patil, Krishnakumar Joshi: 3-Dimensional Experimental and Finite Element Stress Analysis of C.I. Wedge of Sluice Valve; Machine Design, Vol.5(2013) No.1, ISSN ; pp geometry, it was decided to split the pattern into two. Finally, we prepared the photoelastic model of the wedge using this mould and photoelastic material (Araldite AY- 103 and Hardener HY-951.) as shown in Figure 2. disc was also placed in a separate loading frame and kept inside the stress freezing oven. A stress freezing oven (Han-Yong make) having PX09 Process Controller was used for stress freezing purpose. The total Stress Freezing cycle lasts for 34 hours. This cycle involves three distinct steps, Heating, Soaking and Cooling, as shown in figure. Figure 5. Fig.2. Mould of a Wedge 2.2. Design and Development of Loading Fixture The basic requirement of the loading frame is to simulate the actual working conditions of the Wedge. Working conditions involve, tangential force caused due to fluid pressure, frictional force between valve seat & wedge seat and Weight of the wedge, as shown in Figure 3 While designing the loading frame, following aspects have been considered: a. Construction of the frame should be as simple as possible. b. It should be capable to apply the required load which will simulate the loading conditions of the prototype. c. It should provide proper constraint to the model. d. It should provide easy and quick replacement of the model. It consists of a base plate on which a taper plate is mounted. This plate simulates the actual support for the surface of the wedge. The dead weight and lever arrangement is used to apply forces in vertical direction. A clamping pressure plate is used to apply pressure. Figure Slicing Fig.5. Stress Freezing Cycle Generally, in 3D Photoelasticity, the analysis is performed on slices cut from the model after stress freezing. The model is sliced to remove the planes of interest which can then be examined individually to determine the state of stress existing in that particular plane or slice. The slice preparation involves three steps viz. Layout, Cutting of the slices and flattening of the slice surfaces. The particular slicing plan employed in sectioning a 3D Photoelasticity model will of course depend upon the geometry of the model and information being sought in the analysis. The figure 6 shows the slicing plan employed for the wedge model to remove and examine the slices. The slices were cut by using high speed horizontal milling machine and slicing saw of 1mm thickness. The sufficient amount of cutting oil is spread at the time of cutting. The slicing thickness was kept 3mm. After cutting the slices, the surface of each slice was finished manually with the help of Zero number polish paper. Fig.4. Loading Frame 2.3. Stress Freezing The stress freezing technique is used to lock the stresses permanently in the model. After mounting the model on a loading frame, the model along with the loading frame was placed in the stress freezing oven. The calibration 58 Fig.6. Slicing Plan 2.5. Experimental Observations The finished slices were observed under polariscope to find out fringe order at a point of interest to evaluate the

67 Narayan Dharashivakar, Prashant Patil, Krishnakumar Joshi: 3-Dimensional Experimental and Finite Element Stress Analysis of C.I. Wedge of Sluice Valve; Machine Design, Vol.5(2013) No.1, ISSN ; pp stress pattern at that point, as shown in Figure 7. Also we found out Fringe Order at the centre of circular calibration disc to evaluate Material Fringe Value. Slice 1 Slice 2 3. FINITE ELEMENT ANALYSIS: ANSYS software is used to verify the results obtained by Experimental method. The entire wedge model was descretized using Solid 92 element. Figure 9 shows the meshing and boundary conditions applied to the wedge model and Figure 10 shows the variation of the stresses along the width of post system of the wedge. Table II shows the comparative study of the results obtained by both the methods. Slice 3 Slice 4 Fig.7. A Photoelastic fringe pattern showing the isochromatics in longitudinal slices. The actual stresses produced in the wedge were obtained experimentally are as shown in the Table I. Fig.9. Meshing & boundary conditions applied to wedge model Table 1. Stress Values Obtained By Experimental Method Stresses Produced (N/mm2) for Slice No. 343 N Tensile Force A B C D It has been found that most of the failures take place at the sections where the stress concentration is present that is at places where abrupt changes in the form of part occur. Even though the average stresses for cross section is kept far below the elastic limit of the material, failures are found to occur without warning.. / k. / Actual stress concentration factor (k a ) has been experimentally found to be By referring, Leven-Hartman Chart, the theoretical stress concentration factor (k t ) is found to be Figure 8 Fig.8. Leven-Hartman Chart Fig.10. Variations of the Stresses 4. DISCUSSION AND CONCLUSION: In sluice and gate valves, wedge used to control the flow of fluid is a very important component. The gland is loaded in a similar manner to that of tensile loading by means of a T-Nut. The reaction of this tensile loading is on post system of the wedge symmetrically placed tending it to fail due to bending and tension. The stresses produced at any discontinuity are different in magnitude from those calculated by elementary formulae. It is observed that a small discontinuity in a part will have effect of increasing the magnitude of stresses. The measurement of stress concentration in machine components is of considerable importance because mechanical failures of such parts are frequently due to fractures which are initiated through the sections having stress concentrations. Experimental stress analysis of wedge was carried out considering the importance of the wedge in controlling the flow of fluid, its geometry and loading. It has been observed that 3 dimensional photoelastic stress analysis technique using stress freezing phenomenon is the best method for experimental stress analysis of gland. 59

68 Narayan Dharashivakar, Prashant Patil, Krishnakumar Joshi: 3-Dimensional Experimental and Finite Element Stress Analysis of C.I. Wedge of Sluice Valve; Machine Design, Vol.5(2013) No.1, ISSN ; pp Low temperature curing epoxy mixture is a good material for casting such type of component. Good castings can be obtained by using silicon rubber molds. Only longitudinal slices are helpful in exhibiting the stress pattern in the wedge. The stress concentration factor at stress concentration region is observed to be more that theoretically required. Referring to Figure 8 and 11, it is observed that stress concentration factor goes on reducing with r/d ratio. If we neglect tensile stresses created by bending at inner edge, there is fairly good agreement between theoretical, experimental and FE analysis as in Table 2. There seems to be condition of impact between the surface of T-Nut and that of post system of wedge as there is a gap of 3-4mm between these two. The failure may be because of the impact of T-Nut on surface of wedge. Referring to Figure 12, it is observed that the stress concentration factor goes on varying along the width of the wedge Stress Concentration factor A A A B C C D B C Distance (mm) Slice1 Slice2 Slice3 Slice4 Fig.12. Variation of Stress concentration factor along the width of the wedge D D Fig.11. Variation of Stress concentration factor with r/d ratio Table 2. Stresses Produced (N/mm 2 ) for 343 N Tensile Force Slice A B No. 60 *E *A *V *E *A *V Slice No. C *E *A *V *E *A *V *E: Experimental Value *A: ANSYS Value *V: Variation in Percentage D 5. REFERENCES: [1] Emerging Trends in Design Engineering (Vol.II) [2] Ball & Roller Bearing Engineering, Published by Fag Kugel Fischer and George Schafer & Co. [3] Instruction Manual- Silicon Mould making materials Published by Metroark Pvt. Ltd. [4] 3D Photoelastic Photo analysis of Femur-C. Y. Lau and H. Teoh, 1998 [5] 3D Photoelastic study of stresses in Rack gearsexperimental Mechanics, May 1979 by N. A. Rubayi and H.W. Tam [6] Designing by Photoelasticity, Chapman & Hall Ltd. by R.B. Heywood [7] 3D Photoelasticity by stress freezing-experimental Mechanics-Dec by Jan Cernosek [8] Machine Design- Paul H. Black, O. Eugene Adams Jr. Published by McGraw-Hill Publishing, International Edition, pp [9] Experimental Stress Analysis by J. W. Dally and W. F. Riley- McGraw-Hill Book Co., pp [10] Experimental Stress Analysis by Sadhu-Singh, pp [11] PSG Design data Book.PSG College of Engg. Coimbatore [12] Introduction to Finite Element in Engineering. Ashok D. Belegundu