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BULETINUL 

INSTITUTULUI 

POLITEHNIC 

DIN IAŞI 

Tomul LVIII (LXII) 

Fasc. 1 

CONSTRUCłII DE MAŞINI 

2012 Editura POLITEHNIUM

BULETINUL INSTITUTULUI POLITEHNIC DIN IAŞI 

PUBLISHED BY 

“GHEORGHE ASACHI” TECHNICAL UNIVERSITY OF IAŞI 

Editorial Office: Bd. D. Mangeron 63, 700050, Iaşi, ROMANIA 

Tel. 40-232-278683; Fax: 40-232-237666; e-mail: polytech@mail.tuiasi.ro 

Editorial Board 

President: Prof. dr. eng. Ion Giurma, Member of the Academy of Agricultural 

Sciences and Forest, Rector of the “Gheorghe Asachi” Technical University of Iaşi 

Editor-in-Chief: Prof. dr. eng. Carmen Teodosiu, Vice-Rector of the 

“Gheorghe Asachi” Technical University of Iaşi 

Honorary Editors of the Bulletin: Prof. dr. eng. Alfred Braier, 

Prof. dr. eng. Hugo Rosman 

Prof. dr. eng. Mihail Voicu, Corresponding Member of the Romanian Academy, 

President of the “Gheorghe Asachi” Technical University of Iaşi 

Editors in Chief of the MACHINE CONSTRUCTIONS Section 

Prof. dr. eng. Radu Ibănescu, Assoc. prof. dr. eng. Aristotel Popescu 

Honorary Editors: Prof. dr. eng. Gheorghe NagîŃ, Prof. dr. eng. Cezar Oprişan 

Associated Editor: Assoc. prof. dr. eng. Eugen Axinte 

Editorial Advisory Board 

Prof.dr.eng. Nicuşor Amariei, „Gheorghe Asachi” Technical Prof.dr.eng. Dorel Leon, „Gheorghe Asachi” Technical 

University of Iaşi 


Assoc.prof.dr.eng. Aristomenis Antoniadis, Technical Prof.dr.eng. James A. Liburdy, Oregon State University, 

University of Crete, Greece 

Corvallis, Oregon, SUA 

Prof.dr.eng. Virgil Atanasiu, „Gheorghe Asachi” Technical Prof.dr.eng.dr. h.c. Peter Lorenz, Hochschule für Technik 


und Wirtschaft, Saarbrücken, Germany 

Prof.dr.eng. Petru Berce, Technical University of 

Prof.dr.eng. Nouraş -Barbu Lupulescu, University 

Cluj-Napoca 

Transilvania of Braşov 

Prof.dr.eng. Ion Bostan, Technical University of Chişinău, Prof.dr.eng. Fabio Miani, University of Udine, Italy 

Republic of Moldova 

Prof.dr.eng. Mircea Mihailide, „Gheorghe Asachi” Technical 

Prof.dr.eng. Walter Calles, Hochschule für Technik und University of Iaşi 

Wirtschaft des Saarlandes, Saarbrücken, Germany Prof.dr.eng. Sevasti Mitsi, Aristotle University of 

Prof.dr.eng. Doru Călăraşu, „Gheorghe Asachi” Technical Thessaloniki, Salonic, Greece 


Prof.dr.eng. Vasile Neculăiasa, „Gheorghe Asachi” Technical 

Prof.dr.eng. Francisco Chinesta, École Centrale de Nantes, University of Iaşi 

France 

Prof.dr.eng. Fernando José Neto da Silva, University of 

Assoc.prof.dr.eng. Conçalves Coelho, University Nova of Aveiro, Portugal 

Lisbon, Portugal 

Prof.dr.eng. Dumitru Olaru, „Gheorghe Asachi” Technical 

Prof.dr.eng. Juan Pablo Contreras Samper, University of University of Iaşi 

Cadiz, Spain 

Prof.dr.eng. Manuel San Juan Blanco, University of 

Assoc.prof.dr.eng. Mircea Cozmîncă, „Gheorghe Asachi” Valladolid, Spain 

Technical University of Iaşi 

Prof.dr.eng. Loredana Santo,University „Tor Vergata”, 

Prof.dr.eng. Spiridon CreŃu, „Gheorghe Asachi” Technical Rome, Italy 


Prof.dr.eng. Cristina Siligardi, University of Modena, Italy 

Prof.dr.eng. Gheorghe Dumitraşcu, „Gheorghe Asachi” Prof.dr.eng. Filipe Silva, University of Minho, Portugal 

Technical University of Iaşi 

Prof.dr.eng. LaurenŃiu Slătineanu, „Gheorghe Asachi” 

Prof.dr.eng. Cătălin Fetecău, University „Dunărea de Jos” of Technical University of Iaşi 

GalaŃi 

Lecturer dr.eng. Birgit Kjærside Storm, Aalborg 

Prof.dr.eng. Mihai GafiŃanu, „Gheorghe Asachi” Technical Universitet Esbjerg, Denmark 


Prof.dr.eng. Ezio Spessa, Politecnico di Torino, Italy 

Prof.dr.eng. Radu Gaiginschi, „Gheorghe Asachi” Technical Prof.dr.eng.Roberto Teti, University „Federico II”, Naples, Italy 


Prof.dr.eng. Alexei Toca, Technical University of Chişinău, 

Prof.dr.eng. Francisco Javier Santos Martin, University of Republic of Moldova 

Valladolid, Spain 

Prof.dr.eng. Hans-Bernhard Woyand, Bergische University 

Prof. dr. Dirk Lefeber, Vrije Universiteit Brussels, Belgium Wuppertal, Germany

Papers presented at the 

6 th INTERNATIONAL CONFERENCE on 

MANUFACTURING SYSTEMS 

Iaşi, October 20 th – 21 st , 2011 

organized by the DEPARTMENT OF MACHINE TOOLS, 

Faculty of MACHINE MANUFACTURING&INDUSTRIAL MANAGEMENT 

Papers published with the support of 

NATIONAL AUTHORITY for SCIENTIFIC RESEARCHERS 

EDITORIAL BOARD 

MACHINE CONSTRUCTIONS 

Fascicle 1 

Conf.univ.dr.ing. Irina Cozmîncă 

Conf.univ.dr.ing. Cătălin Ungureanu 

Sef lucrari.dr.ing. Bruno Rădulescu 

Drd..ing. Ana Maria Matei

B U L E T I N U L I N S T I T U T U L U I P O L I T E H N I C D I N I A Ş I 

B U L L E T I N O F T H E P O L Y T E C H N I C I N S T I T U T E O F I A Ş I 

Tomul LVIII (LXII), Fasc. 1 2012 


S U M A R 

MIHAI AFRĂSINEI, DORU CĂLĂRAŞU, VASILE DAMASCHIN, IRINA 

MARDARE şi ADRIAN OLARU, Analiza prin simulare numerică a 

unui sistem hidraulic autoadaptiv destinat turbinelor eoliene (engl., rez. 

rom.)........................................................................................................... 

ADRIAN SORIN AXINTI şi GAVRIL AXINTI, Modele dinamice pentru 

sisteme de tracŃiune (engl., rez. rom.)......................................................... 

GAVRIL AXINTI şi ADRIAN SORIN AXINTI, Modele pentru disipatoare 

hidraulice de energie seismică (engl., rez. rom.)......................................... 

CARMEN BAL, CARMEN IOANA IUHOS şi NICOLAIE BAL, Cercetări 

experimentale privind efectele căldurii într-o instalaŃie de acŃionare cu 

debite armonice. (I) InstalaŃie de acŃionare cu debite armonice − Montaj 

în paralel (engl., rez. rom.).......................................................................... 23 

VLAD BOCĂNEł, HORIA ABĂITANCEI, CONSTANTIN CHIRIłĂ, 

MARIUS DENEŞ-POP şi LIVIU BĂRNUłIU, Analiza dinamică a unui 

sistem de acŃionare hidraulic folosit în acŃionarea unui sistem care se 

deplasează cu viteză redusă, în condiŃii de sarcină mare (engl., rez. rom.) 

ILARE BORDEAŞU, MIRCEA POPOVICIU, ADRIAN KARABENCIOV, 

ALIN DAN JURCHELA şi CONSTANTIN CHIRITA, Noi contribuŃii 

în corelarea proprietăŃilor mecanice cu rezistenŃa la cavitaŃie a oŃelurilor 

inoxidabile (engl., rez. rom.)..................................................................... 35 

CONSTANTIN CHIRIłĂ, ANDREI GRAMA şi DUMITRU ZETU, 

Cercetări privind pierderile prin frecare în dispozitivele de tensionare la 

mersul în sarcină (engl., rez. rom.)........................................................... 43 

CORNELIU CRISTESCU, PETRIN DRUMEA, CĂTĂLIN DUMITRESCU 

şi DRAGOŞ ION GUŞĂ, Cercetări experimentale privind comportarea 

dinamică a servo-sistemelor hidraulice liniare (engl., rez. rom.) ............. 

FLORINA-CRISTINA FILIP, Metode eficiente de repartizare a costurilor 

directe, a costurilor cu forŃa de muncă şi a costurilor de regie 

(engl., rez. rom.)......................................................................................... 

Pag. 

1 

9 

17 

29 

51 

59

ANDREI GRAMA, CONSTANTIN CHIRIłĂ, DUMITRU ZETU şi MIHAI 

AXINTE, Analiza prin metoda elementului finit a ansamblului bacuri de 

tragere – bucşă port-bacuri (engl., rez. rom.)............................................. 

DANIELA IONESCU, ION BOGDAN şi GABRIELA APREOTESEI, Asupra 

proprietăŃilor controlabile ale straturilor subŃiri de YIG cu aplicaŃii la 

fabricarea dispozitivelor de microunde (engl., rez. rom.) .......................... 75 

ALIN LUCA, MIRCEA COZMÎNCĂ şi ANA MARIA MATEI, Stabilirea 

relaŃiei dintre Ra şi Rz la strunjirea oŃelului OL50 (engl., rez. 

rom.)............................................................................................................ 

IRINA MARDARE şi IRINA TIłA, Senzor de forŃă inclus într-un sistem 

WIM hidrostatic (engl., rez. rom.).............................................................. 91 

ANA-MARIA MATEI, MIRCEA COZMÎNCĂ şi ALIN LUCA, Cercetări 

privind uniformizarea forŃelor de aşchiere la frezarea frontală 

(engl., rez. rom.) ......................................................................................... 

MEHRAN DOULAT ABADI (Malaezia) şi SHA’RI MOHD YUSOF 

(Malaezia), Studiu preliminar asupra factorilor cheie care susŃin 

principiile managementului calităŃii totale (engl., rez. rom.) ..................... 

MARIUS MILEA şi MIRCEA COZMÎNCĂ, Sinteză asupra evaluării 

coeficientului de deformare plastică CD (engl., rez. rom.) ........................ 117 

EUGEN-VLAD NĂSTASE şi DORU CĂLĂRAŞU, InfluenŃa variaŃiei corzii 

asupra performanŃelor unei miniturbine cinetice (engl., rez. rom.).......... 121 

EUGEN-VLAD NĂSTASE şi DORU CĂLĂRAŞU, Cercetări teoretice 

privind influenŃa numărului de pale asupra eficienŃei unei miniturbine 

(engl., rez. rom.).......................................................................................... 

VASILE NĂSUI, Modelarea controlului mişcării la actuatorii liniari 

electromecanici (engl., rez. rom.)................................................................ 129 

IOANA PETRE, TUDOR DEACONESCU, ANDREA DEACONESCU şi 

DAN PETRE, Modelarea cu element finit a unui echipament de 

reabilitare (engl., rez. rom.)........................................................................ 

IOANA PETRE, TUDOR DEACONESCU, ANDREA DEACONESCU şi 

DAN PETRE, ConsideraŃii privind calculul volumului muşchiului 

pneumatic (engl., rez. rom.)........................................................................ 

TEODOR COSTINEL POPESCU şi IOAN LEPĂDATU, Tehnici moderne de 

experimentare a pompelor hidrostatice reglabile (engl., rez. rom.) ......... 149 

OCTAVIAN PRUTEANU, CONSTANTIN CĂRĂUŞU şi LUCIAN 

TĂBĂCARU, ConsideraŃii asupra deformării plastice la rece a inelelor 

de rulmenŃi (engl., rez. rom.) .................................................................... 

BRUNO RĂDULESCU şi MARA-CRISTINA RĂDULESCU, Estimarea 

costului de producŃie în cazul produselor (engl., rez. rom.) ....................... 169 

65 

85 

97 

105 

125 

137 

143 

157

LUCIANA-CRISTIANA STAN şi VLADIMIR MĂRĂSCU-KLEIN, Rolul 

competenŃelor profesionale în activitatea antreprenoriala (engl., rez. 

rom.) ........................................................................................................... 

IRINA TIłA şi IRINA MARDARE, Aspecte privind sisteme hidraulice cu 

reglare secundară (engl., rez. rom.) ............................................................ 

ALEXANDRA TOMA şi CORNEL CIUPAN, Factorii umani asociaŃi 

designului ergonomic al unei interfeŃe om-maşină (engl., rez. rom.)......... 

LILIANA TOPLICEANU, ADRIAN GHENADI şi MARIUS PASCU, Studiu 

despre rolul acumulatoarelor în funcŃionarea sistemelor hidraulice cu 

reglaj secundar (engl., rez. rom.)................................................................ 

CĂTĂLIN UNGUREANU, IRINA COZMÎNCĂ şi RADU IBĂNESCU, 

ConsideraŃii privind fişele de control utilizate în controlul statistic al 

proceselor. (II) Fişe de control pentru date rare şi cumulative 

(engl., rez. rom.) ......................................................................................... 

DĂNUł ZAHARIEA, Diagrame funcŃionale pentru modelarea procesului de 

golire a unui rezervor printr-un orificiu de golire (engl., rez. rom.)........... 

DĂNUł ZAHARIEA, Diagrame funcŃionale pentru modelarea procesului de 

golire a unui rezervor printr-o conductă de golire (engl., rez. rom.)........... 

DAN POPESCU, O pledoarie pentru modernizarea maşinilor unelte actuale din 

industria naŃională (engl., rez. rom.)........................................................... 

177 

185 

191 

197 

203 

211 

219 

227

B U L E T I N U L I N S T I T U T U L U I P O L I T E H N I C D I N I A Ş I 

B U L L E T I N O F T H E P O L Y T E C H N I C I N S T I T U T E O F I A Ş I 

Tomul LVIII (LXII), Fasc. 1 2012 

MACHINE CONSTRUCTION 

MIHAI AFRĂSINEI, DORU CĂLĂRAŞU, VASILE DAMASCHIN, IRINA 

MARDARE and ADRIAN OLARU, Analysis by Numerical Simulation 

of a New Self-Adaptive Hydraulic System Used at Wind Turbines 

(English, Romanian summary)................................................................... 

ADRIAN SORIN AXINTI and GAVRIL AXINTI, Dynamic Models for 

Traction Systems (English, Romanian summary)...................................... 

GAVRIL AXINTI and ADRIAN SORIN AXINTI, Models For The Hydraulic 

Seismic Energy Dissipaters (English, Romanian summary)....................... 

CARMEN BAL, CARMEN IOANA IUHOS and NICOLAIE BAL, Research 

on Experimental Heat Effects in a Flow with Harmonic Drive 

Installation. (I) Drive Installation of Harmony Flow − Assembly in 

Parallel, (English, Romanian summary)..................................................... 

VLAD BOCĂNEł, HORIA ABĂITANCEI, CONSTANTIN CHIRIłĂ, 

MARIUS DENEŞ-POP and LIVIU BĂRNUłIU, Dynamic Analysis of a 

Hydraulic Actuation System of Very Slow Moving Devices 

(English, Romanian summary) .................................................................. 

ILARE BORDEAŞU, MIRCEA POPOVICIU, ADRIAN KARABENCIOV, 

ALIN DAN JURCHELA and CONSTANTIN CHIRIłĂ, New 

Contributions in the Correlation of Mechanical Properties with the 

Cavitation Resistance of Stainless Steels (English, Romanian summary)........................................................................................................... 

CONSTANTIN CHIRIłĂ, ANDREI GRAMA and DUMITRU ZETU, 

Research on Frictional Losses in Tensioning Devices at Full Load 

(English, Romanian summary)................................................................... 

CORNELIU CRISTESCU, PETRIN DRUMEA, CĂTĂLIN DUMITRESCU 

and DRAGOŞ ION GUŞĂ, Experimental Research Regarding the 

Dynamic Behaviour of Linear Hydraulic Servo-systems (Engliah, 

Romanian summary)................................................................................... 

CORNELIU CRISTESCU PETRIN DRUMEA, CĂTĂLIN DUMITRESCU i DRAGO ION GU Ă, Experimental Research Regarding the Dynamic Behavior of Linear Hydraulic Servo-Systems (English, Romanian Summary) ................................................................................ 

C O N T E N T S 

.FLORINA-CRISTINA FILIP, Effective Methods of Cost Breakdown for 

Direct Costs, Tools and Product Costs (English, Romanian 

summary)..................................................................................................... 

Pp. 

1 

9 

17 

23 

29 

35 

43 

51 

59

ANDREI GRAMA, CONSTANTIN CHIRIłĂ, DUMITRU ZETU and 

MIHAI AXINTE, Analysis by Finite Element Method of Assembly 

Wedge Grips - Mantle Corbel (English, Romanian summary)................... 

DANIELA IONESCU, ION BOGDAN and GABRIELA APREOTESEI, 

About the Tunable Properties of the YIG Films with Applications in 

Microwave Devices Manufacturing (English, Romanian summary) ......... 

.ALIN LUCA, MIRCEA COZMÎNCĂ and ANA MARIA MATEI, 

Experimental Determination of Ra and Rz on Turning Steels 

(English, Romanian summary).................................................................... 

IRINA MARDARE and IRINA TIłA, Force Sensor in a WIM Hydrostatic 

System (English, Romanian summary)....................................................... 

ANA-MARIA MATEI, MIRCEA COZMÎNCĂ and ALIN LUCA, Researches 

Concerning the Uniformization of Cutting Forces in Face Milling 

(English, Romanian summary) .................................................................. 

MEHRAN DOULAT ABADI (Malaysia) and SHA’RI MOHD YUSOF 

(Malaysia), A Preliminary Study of the Key Factors for Sustaining Total 

Quality Practices (English, Romanian summary)....................................... 105 

MARIUS MILEA and MIRCEA COZMÎNCĂ, Synthesis on the Assessment 

of Chips Contraction Coefficient CD (English, Romanian summary) ........ 117 

EUGEN-VLAD NĂSTASE and DORU CĂLĂRAȘU, Influence of Chord 

Variation on the Performance of a Kinetic Miniturbine (English, 

Romanian summary)................................................................................... 121 

EUGEN-VLAD NĂSTASE and DORU CĂLĂRAȘU, Theoretical Research 

Regarding the Blades Number Influence of the Miniturbine Efficiency 

(English, Romanian summary)................................................................... 125 

VASILE NĂSUI, About Modelling The Movement Control of the 

Electromechanic Linear Actuator (English, Romanian summary)............. 129 

IOANA PETRE, TUDOR DEACONESCU, ANDREA DEACONESCU and 

DAN PETRE, Finite Element Modeling of a Knee and Hip 

Rehabilitation Equipment (English, Romanian summary).......................... 137 

IOANA PETRE, TUDOR DEACONESCU, ANDREA DEACONESCU and 

DAN PETRE, Some Considerations Regarding Pneumatic Muscle 

Volume (English, Romanian summary)...................................................... 143 

TEODOR COSTINEL POPESCU and IOAN LEPĂDATU, Modern 

Techniques for Experimentation of Adjustable Hydrostatic Pumps 

(English, Romanian summary).................................................................... 

OCTAVIAN PRUTEANU, CONSTANTIN CĂRĂUŞU and LUCIAN 

TĂBĂCARU, Some Considerations about Cold Plastic Deformation of 

Bearing Rings (English, Romanian summary) ........................................... 157 

65 

75 

85 

91 

97 

149

BRUNO RĂDULESCU and MARA-CRISTINA RĂDULESCU, The Costs 

Estimation for the Mechanical Production (English, Romanian 

summary).................................................................................................... 

LUCIANA-CRISTIANA STAN and VLADIMIR MĂRĂSCU-KLEIN, The 

Role of Professional Competence in Business Entrepreneurship (English, 

Romanian summary.) ................................................................................. 

IRINA TIłA and IRINA MARDARE, Some Aspects Regarding Hydraulic 

Systems with Secondary Control (English, Romanian summary).............. 

ALEXANDRA TOMA and CORNEL CIUPAN, Man Machine Interface 

Ergonomic Design Related to Human Factors (English, Romanian 

summary)..................................................................................................... 

LILIANA TOPLICEANU, ADRIAN GHENADI and MARIUS PASCU, 

Study about the Role of Accumulators in Hydraulic Secondary Control 

Systems (English, Romanian summary).................................................... 

CĂTĂLIN UNGUREANU, IRINA COZMÎNCĂ and RADU IBĂNESCU, 

Considerations about Control Charts Used in Statistical Process Control. 

(II) Control Charts for Infrequent and Cumulative Data (English, 

Romanian summary) .................................................................................. 

DĂNUł ZAHARIEA, Functional Diagrams for Modeling the Reservoir 

Emptying Process through a Small Emptying Orifice (English, 


DĂNUł ZAHARIEA, Functional Diagrams for Modeling the Reservoir 

Emptying Process through an Emptying Pipe (English, 


DAN POPESCU, A Pleading for the Modernization of Machine Tools in 

National Industry (English, Romanian summary)...................................... 

169 

177 

185 

191 

197 

203 

211 

219 

227


Publicat de 

Universitatea Tehnică „Gheorghe Asachi“ din Iaşi, 

Tomul LVIII (LXII), Fasc. 1, 2012 

SecŃia 


ANALYSIS BY NUMERICAL SIMULATION OF A NEW SELF- 

ADAPTIVE HYDRAULIC SYSTEM USED AT WIND TURBINES 

BY 

MIHAI AFRĂSINEI 1 , DORU CĂLĂRAŞU 2 , VASILE DAMASCHIN *1 , 

IRINA MARDARE 2 and ADRIAN OLARU 2 

”Gheorghe Asachi” Technical University of Iaşi, 

1 Department of Machine Tools 

2 Department of Fluid Mechanics, Hydraulic Machines and Drives 

Received: September 12 2011 

Accepted for publication: September 20, 2011 

Abstract. The paper presents the methodology and the results from the 

analysis by numerical simulation of a self adaptive hydraulic transmission in 

closed circuit for low power wind turbines. In order to analyze the simulation 

model, realised using Simulink, the pump drive speed was considered as input 

value. Based on the resulted unit step responses, the system has a good working 

behaviour, stable dynamically to the wind speed variation and also to the load 

variations at the motor shaft. 

Key words: hydraulic transmission, wind turbine, Simulink, unit steps. 

1. Introduction 

The wind energy transmitted hidraulically to the ground represents an 

actual research trend from complex programs, dealing with the non 

conventional energy sources. The wind turbines with horizontal axis and low 

power, fitted with adaptive hidraulic transmissions, may run with variable speed 

(Bej, 2001), (Spera, 1994). 

* Corresponding author: e-mail: vasile.damaschin@hydramold.ro

2 Mihai Afrăsinei et al. 

The wind turbines run under a rigorous control of speed and power at the 

electrical generator shaft. Specialized literature relieved two ways of wind 

turbine control and running when the wind speed and/or its direction fluctuates 

(Bej, 2001): 

i) By maintaining constant the speed value of the turbine axis, using 

different solutions in order to modify the incidence angle of the pales or to 

determine the removing of the air flow on the pales. 

ii) By maintaining constant the speed value of the generator shaft using 

adaptive hydraulic transmissions. 

First method is recommended especially for the medium and high power 

turbines, connected to a local or national network, because this type of solutions 

are increasing the turbine complexity, the costs and the static and dynamic 

loads, determining loss of reliability and great maintenance costs. For low 

power turbines the second method is preferred, using adaptive hydraulic 

transmissions which could give good results when the wind speed varies into 

acceptable limits (Bugarschi & Galeriu, 1997). 

The authors are proposing a structure for a closed circuit hydraulic 

transmission, self adaptive. This transmission is then studied when running in 

dynamic regime by analyzing the unit step responses at step variations of the 

pump speed and consumer´s load. 

2. The Structure of the Adaptive Hydraulic Transmission 

The hydraulic transmission with closed circuit and self adaptive control 

has the structure presented in Fig. 1. The transmission module consists in the 

pump 12 with variable flow and the hydraulic reversible motor 16, connected in 

closed circuit. The pump 12 is a double pump with low flow, which supplies the 

closed circuit, in order to compensate the oil losses from the circuit and to 

command the flow used in the circuit for setting up the pump 12 disc inclination 

angle. 

Simulation of the wind speed variation, which determines a speed 

variation at the pump 12 axis, is done with the capacity motor 13, remote by the 

frequency converter CF. Simulation of the load at the rotary hydraulic motor 

shaft 16 is realized through the loading module. The loading module consists by 

the rotary hydraulic motor 17, running as a pump and driven by the axis 19. 

The load value for the motor 16 to drive the pump 17 is achieved using 

the outlet 20. Its output pressure is measured with the transducer 18. The flow 

for supplying the pump 16 is realized through the flow transducer 21. 

The hydraulic transmission may run in adaptive regime only if the speed 

at the admission axis 19 is constant. If this condition is satisfied, the adjustment 

of the electrical generator working parameters to the network parameters is very 

easy, so no more adapting elements are necessary.

Bul. Inst. Polit. Iaşi, t. LVIII (LXII), f. 1, 2012 3 

Fig. 1 – The structure of the adaptive hydraulic transmission.


3. Numerical Simulation of the Adaptive Hydraulic Transmission 

In order to complete the numerical simulation of the adaptive hydraulic 

system the simulation model was realized, using MATLAB and Simulink 

(Călăraşu et al., 2008). To analyze the model´s behavior and, implicitly the 

system´s functioning, the input value was considered the pump drive speed and 

a signals simulation block was provided for it. The simulation model for the 

hydraulic system is presented in Fig. 2. 

Fig. 2 –The simulation model of the adaptive hydraulic transmission. 

The aim of the experimental tests by numerical simulation is to 

determine the unit responses of the hydraulic adaptive system at a step signal 

variation. The reference value for the simulation tests is the speed of the shaft 

19 and the input values are the speed drive of the pump 12 and the load at the 

hydraulic motor shaft 16 (Fig. 1). 

4. Unit Step Responses at Step Variation of the Pump Drive Speed, 

with Reference Value of 130 rad/s and Constant Load 

Fig. 3 presents the unit step responses for the variation of the angular 

velocity ωM(t) of the hydraulic motor (Fig. 3a), of the pump flow rate with 

variable unitary volume QP(t) (Fig.3b), of the pump control element stroke c(t) 

(Fig.3c) and of the pressure drop on the hydraulic motor ∆p(t) (Fig.3d) at load 

step variations M of the rotary hydraulic motor. The adaptive hydraulic 

transmission must ensure a constant speed of the motor axis ωM = const. at step 

variations of the pump speed.


a. b. 

c. d. 

Fig. 3 –The unit step responses at hydraulic motor load variation. 

From Fig. 3a, presenting the unit step response of the motor axis for 

constant speed ωP=const. at pump axis and two load variations ∆M1, ∆M2 we 

can conclude that, when load step occurs, the response is periodically damped, 

being stabilied on the reference value. If the pump speed changes, the variation 

tendency of the motor speed – registered by the speed transducer – is 

transmitted as a reaction signal to the servo mechanism, which modifies the 

stroke of the variable flow rate pump drive element. The increasing of the pump 

speed leads initially to a growing of the flow rate, the pump drive element 

stroke is then modified in order to decrease the flow rate and correct the speed 

error. This way, the motor axis speed remains constant. 

From the unit responses we find out that, at speed increasing a tendency of 

motor speed growing occurs and then it come back to the reference value. The 

system is stable of oscillatory damped type. 

Fig. 3b presents the unit step response of the pump flow rate QP(t) for 

constant load at hydraulic motor M=const. The reference value for ωM is 

constant. Two step variations ∆ωP1, ∆ωP2 for the pump axis speed are


considered. The conclusion is that when the speed increases a tendency of pump 

flow growing occurs, but after that it comes back to the initial value. The 

hydraulic motor supply flow remains at the imposed value so the speed keeps 

constant ωM=const. The functionning regime is stable, of oscillatory damped 

type. 

Fig.3c presents the unit step response of the stroke variation c(t) for M= 

=const. The reference value for ωM is constant. The experiments are achieved 

for two values ∆ωP1, ∆ωP2 of the pump speed. Analyzing the dynamic regime 

we can conclude that the system is stable, of oscillatory damped type. When the 

speed increases, the control element stroke value is diminishing. 

Fig 3d presents the unit step response for the pressure drop ∆p(t) when the 

load is constant M=const. Two step variations ∆ωP1, ∆ωP2 for the pump axis 

speed are considered. The reference value for ωM remains constant. The 

pressure drop on the rotary hydraulic motor does not change. For the inferior 

limit of the speed value the regime has a tendancy of instability. 

4. Conclusions 

1. The numerical simulation of the new hydraulic system confirmed a 

good functioning, stable at the wind speed variation (variable speed at the pump 

axis) and also at load variation of the generator motor axis. 

2. The numerical analysis aims to determine the evolution of unit step 

responses of the angular velocity at the hydraulic motor, of the flow for the 

variable flow pump, of the control element stroke and of the pressure drop on 

the rotary motor at the step variation of its load. 

3. The analysis of the dynamic regime stressed for the considered 

experimental conditions tha hydraulic system stability of oscillatory damped 

type. 

REFERENCES 

Bej A., Optimizarea construcŃiei turbinelor eoliene cu autoplafonare de putere şi 

frânare aerodinamică. Teză de doctorat, Universitatea „Politehnica” din 

Timişoara, 2001. 

Bugarschi A., Galeriu C.D., La simulation des sillages des agrégats éoliens par des 

modèles statiques à tourbillon. Buletinul ŞtiinŃific şi Tehnic al UniversităŃii 

„Politehnica” din Timişoara, s. Mecanica, 42 (56), Timişoara (1997). 

Călăraşu D., TiŃa I., Ciobanu B., Scurtu D., Simulation Results for a Hydrostatic 

Transmission for Use in Association with a Wind Turbine. Proc. of the 

International Conference on Hydraulic Machinery and Equuipments, HME 2008, 

Timişoara, Romania; The Scientific Bulletin of Politehnica University of 

Timişoara, Transactions on Mechanics, 53 (67), Special Issue, 13-18 (2008).


NăşcuŃiu L., Banyai D., OpruŃa D., Control System for Hydraulic Transmissions 

Specific to Wind Machines. Ed. Politehnica, Timişoara, 2010, pp. 475-482. 

Spera D.A., Introduction to Modern Wind Turbines/ Wind Turbine Technology. ASME 

Press, New York, USA, 1994. 

ANALIZA PRIN SIMULARE NUMERICĂ A UNUI SISTEM HIDRAULIC 

AUTOADAPTIV DESTINAT TURBINELOR EOLIENE 

(Rezumat) 

În lucrare sunt prezentate metodologia şi o serie de rezultate obŃinute la analiza 

prin simulare numerică a unei transmisii hidraulice în circuit închis, autoadaptive, 

destinată acŃionării turbinelor eoliene de putere mică. Pentru analiza comportării 

modelului de simulare, conceput cu ajutorul bibliotecii de functii Simulink, s-a 

considerat ca mărime de intrare turaŃia de antrenare a pompei. Pe baza răspunsurilor 

indiciale obŃinute se poate constata buna funcŃionare a sistemului, stabilă, în regim 

dinamic, atât la variaŃia vitezei vântului, cât şi la variaŃia sarcinii la arborele motorului 

generator.


Publicat de 

Universitatea Tehnică „Gheorghe Asachi” din Iaşi 


SecŃia 


DYNAMIC MODELS FOR TRACTION SYSTEMS 

BY 

ADRIAN SORIN AXINTI ∗ and GAVRIL AXINTI 

Received: August 24, 2011 


“Dunărea de Jos” University, GalaŃi 

Department of ŞtiinŃa şi Ingineria Materialelor 

Abstract. The work presents a classification of traction systems as a function 

of structure, the systems being classified into: integrally mechanic-traction 

systems–STIM, mechanic-hydrostatical traction -STMH and integrally traction 

hydrostatical systems STIH. Each structural model contains the main components 

from dynamic point of view: the heat engine as the source of energy, the 

transmission of the traction and the system for movement formed of wheels with 

tyres or of self-propelled equipment. For each structural model is the dynamic 

model to which dynamic analyses should be done by excitation accomplished by 

the road (the runaway). The necessity of these models has been due to the study 

models of dynamic behavior and of traction systems comparison of different 

structures, mechanic, mechanic and hydraulic. 

Key words: dynamic model, transmission, traction, hydraulic. 

1. Classification of Traction Systems 

Traction systems of the consulted technologic equipments are complex 

systems formed of a large number of constitutive hydrostatic, mechanic, or 

combinations of these components which reciprocally interacts in the aim of 

equipment movement in different conditions of movement. Dynamic modeling 

for all the exploitation situations is unaccomplished, at least with the conditions 

and available modeling methods this moment, the study of dynamic behavior 

used dynamic simple models with one or two degrees of freedom, when the 

traction system is substituted by one, two or three concentrated masses 

∗ Corresponding author: e-mail: axinti@ugal.ro

10 Adrian Sorin Axinti and Gavril Axinti 

connected between them with elements of transmission being considered elastic 

systems by a certain rigidity. These models permit as the influence to be studied 

in dynamic behavior of certain components of the transmission (couplings, 

gearbox, cardan drive, differential, the wheel, and the tyre). 

A step towards the generalization of dynamic models for the traction 

systems constitutes the models with a finite number of freedom degrees, which 

are to be considered the factors of amortization entered by the components of 

the system in the model besides the rigidity of the mechanical components 

which compose the traction system structure. The parameters characterizing 

these systems are the rigidities of structural elements written with kij and the 

amortizations of the same elements written with cij. 

In order to accomplish the mathematical models upon which to be studied 

the behavior of the system to excitations produced by the dislevels of the 

runaway it is necessary the realization of a classification of the characteristic 

types of traction systems presented, taking count of the dynamic characteristics. 

The models must answer to the following requirements: 

i) to take count of dynamic features produced by the components of the 

mechanic transmission and the way of coupling of these (gearings, arbors, 

planetary cardans, couplings, etc.); 

ii) to take count of dynamic features produced by the components of the 

hydrostatical transmission and the way of coupling of these (pumps and 

hydraulic motive rotative printing press, hydraulic meshes, the hydraulic usedup 

environment as the agent of thing, etc.); 

iii) to take count of rigidities and the linear and angular amortization 

being introduced by the organ of movement (tyre or caterpillar) to excitation 

produced by the run way; 

iv) to take count of characteristics visco-elastic of environment from 

which is composed the run away and the degree of its deformability; 

v) to consider the kinematic excitation produced by the dislevels of the 

run away, through the resistant moments of the equipment wheels; 

vi) to take count of the touch moment of the equipment adherence, and 

therefore of the skids of the moment organ in report with the way; 

vii) to take count of the moment angular-speed feature, from the external 

feature the source (thermic engine) and the regulation laws of waterworks; 

vii) to accomplish a simple model but conclusive must estimate, by 

numeric modeling, the influence of the road to dynamic behavior of traction 

system; 

viii) to achieved models for typical traction systems which permit to 

compare to miscellaneous solicitations inducted by the run away. 

1.1. The Complete Integral Mechanical System –STIM 

This traction system is characterized by the ensemble achieved between


the heat engine, as source of energy and the organ of movement of the 

equipment, formed of tyres or caterpillar, is exclusively composed from 

constitutive mechanic in the shape of systems of denticulate wheels, axe and 

organized arbors in specific components (gearboxes, planetary reductors, 

differentials, transmissions cardanics, couplings, etc). This organization of the 

scheme causes the touch of functional parameters bankables for systems but and 

a certain dynamic behavior of this caused by the dynamic parameters of the 

system (moments of inertia, features of rigidity and amortization, etc). The 

tipology of the systems is rendered in Fig. 1. What comprise most the complete 

systems of used-up mechanics for the actuation of technological equipments. 

(Axinti, 2004), (Boazu, 1998), (Gilespi, 1992). 

Fig. 1 – The model of mechanical integral traction systems − STIM: 

a − system with a power line; b − system with the many power-lines; 

MT − thermic engine; Cv − mechanic transmission with denticulate wheels of speeds; 

Tc cardan transmission; Df − differential; OD/SD organs/systems of movement. 

1.2. The Mechanic-Hydrostatical Traction System STMH 

It is characterized as the fact the accomplished system between the heat 

engine as the source of energy and the organ of movement of equipment, 

formed of wheels or caterpillar, it is formed of mechanical as those presented 

before to which are added hydrostatic components as pumps, engine, apparatus 

of casting and protection, hydraulic meshes etc. 

CD 

MT CD 

TC 

TC 

P M 

P 

P 

Fig. 2 – The model of mechanical traction systems – STMH: a − equipments on tyre; b 

− equipments on caterpillar: MT − heat engine; CD, DF, RP − mechanic transmission 

with denticulate wheels (cable box, differential, planetary reductor); P, M − hydrostatic 

transmission; OD/SD organ/system of movement. 

M 

M 

RP 

RP 

a 

b 

SD 

a 

SD b


This, by average used-up liquid as the hydraulic agent, conduces to the 

conduction of the energy of the system to the desirable parameters, but through 

own dynamical characteristics influences the dynamic behavior of draft system 

of the equipment. Elastic characteristics and amortization of hydraulic agent and 

of hydraulic apparatus are added to characteristics of the used-up components, 

causing the dynamic behavior to the whole traction system. The typology of 

these systems is rendered as in Fig. 2 which includes the most mechanical 

systems used-up for the action of technological equipments. (Axinti, 2004), 

(Boazu, 1998), (Gilespi, 1992). 

1.3. The Complete Hydrostatic Traction System STIH 

It is characterized by the action between the heat engine as the source of 

energy and the organ of movement of the equipment, formed of wheels or 

caterpillar it is composed by exclusively from hydrostatical coupled round open 

or closed components. Dynamic characteristics of the system are influenced by 

the dynamic features of hydrostatic used-up components, by the command way 

and regulation of these components and of the way of coupling in system. The 

typology of this system is rendered as in Fig. 3, which includes the most of 

integral hydrostatical used-up systems for the action of technological 

equipments (Axinti, 2004), (Borkowski, 1996), (Boazu, 1998). 

MT P P 

Fig. 3 – The model of hydrostatical systems traction − STIH. MT − heat engine; 

P, M − hydrostatic transmission; OD/SD organ/system of movement. 

In the case of this system, the heat engine acts directly the pump or the 

volumic pumps (mount the tandem), and the volumic engines, acts directly the 

organ of movement of the equipment. The dynamic behavior of draft system is 

to influence the dynamic behavior of hydrostatical link (pump). 

2. The Dynamic Fashions 

2. 1. Structure of Dynamic Suggested Model for - STIM 

The integral mechanical traction system − STIM, having the same 

mechanical components, is reduced down to a dynamic equivalent to model two 

degrees of freedom, one represents the heat engine of the equipment, as source 

of autonomous energy, and the other of the organ of movement of the 

equipment, respectively the wheel with its tyres or caterpillar. Between these 

two ascertainable components is interposed the transmission mechanical of the 

equipment to characterize the equivalent rigidity Ktr of the factor of equivalent 

amortization of the transmission Ctr. The system achieved is reduced to the 

M 

SD 

M SD


motive thermic axle, characterized of the angle of rotation. The moment of 

inertia mechanic Js is similar reduced to the motive thermic axle and contains 

the own moment of engine but and the moments of components of components 

of the mechanical transmission. The moment Ms represents the active moment 

applied to the transmission and results from external characteristic thermic 

engine definite to a formal law Ms=f ( ϕɺ s ). The moment of inertia mechanic JR 

represents the moment of inertia of the organ from the same axis of the 

equipment (the wheels with tyres from the caterpillars). To this moment of 

inertia is added the moment of reduced inertia of the equipment as far as Ju the 

touch of the moment of adhesive MA (Fig. 4). In the same way proceed to 

equipments with many axe assets. (Fig. 4b). In the case of the gearboxes with 

many stairs of brave the features of rigidity, amortization, reduced moment of 

inertia and the active moment shall be evaluated for each step and used 

accordingly in the dynamic model. 

J S 

J S 

MS; ϕ 

S 

K TR 

C TR 

K TR 

C TR 

K TR 

C TR 

TR 

C 

K TR 

JJR R 

JR 

JR 

J R 

MR; ϕ 

R 

Fig. 4 – The dynamic model for the complet mechanical systems – STIM: 

a − system with a power line; b − system with many power lines. 

JS,MS, φS − dynamic parameters of heat engine h; JR, MR, φR − dynamic parameters of 

the organ of movement (wheels or caterpillar); JU, MU, φU - dynamic parameters of the 

equipment; KTR, k φ - the transmission of the mechanic and the active element of the 

system of movement (tyres); cTR, cφ − the factors of proper amortization of the 

transmission and systems of movement; MA of adherence to the organ of movement of 

the equipment. 

In the situations presented, mostly the aspect of accomplishing of the 

dynamic model is adverted to the determination rigidities and the factors of 

equivalent amortization have the components of used-up mechanical in the 

transmission mechanic. Shaping the traction system is considered for the case 

Kϕ 

Cϕ 

Kϕ 

M A 

Cϕ 

MA 

Kϕ 

Cϕ 

C 

Kϕ 

ϕ 

M A 

M A 

J U 

MU; ϕ 

U 

JU a 

MU; ϕ 

U b


presented the in Fig. 4 of equipment with an only motor deck and four wheels 

with tyres. 

2.2. The Structure of Dynamic Model Suggested for STMH 

The mechanic –hydraulical traction system-STMH, is reduced to a 

dynamic equivalent model formed of two or many systems with two or three 

degrees of freedom, which models the dynamic behavior of the mechanical 

components (gearboxes, reductors, etc), bound between them through 

hydrostatical components (pump) (Fig. 5). 

J J J 

s 

J Jp 

m Jr 

k cd 

ktm 

Kϕ 

C cd 

P M 

C 

tm Cϕ 

MA 

Ms Ms Ms s M 

p p 

M 

m m r Mr r M 

u u a 

Ju 

Jm Jr 

J 

s 

J p 

k cd 

C 

cd cd cd cd cd cd 

M MMs Ms Ms Ms s s M 

p p 

P M 

P M 

Fig. 5 – The dynamic model for the mechanic –hydraulical system-STMH: 

a − models with one power line; b − model with two power lines. 

For the one-track energetic systems, as are the equipments on tire with 

only one motor deck (4×2) or the equipments on tire with two or many motor 

decks (4×4; 6×4; 6×6), to which hydrostatical link (P-M) is interposed between 

source of energy (MT) and the draft system achieved whole mechanic, the 

suggested model is one from Fig. 5. The heat engine and the inclusive cable box 

are modelled as a system with two degrees of freedom, characterized of rigidity 

and the factor of equivalent amortization, which acts primary constitutive the 

hydrostatical pump. The draft system, inclusively the system of movement, is 

modeled as a system with three degrees of freedom, set secondary constitutive 

the hydrostatical motor-M system. The mechanical components of the 

transmission are modeled as visco-elastic elements to characterize the 

equivalent rigidity and the factor of equivalent amortization ctm. By dint of these 

is set the organ of movement to characterize elements sequence of the rigidity 

and the factor of amortization kφ cφ, by dint is set g the equipment (Ju). 

In the case of draft systems with two power lines, which is the case of 

fitting-out of the equipments on caterpillars (bulldozers, chargers, dredgers, 

tractors, special equipments, etc), the equipments on tire with direction through 

k tm 

C tm 

M 

m m rM 

r r 

J m J r 

k tm 

C 

tm tm 

M 

m m r M r r 

Kϕ 

Cϕ 

Kϕ Kϕ 

Cϕ 

Ju 

MA 

MA MA 

M u 

u b


side-slip (chargers, multipurpose machines, etc) or the equipments with 

tractional axle (active trailers, transport platforms, etc), the model described 

previously is reprographic with the number of power lines, in the case from Fig. 

5b, with two power lines are enforced to the solicitation inducted of the 

runaway of the two organs of equipment run (deck with tyres or caterpillar). 

And in the case of the loss state analysis of adhesive the equipment is to 

consider through the touch boundary value of adhesion -MA. 

2.3. Structure of Suggested Dynamic Model for STIH 

The draft complete system hidrostatic-stih, is modeled as a dynamic 

system with two degrees of freedom, to which the connections to the dynamic 

inertial components of heat engine and of the system of run (tyres or 

caterpillars) are realized of hydrostatic components, without pure mechanic 

components (denticulate wheels, couplings etc). 

J 

s 

J 

s 

Ms Ms Ms Ms Ms Ms s 

Ms Ms Ms Ms Ms Ms 

s 

P M 

P P 

Jr 

u b 

Fig. 6 – The dynamic model for the complete hydraulics - STIH system: 

a − models with a single power line; b − model with two power lines. 

In the dynamic model is considered an only visco-elastic connection, or 

except elastic realized of the organ of movement of the equipment (wheels with 

tires or caterpillars). This connection is interposed between the run away and 

the equipment, being the components of the model through with is inducted he 

draft force produced by the system of action 


1. The draft systems of technological self-propelled equipments are 

characterized through three types of structures, which cover most of the 

C 

M 

r r 

M 

M 

Jr 

k 

C 

k 

JMr 

r r 

k 

C 

M 

r r 

MA 

MA 

MA 

Ju 

Ju 

u 

M u 

u a 

M u


practical situations. One can see the discrepancies which characterizes the 

dynamic models of STIM, STMH, STIH. 

2. For each of ascertainable models achieved numerical analyses the 

experimental and the results are presented in another scientific works of the 

main author whose conclusion is detached as for shaping of the process in a 

draft system of self-propelled equipments age necessity of the model. The 

conclusions contain the elemental structures of the equipment: heat engine of 

the draft system formed of transmission and the system of movement (the 

wheel, caterpillar) runaway. 

3. The runaway constitutes the factor of excitation of draft system to 

miscellaneous disturbances produced by the dislevels, the states, the humidity, 

consistence etc. 

REFERENCES 

Axinti G., Contributii la modelarea proceselor dinamice din actionarea hidrostatica a 

sistemului de deplasare a utilajelor tehnologice autopropulsate. Proc. of the 6 th 

Int. Conf. on Hydraulic Machinery and Hydrodynamics, Timişoara, 2004, pp. 

292-298. 

Borkowski W., Konopka S., Prochowski L., Dynamika maszyn roboczych. Podreczniki 

Akademickie. Wydawnictwa Naukowo- Techniczne, Mechanika, Warszawa, 156- 

168,172-185 (1996). 

Boazu D., Contributii privind analiza vibraŃiilor provocate de angrenaje. Teză de 

doctorat, Univ. “Dunărea de Jos” din GalaŃi, 1998, pp. 32-41. 

Gilespi T., Fundamentals of Vehicle Dynamics. Society of Automotive Engineers, 

Warrendale, USA, 1992. 

Mladin Gh., Maşini de tracŃiune şi sisteme de transport. Vol. I, II, Ed. Impuls. 

Bucureşti, 1999. 

Untaru M., Pereş Gh. et al., Dinamica autovehiculelor pe roŃi. Ed. Didactică şi 

Pedagogică, Bucureşti, 1981. 

MODELE DINAMICE PENTRU SISTEME DE TRACłIUNE 

(Rezumat) 

Se prezintă o clasificare a sistemelor de tracŃiune după structura acestora. Sistemele 

sunt clasificate în: sisteme de tracŃiune integral mecanice –STIM, sisteme de trac- 

Ńiune mecano-hidrostatice – STMH şi sisteme de tracŃiune integral hidrostatice − STIH. 

Fiecare model structural conŃine componentele esenŃiale din punct de vedere dinamic: 

motorul termic, ca sursă de energie; sistemul de transmisie al sistemului de tracŃiune; 

sistemul de deplasare format din roŃile cu pneuri sau şenilele utilajului autopropulsat şi 

calea de rulare. Modelele au permis autorilor să analizeze comportarea dinamică produsă 

de denivelările drumului, consistenŃa drumului, aderenŃa drumului, etc. Necesitatea 

realizării acestor modele s-a datorat nevoilor de studiu a comportării dinamice şi de 

comparare a sistemelor de tracŃiune de diverse structuri, mecanice, mecano-hidralice şi 

hidraulice.


Publicat de 



SecŃia 


MODELS FOR THE HYDRAULIC 

SEISMIC ENERGY DISSIPATERS 

BY 

GAVRIL AXINTI ∗ and ADRIAN SORIN AXINTI 



“Dunărea de Jos” University, GalaŃi 

Department of ŞtiinŃa şi Ingineria Materialelor 

Abstract. This article refers to the complex seismic energy dissipaters based 

on hydrostatic equipments that are capable of destroying the energy of the 

earthquake, shock or vibration, for various frequencies, dissipaters that connect 

the body system object to the dynamic phenomenon with its ground basis. 

Key words: models, hydraulic, seismic energy, dissipater, etc. 


At national and international level there are a lot of researches on the 

developing and perfecting methods for insulating, buffering and protecting the 

structures and the human beings to the effects of shocks, vibrations and 

earthquakes. We know methods of insulating or absorbing the technological 

shocks and vibrations, based on the same principles. 

There are recent European researches, as ECOLEADER programme, 

conducted between 2001-2005 by an university consortium made of Patras 

University (Greece), Roma 3 University (Italy), Pescara University (Italy), 

Ancona University (Italy), FIP Industriale (Italy), TARK (Great Britain). Such 

studies show the actuality of the researches in the field of annihilating the 

effects of the previously specified dynamic phenomena (PătruŃ et al., 2005). 

∗ Corresponding author: e-mail: gaxinti@ugal.ro

18 Gavril Axinti and Adrian Sorin Axinti 

2. The Description of the Hydraulic Seismic Energy Dissipater 

The hydraulic energy dissipater is made as a cylinder, 1 with two 

chambers separated by a piston, 2, chambers filled with a viscous environment 

(synthetic oil, silicon oil, particles in suspension, liquid metals etc.). The piston 

is connected to a bar, 3 and a compensation tube that crosses on either side of 

the piston the two chambers. A third chamber, 4, is situated inside the 

compensation tube and has the role of compensating the dilatation or the 

contraction of the viscous environment under thermal effect. The dissipater’s 

bar is connected at the base of the construction (I) and the body is connected at 

the foundation of the construction, (II) with plane or spatial joints, 5. The device 

works reversibly to the alternative traction and compression movements and the 

dynamic behavior depends on the instantaneous frequency (speed) of the 

excitation generated by the earthquake, mechanical shock or vibration. (Fig.1). 

This action is generated by a system of pressure adjusting equipments 6, 7, 8 

(energetic regulators) created for various dissipation domains, equipments 

placed outside the dissipater. The opening of a regulator or more, or their 

closing, depends on the commands given by an electronic monitoring system of 

the earthquake and technological shock or vibration, according their frequency 

and intensity, which leads to various stages of dissipation. 

The construction of the linkage is presented in Fig.1. 

Fig.1 – Construction of the hydraulic energy dissipater. 

3. The Mathematical, Nonlinear Model of the Dissipater 

The dynamic model of the energy dissipater results from the flow 

equation and from the equation of the forces applied on the piston (d’Alembert 

principle). The non linear mathematical model is described by Eqs. (1) 

where 

4 

5 

6 7 8 

2 

dy D 

= 

dt A 

3 H 

p − 

A 

V0 

dp 

p + , 

AE dt 

2 

d y C dy 

K g A 

+ + y = − p , 

2 

dt M dt 

M 10 M 

3 

5 

(1)

Bul. Inst. Polit. Iaşi, t. LVIII (LXI I), f. 1, 2012 19 

2 3 

π d 2 

D = and 

4k ξρ 

r 

2 

H = πdδ , 

ξρ 

represent the dissipater constants, where: d is diameter of the pressure regulator 

valve; δ is the passive run of the pressure regulator until the total opening of 

the regulator; kr – the rigidity of the dissipater pressure regulator; ξ is the 

coefficient of local load losses from the hydraulic agent circulation through the 

2 

dissipater regulator; a = πd 4 is area of the front surface of the pressure 

regulator; K – the rigidity of the hydraulic dissipater; C is the absorption factor 

of the dissipater; M is the mass suspended on the dissipater; A is the area of the 

frontal face of the hydraulic dissipater piston; V0 is the volume of hydraulic 

agent (mineral oil) situated between the piston ad the dissipater regulator; 

ρ, E are the density and the elasticity of the hydraulic agent. The model 

variables are y–momentary movement of the hydraulic dissipater piston; p is the 

momentary regulated pressure. 

4. Numerical Case 

For the numerical case presented, with the values: kr=104 daN/cm, 

d=0.45 cm, ρ= 0.0009 kg/cm 3 , ξ = 1.8 , A=115.4 cm 2 , V0=3460 cm 3 , E=16900 

daN/cm 2 , M=60000 kg, g=9.81 m/s 2 , δ = 0.9 cm, K=8 daN/cm, C= 

y ∈ 0,y 

. We have the 

=0.001daNs/cm, pmax=700 bar, p∈ [0, pmax], y [cm], [ ] 

dissipater variables as functions of time according the following equations 

y = y( t), p = p( t), F = F( t) = Ap( t) and F = F( y) 

. (2) 

The model presented is made under the following hypothesis: 

i) We disregarded the flow losses through gaps, because the piston sealing 

do not allow losses of volume, being self deformable on the cylinder bore. 

ii) The flow loses on the regulator between the high pressure chambers 

and the low pressure chambers are considered neglectable taking into view the 

small dimensions of the regulator compared to the dissipater cylinder. 

( d Dd ≤ 10 150 = 1 15 ), where: d – the diameter of the pressure regulator valve, 

Dd – the diameter of the dissipater cylinder bore. 

iii) The friction forces are considered to be neglectable on the hydraulic 

regulator, compared with the direct applicable forces. 

The pressure characteristic p [bar], a function of the hydraulic dissipater 

run y [m −2 ], respectively p=p(y), obtained by modeling model (1), leads to 

graphics such as those in Fig. 2a. 

0


The pressure characteristic p [bar], a function of the hydraulic dissipater 

run y [m −2 ], respectively p=p(y), obtained by experiments, leads to graphics 

such as in Fig. 2b. 

p[bar] 

600 

400 

200 

0 

20 40 60 

80 100 120 

y[cm] 

p[bar] 

600 

400 

200 

0 

20 40 60 

y[cm] 

a b 

Fig. 2 − Characteristic p=p(y): a − theoretical characteristic; b − experimental 

characteristic at constant speed v0=0.2 m/sec. 

5. The Linear Mathematical Model of the Dissipater 

We may notice from both the behavior of the experimental model and the 

theoretical model (1) that the movement process of the body of mass M, 

suspended on the dissipater, occurs with constant speed, v 0 , after going through 

the transition period. Under these terms we can consider that 

dy 

= v0 

= const. 

(3) 

dt 

With Eq. (3) the model (1), becomes 

V dp 

− + = and Ky = 0.1Mg 

− Ap − Cv0 

. (4) 

E dt 

3 0 

D p H p Av0 

dp dp 

From the first Eq. (4), changing the variable = v0 

, results 

dt dy 

dp 

AE HE DE 

dt 

V v V v V 

3 

= + p − p , (5) 

0 0 0 0 0 

and from the derivative of the second Eq.(4) with respect to y, results 

dp 

K 

= − . (6) 

dy 

A 

From the Eq. (6) we confirm that the variation of the pressure for the run 

y has an ascendant slope, meaning that the variation law of the pressure is

Bul. Inst. Polit. Iaşi, t. LVIII (LXI I), f. 1, 2012 21 

K 

p = p0 − y , (7) 

A 

where the pressure p 0 is the regulated pressure of the hydraulic regulator. 

From Eq. (5) results the expression that connects the building parameters 

of the dissipater, the parameters of the hydraulic environment and the initial 

p , from there we can deduce the dissipater’s rigidity (8), or its own 

pressure 0 

pulsation (9), that is 

ADE 3 AHE 

2 

A E 

v0V0 0 

v0V0 0 

V0 

K = p − p − ; 

(8) 

3 

AE ⎡ D p0 H p ⎤ 

0 

ω = ⎢ − − A⎥ 

. (9) 

V0M ⎢ v0 v0 

⎣ 

⎥ 

⎦ 


1. The non linear model (1) satisfies accurately enough the real behavior 

of a seismic energy hydraulic dissipater. The characteristics p = p( y) 

, 

theoretical and real are alike at least in regard to the behavior in a permanent 

working regime. 

2. Imposing the conditions specific to the permanent working regime, 

with a constant speed v= v 0 =const. , we create the linear model of the dissipater, 

and from there we obtain the descending linear character of the pressure 

variation depending on the variable run of the dissipater piston. 

3. We demonstrated theoretically that the own pulsation of the 

mechanical linkage made with hydraulic dissipaters depends on: the area of the 

hydraulic piston’s surface, the elasticity of the hydraulic environment, the 

volume of hydraulic agent between the active chamber of the cylinder and the 

pressure regulator, the suspended mass, the regulated pressure, the piston 

movement speed, the characteristics of the hydraulic regulator. The own 

pulsation characterizes the transitory regime of the rigid links connected with 

joints to form some mechanical structures, provided with hydraulic dissipaters, 

(Axinti et al., 2008), (Axinti, 2003). 

REFERENCES 

Axinti G., NedelcuŃ F., Axinti A. S., Caracteristici ale legăturilor hidraulice disipative. 

The Annals of „Dunărea de Jos” University of GalaŃi, XIV, Mechanical 

Engineering, 1224-5615 (2008).


Axinti G., Dinamica solidului rigid cu legături hidraulice. Lucrările celei de-a XXVII a 

ConferinŃe de Mecanica Solidelor, Buletinul ştiinŃific al UniversităŃii din Piteşti, 

seria Mecanica Aplicată, 1(7), 21-31 (2003). 

Denis D., Pont sur le Var à Saint-Isidore. Exemple de conception parasismique. 

Ouvrages d’art, 45, 18-28 (2004). 

Pătrut P., Betea S., Crainic L. et al., Sistem integrat de protecŃie a clădirilor la solicitări 

seismice. Buletinul celei de-a 3-a ConferinŃe NaŃionale de Inginerie Seismică, 

Bucureşti (2005). 

* * *Program “ECOLEADER”. Comunicare program internet “ Laboratoire d’ Études de 

Mécanique Séismique, 2005, pp. 1-14. 

MODELE PENTRU DISIPATOARE HIDRAULICE DE ENERGIE SEISMICĂ 

(Rezumat) 

Se prezintă un model matematic creat de autori pentru sistemele hidrostatice 

disipatoare seismice. Se deduc caracteristicile dinamice ale acestor aparate, 

caracteristici care ajută la realizarea legăturii dintre disipator şi parametrii dinamici ai 

seismului. Se realizează un studiu de caz pentru o tipodimensiune de disipator la care se 

compară rezultatele teoretice cu cele experimentale, fapt ce confirmă că modelul 

matematic reproduce suficient de bine comportarea reală.


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RESEARCH ON EXPERIMENTAL HEAT EFFECTS IN A FLOW 

WITH HARMONIC DRIVE INSTALLATION 

I. DRIVE INSTALLATION OF HARMONY 

FLOW − ASSEMBLY IN PARALLEL 

BY 

CARMEN BAL ∗1 CARMEN IOANA IUHOS 2 and NICOLAIE BAL 3 



1 Technical University, Cluj-Napoca, 

Department of Teacher Education and Training 

2 S.C. Borker S.A., Cluj-Napoca, 

3 Technical University, Cluj-Napoca, 

Department of Strength of Materials 

Abstract. Paper aims to highlight the phenomenon of heat conducted into a 

sonic installation. Heat is produced and sent away by vibration. 

Key words: sonic pressure, sonic flow, friction resistance, sonic condenser, 

temperature. 


One of the most interesting applications of sonic proposed by Gogu 

Constantinescu, is the production and transmission of heat away by vibration. 

In this paper the proposed study the lows production and heat waves 

sonic and practical implementation of a stand to make it possible to achieve this 

objective. To produce heat vibrations to build a sonic generator phase, this 

consists of a pump equipped with a moving piston and a cylinder alternative. 

Pump speed is given by a DC electric motor with variable speed. The cylinder 

leaves a pipe to a condenser (capacitive cylinder) filled with liquid steel. As 

fluid is preferably water, with a coefficient of elasticity than oil. 

∗ Corresponding author: e-mail: balcarmen@yahoo.com

24 Carmen Bal et al. 

To protect the system against rust oil was used. This capacitor can be 

considered equivalent to a capacitor of electricity called capacitor. From the 

other end of the condenser leaving a pipeline that is connected to a tube of small 

diameter, the shape of a coil spring. Tubing (resistance of friction which acts as 

an electrical resistance) is linked with a second capacitor (capacitive cylinder) 

filled with liquid. This assembly of hydraulic viewpoint is nonsense as classical 

hydraulic fluid compressibility is not taken into account (Fig. 1). 

If you take into account the liquid compressibility factor can be put in 

motion generator through a mechanism with eccentric (or rod crank), which 

produces alternative movement of the piston. As a result of the reciprocating 

piston pulsations occur in the first cylinders. Thus the tank becomes a kind of 

sonic generator. 

Sonic waves are forced to pass inside the friction resistance and capacitor 

to reach its end. Movement is possible because of compressibility energy 

transmission waves. Alternative energy via friction resistance thin tube made 

sonic friction loss, such losses caused by passing electric current through ohmic 

resistance to electricity. 

1 2 3 4 5 6 7 8 9 

M 

e 

A B C 

where 1 − DC electric motor M, 2 − proximity sensor, 3 − elastic coupling, 4 − 

pump sonic, 5, 8, 10 − pressure sensor, 6 − temperature sensor, 7 − friction 

resistance, 9 − CM - large condenser, 11 − cm - small condenser, 13 − pump to 

achieve static pressure, 14 − valve, 15 − oil tank, for static pressure. 

The installation is complete, multifunctional, and allows, starting from a 

sonic source, determination of thermal effects. 

2. Experimental Research in the Harmonic Parallel Installation 

Research focused on obtaining experimental heat effect as a result of heat 

A B C 

A B C 

10 11 12 13 

Fig. 1 – Block diagram of sonic installation: 1 − DC electric motor M, 2 − 

proximity sensor, 3 − elastic coupling, 4 − pump sonic, 

14 

15


transmission remote vibration (sonic waves in liquids). These studies were 

conducted on the stand presented in Fig. 1, starting at different frequencies of 

the engine that drives the piston sonic generator. For each frequency 

measurements were performed for various static pressure in the system (0.25, 

0.5) Pa. 

Stand in Fig. 1, is small capacitor mounted in parallel with the resistance 

of friction. 

Fig. 2 – Evolution mounting pressure over time for small capacitor in parallel. 

After processing the files with experimental data, from three sensors 

mounted in the system, resulting graphics illustrating developments primary 

generator pressures and two capacitors, the shape of the graphics represented in 

Fig. 2. You can also view the generator speed (position viewed by curve 

generator). Evolutions of pressure curves reveal a phase shift between pressure 

from the pressure generator and capacitors. 

n = 600 rpm ps = 0,25E+05 Pa 

Fig. 3 – Diagram of pressures and temperature variation 

with time in static pressure of 0,25E+05 Pa.


ps = 0,25E+05 Pa 

Fig. 4 – Diagram of pressures and temperature variation of speed 

according to the static pressure of 0,25E+05 Pa. 

The graphs in Figs. 3 and 4 amounted to a static pressure of 0,25E+05 Pa 

starting a speed n = 600 rpm stabilize at 560 rpm. Pressure generator is 

stabilizing after a minute of starting the installation around 70E+05 Pa and the 

pressure from the large cylinder 50E+05 Pa, pressure drop across the resistance 

friction being 30E+05 Pa. The surface temperature of frictional resistance 

increased after about 1 minute to 85ºC, (Bal, 2006). 

The results noted with graphics: ∆G − sonic pump pressure variation on 

the sensor 5; ∆S1 − pressure variation obtained from pressure sensor 8; ∆S2 - 

pressure variation obtained from pressure sensor 10; T − temperature. 

The graphs in Figs. 5 and 6 were built for a static pressure of 0,5E+05 Pa. 

and n = 1000 rpm power. Temperature reached after about one minute and a 

half working at 86ºC continued to rise further to stabilize (Bal, 2006). Pressure 

sensor to rise around de 70E+05 Pa and at the large cylinder at a pressure of 

37E+05 Pa, the pressure drop on the resistance of friction is equal to 28E+05Pa. 

n = 1000 rpm ps = 0,5E+05 Pa 

Fig. 5 – Diagram of pressures and temperature variation 

with time in static pressure of 0.5E+05 Pa.



The analysis of diagrams for assembly in parallel in Fig. 1, allows to 

draw the following conclusions: 

1. Pressure drop across the resistance friction is around 30E+05 Pa to the 

computer is 42.5 E+05 Pa. Pressure deviation between the two is 29% 

acceptable deviation taking into account losses that occur across the system. 

2. After stabilization is constant pressure to maintain constant speed - 

static pressure in the system does not influence significantly the pressure drop. 

3. Based on other measurements we concluded that the optimal speed for 

the resistance of friction with diameter of 3mm and length 1m is comprised 

between 600 and 1000 rpm as confirmed by calculation. 

4. The two capacitors sonic, in parallel linked via a pipe with a small 

diameter acting as “capillary type hydraulic resistance”, which aims to 

transform the sonic waves produced by friction fluid environment with walls, 

into heat 

Acknowledgements. The authors would like to thank Prof. Ioan I. Pop, Ph.D for 

constructive support and help given to conducting experiments in the field of sonics. 

REFERENCES 

ps = 0,5E+05 Pa 

Fig. 6 – Diagram of pressures and temperature variation of speed 

according to the static pressure of 0,5E+05 Pa. 

Constantinescu G., The Theory of the Sonicity. Ed.Academiei, Bucureşti, 1985, pp. 7-8. 

Bal C., Caloric Effect in the Circuits by Harmonic Flow. Ed. Alma Mater, Cluj Napoca, 

2007, pp. 17, 75, 111. 

Bal C., Research and Contributions about the Drive Systems with the Harmonic Flow. 

Doctoral Thesis, Technical University of Cluj Napoca, 2006, p. 97.


Pop I. Ioan, Bal C., Marcu L. et al., The Sonicity Applications. Experimental Results. 

Ed. Performantica, Iaşi, 2007. 

CERCETĂRI EXPERIMENTALE PRIVIND EFECTELE CĂLDURII ÎNTR-O 

INSTALATIE DE ACłIONARE CU DEBITE ARMONICE 

I. InstalaŃie de acŃionare cu debite armonice – Montaj în paralel 

(Rezumat) 

Transmisiile sonice se realizează prin vibraŃii, iar la începutul secolului, se 

considera că energia de vibraŃie constituie o formă de energie degradantă care nu se mai 

poate transforma decât în căldură. Era de neconceput că dintr-un sistem de vibraŃii se 

poate obŃine lucru mecanic cu randament ridicat. 

Cercetările privind sistemele de acŃionare cu debite armonice s-a bazat pe o 

aplicaŃie interesantă a sonicităŃii care este transmisia căldurii la distanŃă într-un mod 

similar încălzirii electrice. Conducta conŃinând un lichid (de preferinŃă apa) poate 

transmite energie de la un generator la o rezistenŃă sonică care se poate găsi la o mare 

distanŃă. Aceasta este transformată în căldură într-un tub spiral în timp ce conducta 

rămâne rece. Limita la care temperatura din tub va ajunge depinde de punctul de 

fierbere al lichidului din interior care este influenŃată de presiune. 

Cercetările experimentale au vizat obŃinerea efectului caloric ca urmarea a 

transmiterii călduri la distanŃă prin vibraŃii (prin unde sonice în lichide). Aceste cercetări 

s-au realizat pe standul prezentat în Fig. 1, pornind de la frecvenŃe diferite ale motorului 

de acŃionare a pistonului generatorului sonic. Pentru fiecare frecvenŃă s-au efectuat 

măsurători pentru diferite presiuni statice în instalaŃie (0,25;0,5; 0,75; 1) şi pentru 

fiecare presiune statică câte 3 măsurători. Partea experimentală s-a realizat pe standul 

construit în diferite variante de aranjare a cilindrilor capacitivi şi a rezistenŃei de 

fricŃiune. O variantă constructivă este prezentată în aceasta lucrare şi anume varianta 

montajului în paralel (condensatorul mic montat în paralel cu instalaŃia).Important este 

ca în urma experimentărilor s-a putut demonstra ca în circuitul din Fig. 1 se obŃine 

căldură în rezistenŃa de fricŃiune, iar conductele de legătura rămân reci.


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DYNAMIC ANALYSIS OF A HYDRAULIC ACTUATION 

SYSTEM OF VERY SLOW MOVING DEVICES 

BY 

VLAD BOCĂNEł ∗1 , HORIA ABĂITANCEI 2 , CONSTANTIN CHIRIłĂ 3 , 

MARIUS DENEŞ-POP 1 and LIVIU BĂRNUłIU 1 


Accepted for publication: August 31, 2011 

1 Technical University, Cluj Napoca, 

Department of Machine Tools and Industrial Robots 

2 Transilvania University, Braşov, 

Department of Motor Vehicles and Engines 

3” Gheorghe Asachi” Technical University, Iaşi 

Department of Machine Tools 

Abstract. The paper presents research conducted for a hydraulic actuation 

system designed to propel a slow moving device. During system run, the stability 

of the system has to be assured, considering the slow moving velocity of the 

device. In order to analyse the system, a multi-domain model was developed to 

identify the influences of geometry and running conditions on system behaviour. 

In order to appreciate the influences of different running parameters, the 

propelling speed, pump displacement and motor displacement are considered. 

Instabilities at system run-up are identified and are relative rapidly damped. 

Key words: fluid power, slow moving system, multi-domain simulation etc. 


High load and slow moving devices are systems that may be actuated in 

many cases only using fluid power systems, due to their capability of 

developing high power densities and precise actuation. If the system has to 

work under moving conditions, additional requirements related to mass density, 

unexpected load changes and safety have to be fulfilled. Another important 

∗ Corresponding author: e-mail: vlad.bocanet@gmail.com

30 Vlad BocăneŃ et al. 

requirement of the system is the broad range of powers that have to be achieved. 

First design issues for such a system have to rely on important simulation data 

to assure time and resource consuming aspects. In order to meet these tasks, a 

multi-domain model was used, capable of including time-dependent phenomena 

(AMESIM, User Guide). 

2. The Actuated System 

The hydraulic powered device is a slow moving vehicle that has to meet 

different running conditions. One important task is the displacement on 

conventional running conditions, including different displacement surface 

qualities and surfaces slopes. For this case the 5 tonne device has to be moved 

on a 2% slope on a surface having the rolling resistance coefficient µ, 0,015 for 

asphalt/concrete road and 0,3 for heavy duty earth conditions. The running 

resistance that has to be covered by the propelling system is given by 

2 

v 

R = µmg cosα + mg sin α + 0,5c 

x ρA + ma 

(1) 

2 

expressed as a resistance force. Parameters in Eq. (1), have the following 

meaning: µ − rolling resistance coefficient, m – vehicle mass, g – acceleration 

due to gravity, α – road slope, cx – aerodynamic coefficient, ρ – air density, A – 

cross section area of the vehicle, v – vehicle translational speed, a – vehicle 

linear acceleration. The aerodynamic drag force, the third term in Eq. (1) is 

neglected for the slow running condition. The acceleration force may be 

significant if a given acceleration is required so that the system reaches a given 

displacement velocity in a given time or length restriction. A second, important 

requirement is the very slow moving condition in the range of 1…2 m/s, 

maintaining the moving stability. 

To assure this second task, a hydraulic propulsion system is analysed. 

3. The Multi-Domain Model of the Actuating Hydraulic System 

The model, presented in Fig. 1 includes a model 1 for the internal 

combustion engine as variable speed source, driving the variable displacement 

pump 3 using a gear box 2. The liquid flow provided by the pump propels the 

hydraulic motor having a fixed displacement 5 by the hydraulic hose 4. 

Additional gears 6 and the final system transmission gear 7 are meant to avoid 

minimum stable working speeds of the hydraulic motor, that are often 

connected with stability and lubrication issues, especially when high load 

conditions are needed. For a realistic modelling, a vehicle model 8 was used 

that allow to include rolling resistance, slope resistance 10 and to also simulate 

braking conditions by supplying a braking torque at port 9.


Fig. 1 – The multi-domain model of the hydraulic propelling system. 

4. The Simulation Results 

The main parameters were designed using conventional hydraulic 

formulae for a pre-design phase and included in the multi-domain model. The 

goal of the study was to identify system behaviour for different design and 

running parameters (Călăraşu, 2002), having as output goal and optimisation 

criterion, the vehicle displacement and velocity. 

4.1. The Influence of Hydraulic Energy Generation Conditions 

In order to identify the influence of propelling speed, different values 

were chosen. Their influence on vehicle speed is given in Fig. 2a. A 

proportional dependence can be observed. System start-up is affected be 

oscillations that are damped after approximately 2 seconds despite propelling 

speed. A second parameter that has been considered is the volume of the 

hydraulic variable displacement pump. The volume has been modified in the 

range of the actual to total displacement ranging from 0,3 to 1. Engine speed 

and displacement actuation time are modified in the same time range, 5s. The 

vehicle speed dependence on pump displacement having a maximum value of 

23 cm 3 is shown in Fig. 2b. It can be noticed that at small values of the 

displacement, system instability is present at low pump displacements. 

It can be observed that the expected stationary behaviour is confirmed. 

The dynamical behaviour is affected by oscillations at system start-up and small 

scale values of the studied parameters. Inertial effects compensate oscillation 

and system instabilities as far as significant parameters are getting to higher 

scale values. 

The causes of the unstable behaviour are the pressure oscillations 

presented in Fig. 3. It can be observed that overcoming the vehicle inertia, the 

pressure response of the hydraulic system is characterised by high amplitude 

waves of up to 120 bar, compared to system working pressure. The relative 

small absolute values of amplitudes are given by the relative high displacements 

of pump and motors that contribute also to the rapid amplitude reduction.

32 Vlad BocăneŃ et al. 

Fig. 2 – The influence of engine speed and pump displacement on system velocity. 

The translation from accelerating to constant time conditions is also 

associated with the generation of an oscillation that is rapidly damped. This 

oscillation has also significant amplitude, but together with the running in 

oscillations, they have practical known influence on system behaviour. Acoustic 

emissions are expected to be associated with this oscillations change. 

Fig. 3 – Influence of running conditions on geometrical objects: 

a – pressure distribution; b – influence of motor displacement. 

4.2. The Influence of Hydraulic Power Conversion Unit 

The effect of the hydraulic motor volume on system speed is presented in 

Fig. 3b. It can be clearly observed that the smaller displacement motor induces 

higher scale oscillations. The damping effect is supplementary confirmed by the 

pressure oscillation evolution both at system start-up and displacement 

condition. 


1. The first conclusion of the study is that stable running conditions can 

be achieved using the adopted system structure to propel the high load – low 

speed system.


2. The second conclusion reflects that a higher volume system assures 

more stable running conditions contributing to the rapid damping of the 

pressure waves. This damping effect is given by the dynamic properties of the 

liquid, like inertial and compressibility effects. These properties are more 

significant as the volume of liquid is increased. 

REFERENCES 

*** AMESIM User Guide. Release 10 

Călăraşu D., Reglarea secundară a sistemelor de acŃionare hidrostatică în regim de 

presiune cvasiconstantă. Ed. Media-Tech, Iaşi, 1999. 

Călăraşu D., Automatizarea sistemelor hidraulice. Ed. „Gh. Asachi”, Iaşi, 2002. 

ANALIZA DINAMICĂ A UNUI SISTEM DE ACłIONARE HIDRAULIC 

FOLOSIT ÎN ACłIONAREA UNUI SISTEM CARE SE DEPLASEAZĂ 

CU VITEZĂ REDUSĂ, ÎN CONDIłII DE SARCINĂ MARE 

(Rezumat) 

Dispozitivele de putere mare care necesită viteză de mişcare redusă pot fi 

acŃionate de multe ori folosind sisteme hidraulice, datorită densităŃii mari de putere şi a 

preciziei de acŃionare. În plus, dacă este vorba despre implementarea pe o aplicaŃie 

mobilă, trebuie luate în considerare alte aspecte precum masa componentelor, schimbarea 

neprevăzută a sarcinii sau siguranŃă în operare. O altă caracteristică care trebuie 

avută în vedere este domeniul de puteri ce trebuie acoperit. 

Această lucrare îşi propune să prezinte cercetări realizate pe un astfel de 

dispozitiv. Pentru a optimiza consumul de timp şi resurse, s-a realizat un model 

interdisciplinar capabil de a încorpora componentele dependente de timp. 

AplicaŃia vizată presupune punerea în mişcare a unui vehicul cu o masă de 5 

tone pe o pantă de maxim 2% cu o viteză constantă de 1...2 m/s având o distanŃă 

maximă de atingere a acestei viteze. S-a considerat atât situaŃia în care această deplasare 

se realizează atât pe beton/asfalt cât şi pe pământ. 

Modelul include motorul cu ardere internă care pune în mişcare pompa cu debit 

reglabil, conectată la motorul cu debit constant legat de puntea vehiculului prin 

intermediul unor angrenaje care au rolul de a asigura stabilitatea la mişcarea lentă sub 

sarcină mare. În model se ia în considerare rezistenŃa la înaintare, înclinaŃia pantei şi 

momentul rezistent. 

Simularea a avut ca scop analiza comportamentului sistemului pentru diferite 

configuraŃii şi optimizarea acestuia. Prin rularea simulării se confirmă regimul staŃionar 

prevăzut. Regimul dinamic este afectat de oscilaŃiile de presiune prezente la pornirea 

sistemului. De asemenea se poate observa că un motor cu o cilindree mai mică introduce 

oscilaŃii de amplitudine mai mare în sistem. 

În urma studiului se pot trage două concluzii. În primul rând se poate observa că 

această structură poate satisface cerinŃele de viteză mică şi sarcină ridicată. În al doilea 

rând se poate observa că folosind un sistem cu un volum mai mare asigură o amortizare 

accentuată a undelor de presiune şi drept urmare o funcŃionare mai stabilă.


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NEW CONTRIBUTIONS IN THE CORRELATION OF 

MECHANICAL PROPERTIES WITH THE CAVITATION 

RESISTANCE OF STAINLESS STEELS 

BY 

ILARE BORDEAŞU ∗1 MIRCEA POPOVICIU 2 , ADRIAN KARABENCIOV 1 , 

ALIN DAN JURCHELA 1 and CONSTANTIN CHIRIłĂ 3 

1 Politehnica University, Timişoara 

Department of Mechanical Machinery, Equipment and Transport 

2 Romanian Academy, Timişoara Branch 

3 ”Gheorghe Asachi” Technical University, Iaşi, 




Abstract. The paper analyzes the cavitation resistance of twelve stainless 

steels based on the correlations between the main mechanical properties 

(hardness, tensile strength and flow limit) with the 1/MDPR parameter. Data 

regarding the microstructure is presented. The microstructure is important for the 

analysis of cavitation damage. Diagrams presented in the paper also offer the 

ranking of the studied steels in terms of cavitation resistance, using the steels’ 

mechanical properties as landmark. Final conclusions show that steels with 

different values of the mechanical properties and different microstructures can 

have similar cavitaton resistances, thus helping hydromechanical equipment 

manufacturers in the selection of steels according to the hydrodynamic intensities 

of the cavitation conditions. 

Key words: cavitation erosion, mechanical properties, microstructure, 

cavitation resistance. 


An important issue of hydromechanical equipment manufacturers is the 

high price of materials, but choosing the adequate material is also important. 

Today researchers are trying to find the optimal contents of alloying elements 

∗ Corresponding author: e-mail: ilarica59@gmail.com

36 Ilare Bordeaşu et al. 

that will give mechanical characteristics with a positive influence on the 

material’s cavitation resistance. The alloying elements influence both the 

mechanical properties of the material and its microstructure. Therefore this 

paper brings contributions to the correlation of the mechanical properties of 

twelve stainless steels for hydromechanical equipment, in this case for the rotors 

and blades of hydraulic turbines exposed to intense cavitation. The equations 

and the correlation diagrams of the mechanical properties presented in this 

paper are an extension of the results obtained by (Garcia et al, 1960), and 

(Bordeaşu, 2006), through the introduction of the Cre /Nie ratio. 

2. Studied Materials. Testing Method and Used Apparatus 

2.1. Studied Materials 

Eight of the twelve experimental stainless steels (denoted according to 

the rough chromium and nickel content: Cr6/Ni10, Cr10/Ni10, Cr18/Ni10, 

Cr24/Ni10, Cr12/Ni0, Cr12/Ni2, Cr12/Ni6, Cr12/Ni10) were obtained by 

casting at S.C. Zirom S.A. from Giurgiu using the Vacuum melting furnace 

with electron flow EMO 1200 R (Fig.1), equipped with an electron canon with a 

power of 80 kW (manufactured by the Electrical Plants for the construction of 

locomotives “HANS BEIMLER”, Hennigsdorf , Germania). The other four 

stainless steels were acquired from hydromechanical equipment exploiters, as 

follows (Bordeaşu, 2006): OH12NDL – extracted from the blades of turbines 

from Iron Gates I; Stainless steel III-RNR – received from the former 

ICPRONAV Institute from Galati where it was used in experimental ship 

propellers; steels 20Cr130 and X10CrNi18/4PH were used in experiments 

regarding the manufacturing of other components exposed to intense cavitation. 

Fig. 1 – Vacuum melting furnace with electron flow EMO 1200 R. 

All the steels originate from parts or half-finished parts. The heat 

treatments of the stainless steels were achieved using the UTTIS furnace of the 

Politehnica University Bucharest. The structure of the cast parts was marked by 

high chemical inhomogeneity (segregation), due to the fact that the cooling 

process took place at high speeds and the diffusion processes could not take 

place in time.


The annealing process was used for the correction of flaws derived from 

the casting and for the preparation of the half-finished parts for later processing 

and for homogenizing. 

Before the cavitation tests, the stainless steels were subjected to solution 

hardening in order to improve their properties. 

The solution hardening treatment was conducted at 1050°C/1hour/water 

(Fig. 2). 

Fig. 2 – Heat treatment diagrams. 

Using the Schaefller diagram (Bordeaşu, 2006), the microstructures of the 

studied stainless steels were determined, Table 1 (where simplified designations 

were used in order to identify the steels in the diagrams that will follow), in 

order to determine the microstructural constitution, without analysing it’s effect 

on the cavitation resistance, but only through the Cre/Nie ratio. 

Table 1 

The microstructure according to the Cre / Nie ration 

(using the Schaefller Diagram (Bordeaşu, 2006)) 

Designation Material (Ni)e (Cr)e Structure 

% % 

N1 Cr6/Ni10 15.173 10.266 32%M+68%F 

N2 Cr10/Ni10 14.854 14.486 100%A 

N3 Cr18/Ni10 14.138 21.448 98%A+2%F 

N4 Cr24/Ni10 15.101 29.145 81%A+19%F 

C1 Cr12/Ni0 4.81 14.268 75%M+25%F 

C2 Cr12/Ni2 6.25 14.626 90%M+10%F 

C3 Cr12/Ni6 10.145 14.9 100%A 

C4 Cr12/Ni10 14.74 14.668 60%M+40%F 

E1 OH12NDL 4.45 13.2 74%M+26%F 

E2 III - RNR 6.1 16.1 50%M+50%F 

E3 20Cr130 5.4 14.05 34%M+66%F 

E4 X10CrNi18/4PH 8.125 20.23 74%A+5%M+21%F 

Table 2 presents the mechanical properties of the studied steels and of the 

standard steels.


Designation Rm 

N/mm 2 

Table 2 

Mechanical characteristics 

2.2. Testing Method and Used Apparatus 

The cavitation erosion tests were conducted on the magnetostrictive 

vibratory apparatus T1 (Bordeaşu, 2006), (Bordeaşu et al., 2007), using a 

vibratory cavitational specimen (www.astm.org/Standards). The total length of 

the tests for one specimen was 165 minutes. The tests were paused at regular 

intervals, in order to record the mass losses and to examine the surface exposed 

to the cavitation attack. 

The functional parameters of the vibratory apparatus are: power: 500 W, 

vibration fervency: 7000±0.3% Hz, double vibration amplitude: 94 µm, 

specimen diameter: 12 mm, supply voltage: 220 V/50 Hz, specimen type: 

vibratory, working fluid: double distilled water, whose temperature was 

maintained at a constant value of 21 ± 1 0 C for the duration of the tests. 

The experimental procedure complied with the ASTM standards. 

3. Results and Discussions 

Starting from the equations established by Hammitt (Hammitt et al., 

1980) (1/MDPR=C⋅HB D , respectively DPR=C⋅(UR⋅HB) D ) and with coefficients 

established by Garcia (Garcia et al., 1960): 

1 

MDPR 

1 

MDPR 

Rp0,2 

N/mm 

HB 

N1 1550 1120 406 

N2 1450 1020 371 

N3 1335 934 435 

N4 1280 901 253 

C1 1450 1020 461 

C2 1336 935.2 421 

C3 1540 1083 353 

C4 835 626 286 

E1 650 400 225 

E2 550 380 159 

E3 600 300 170 

E4 610 338 185 

0.811 

= 0.998(UR) , (1) 

1.788 

= 0.734(HB) , (2)


1 

MDPR 

2.0 

= 1.43(UR ⋅ HB) , (3) 

new equations were established for the correlation of the studied steels 

properties with their cavitation erosion resistance, expressed by 1/MDPR. 

a b 

c 

Fig. 3 – The variation of cavitation erosion resistance with mechanical properties: a – with 

the hardness; b – with the tensile strength; c – with the flow limit. 

In Table 3 one can see analytical equations for the approximation curves 

from the correlation diagrams (Fig. 3) 

Table 3 

The analytical forms of the curves from the correlation diagrams 

Figure Analitical form Coeficient C Coeficient D 

2 1/MDPR= C⋅HB D 1.54 1,77 

3 1/MDPR= C⋅Rm D 0.54 1,56 

4 1/MDPR= C⋅Rp0.2 D 70.58 1.51


The simultaneous analysis of Table 1 and Table 3, regarding the 

analytical equations and the coefficient values of the main mechanical 

properties Rm, Rp0.2, indicates that the values of the parameters remain 

unchanged, showing only small differences from the values established by 

Hammitt and Garcia (Garcia et al., 1960). These differences show that the 

established formulas can be used for the increase of the generalization degree. 

Fig. 4 – The influence of the mechanical properties and 

the chemical constitution on the cavitation resistance. 

The equations for the curves that define the cavitations resistance 

domains of the stainless steels tested on the magnetostrictive vibratory 

apparatus with nickel tube T1 are: 

i) for curve 1 

ii) for curve 2 

iii) for curve 3 

1 

MDPR 

1 

MDPR 

1 

MDPR 

−0.004778Φ = 775.1869(1 − e ) , (4) 

−0.014426Φ = 129.77(1 − e ) , (5) 

−0,04097Φ = 46.3305(1 − e ) . (6) 

Using the principles that gowern the materials cavitation erosion 

resistance, the diagram in Fig. 4 correlated the normalized resistances with the 

mechanical properties and the chemical constitution contained in a single


formula noted with ψ. This formula represents a generalization of the formula 

established by Sakai-Shima (Bordeaşu, 2006), and generalized by Bordeaşu 

(Grant CNCSIS PN II: ID-34/2007). 

Following the evolution of the curves that establish the difference 

between resistance levels, the following conclusions can be drawn: 

1. The extension of the domains from weak to super resistant is 

justifiable, because creating enhanced properties requires the application of 

adequate treatments and new technologies that result in the chemical 

compositions necessary for obtaining structural constituents with increased 

cavitation resistance. 

2. These diagrams allow the prediction of the cavitation erosion 

resistance for a stainless steel with known mechanical properties. 


1. It seems that the model proposed by Hammitt-Garcia for the variation 

of erosion resistance with the hardness and the other mechanical properties (Rm, 

Rp0,2) keeps its form, with the difference between the scale (C) and form (D) 

parameters. 

2. Eqs. (4) to (6) represent a generalization of the formulas established by 

Sakai-Shima and late by Bordeaşu, with applications for the stainless steels 

tested in the Cavitation Laboratory from Timişoara. 

3. The diagram from Fig. 4 can be used for predicting the cavitation 

behavior of stainless steels used for hydromechanical equipment just by 

calculating the coefficient ψ. 

4. The extension of the cavitation erosion behavior domains from weak to 

super resistant is justifiable because creating increased properties requires the 

application of adequate treatments and new technologies that result in the 

chemical compositions necessary for obtaining structural constituents with 

increased cavitation resistance 

5. To generate a new method for the ranking of stainless steels further 

research is required on vibratory machines with different operating parameters, 

in order to broaden the data base for stainless steels, thus allowing the increase 

of the degree of applicability for the formulas established in this paper. 

Acknowledgments. The present work has been supported from the Grant 

(CNCSIS) PNII, ID 34/77/2007 (Models Development for the Evaluation of Materials 

Behavior to Cavitation), 

REFERENCES 

Bordeaşu I., Eroziunea cavitaŃională a materialelor. Ed Politehnica, Timişoara, 

http://www.grupoogman.com/og_it_manual.html, 2006.


Bordeaşu I., Popoviciu M., Mitelea I., Ghiban B., Bălăşoiu V., łucu D., Chemical and 

Mechanical Aspects of the Cavitation Phenomena. Chem.Abs. RCBUAU, 58, 12, 

1300-1304 (2007). 

Garcia R., Hammitt F. G., Nystrom R. E., Corelation of Cavitation Damage with Other 

Material and Fluid Properties. Erosion by Cavitation or Impingement, ASTM, 

STP 408 Atlantic City, 1960. 

*** Developing Models for Assessing the Behavior of Materials under Cavitation 

Erosion. Grant CNCSIS PN II: ID-34/2007. 

Hammitt F. G., De M., He J., Okada T., Sun B-H., Scale Effects of Cavitation Including 

Damage Scale Effects. Report No. UMICH, 014456 - 75 – I, Conf. Cavitation, 

Michigan, 1980. 

NOI CONTRIBUłII ÎN CORELAREA PROPRIETĂłILOR MECANICE CU 

REZISTENłA LA CAVITAłIE A OłELURILOR INOXIDABILE 

(Rezumat) 

În lucrare se analizează rezistenŃa la cavitaŃie a 12 oŃeluri inoxidabile, pe baza 

corelaŃiilor dintre principalele proprietăŃi mecanice (duritate, rezistenŃa mecanică la 

rupere şi limita de curgere) cu parametrul 1/MDPR. În vederea analizei se oferă şi date 

despre constituŃia microstructurală, importantă în analiza distrugerii prin cavitaŃie. 

Diagramele prezentate oferă şi o ierarhizare a oŃelurilor cercetate, din punct de vedere al 

rezistenŃei la cavitaŃie, folosind ca elemente de baza propietaŃile mecanice. Concluziile 

finale arată că oŃeluri cu proprietăŃi mecanice diferite valoric şi cu constituŃii 

microstructurale diferite pot avea rezistenŃe cavitaŃionale similare, ajutând, astfel 

constructorii de echipamente hidromecanice în selectarea oŃelurilor, funcŃie de 

intensităŃile hidrodinamice ale regimurilor cavitaŃionale.


Publicat de 


Tomul LVIII (LXI I), Fasc. 1, 2012 

SecŃia 


RESEARCH ON FRICTIONAL LOSSES IN 

TENSIONING DEVICES AT FULL LOAD 

BY 

CONSTANTIN CHIRIłĂ, ANDREI GRAMA * and DUMITRU ZETU 



”Gheorghe Asachi” Technical University, Iaşi, 


Abstract. In this paper determination of size of frictional losses at full load 

friction tensioning devices is performed using theoretical calculations of loading 

force and is compared with mesured pulling force developed by tensioning 

device. It was tested a device used for reinforcement tensioning, device designed 

and realized in Hydraulic and Pneumatic Systems Engineering Department, 

Faculty of Mechanical Engineering and Industrial Management, Technical 

University "Gheorghe Asachi" Iasi. The aim of this research is to compare values 

obtained with the values given in literature for tensioning devices made by 

profile companies of the world. 

Key words: pre-tensioning, device for tensioning, frictional force, tendon. 


The stand, subject of research presented in this paper, includes a 

hydraulic tensioning device with mechanically driven blockings. 

The structure of the tensioning device used is shown in Fig. 1 (ChiriŃă et 

al., 2009). 

The construction and functioning of this device must perform the 

following functions: grip the tendon; stretch the tendon until the 

necessary tensioning force is reached; push some mantle corbels in the 

blocking /ancorage slab; unlock the gripping system on the tendon and 

remove the device off the end of tendon (ChiriŃă et al., 2009). 

* Corresponding author: e-mail: andreiasi79@yahoo.com.

44 Constantin ChiriŃă et al. 

Fig. 1 – Tensioning device with mechanically driven blockings (ChiriŃă et al., 2009): 

1 − cylindrical frame, 2 − left hollow rod, 3 − right hollow rod, 4 − piston, 5 − tendon, 6 

− bushing; 7, 8, 9 − wedge grips, tendon blocking system, 10 − pipe frame, 11 − 

bushing, 12 − abutment cup, 13 − bushed bearing with mantle corbel, 14 − guiding 

tendon pipe; 15 −flange, 16 − spring, 17 − protection bushing, 18 − shutter disk; 19 − 

nut with blocking spline. 

During pulling of reinforcement for tensioning, occure friction forces due 

to friction in the working cylinder and other moving mechanical elements which 

oppose to the tension force value. Thus, the active force developed by the 

device must be greater than real pulling force. 

The purpose of this research is to determine the size of losses at full load 

friction tensioning devices used and to compare them with values given in the 

literature, for the tensioning devices made by profile companies. 

2. Experimental Stand for Determining the Frictional Losses in 

Tensioning Devices at Full Load 

Friction forces due to friction tensioning devices are opposed to the 

piston advance. This force must be added to the effective force of tension. 

To determine the theoretical losses of frictional force for single wire 

tensioning devices, is used Eq. (1) which gives theoretical value of loading 

force, 

F = S P, 

(1) 

înc 

where: S − surface of the piston of the tension cylinder who release effective 

strain in cm 2 , P − pressure of power source for pre-stressing in bar. 

To determine the relative losses by friction force is used Eq (2), 

Fînc − Fexp 

∆ Ff 

= 100[%] , 

F 

exp 

were: Fexp is actual force developed by tension device measured with force 

transducer at the output of hydraulic pressure source working. 

(2)


In this paper was submitted to research two tensioning devices made in 

the Department of Hydraulic and Pneumatic Systems Engineering (DISAHP) - 

Department of Machine Tools of the Faculty of Machines Manufacturing and 

Industrial Management , “Gheorghe Asachi” Technical University of Iaşi. The 

devices can strain tendons force 16 tf, first working with a 200mm stroke and 

second 500 mm; both devices having surface of the piston of the tension 

cylinder S=43.3 cm 2 . 

Real pulling force (Fexp) was determined with test stand in Fig. 2. 

Stand blocks 

experimental 

�� 

�� 

Device for 

� 

tensioning 

Pressure 

gauge 

High pressure 

hydraulic unit 

Fig. 2 – Experimental stand to determine of real force (Fexp): 

TF − force transmitter, PT − pressure transmitter. 

�Pressure 

controler 

�Digitally force 

display 

�Digitally pressure 

display 

Stand components to determine of real force (Fexp) of tensioning devices 

are shown in Fig. 2 as follows: high pressure hydraulic unit for tensioning with 

pressure controller; wire TBP9; device measuring service (DMS) for calibrating 

the display of pulling force on the high-pressure hydraulic source; pressure 

transducer of 0…1200 bar; manufacturer SUCO, type 0620, accuracy ± 0.5 % 

full scale at room temperature (http://www.suco-pressureswitches.com); force 

transducer of 0…25 tf; manufacturer Aplisens, type PCE 28, accuracy ± 0.2 % 

(http://www.aplisens.pl/); pressure display digital device; force display digital 

device; power supply, 220V, AC. 

A first set of tests were performed on stand with tensioning device 16 tf 

and 200 mm stroke. 

For each pressure step, were performed three measurements resulting in 

three sets of pull force values. These values, together with theoretical load force 

are shown in Table 1. Here are presented numerical and arithmetic averages of 

three sets of experimental measured values of force and relative losses by 

friction to the device.


Table 1 

Experimental results obtained for losses of 16 tf tensioning device with 200 mm stroke 

Pressure, 

[bar] 

Fexp , [tf] 

1 2 3 

Fînc 

[tf] 

Average Fexp 

[tf] 

Relative loss 

in device, [%] 

60 2.4 2.56 2.61 2.52 2.598 2.87 

70 2.83 2.79 2.68 2.77 3.031 8.72 

80 3.25 3.27 3.26 3.26 3.464 5.89 

90 3.67 3.7 3.71 3.69 3.897 5.23 

100 4.06 4.04 4.12 4.07 4.33 5.93 

110 4.52 4.53 4.61 4.55 4.763 4.40 

120 4.9 5.03 5.03 4.99 5.196 4.03 

130 5.42 5.33 5.43 5.39 5.629 4.19 

140 5.82 5.84 5.82 5.83 6.062 3.88 

150 6.21 6.24 6.28 6.24 6.495 3.87 

160 6.57 6.56 6.6 6.58 6.928 5.07 

170 6.97 6.96 7 6.98 7.361 5.22 

180 7.32 7.41 7.46 7.40 7.794 5.10 

190 7.82 8.02 8.06 7.97 8.227 3.16 

200 8.3 8.45 8.42 8.39 8.66 3.12 

210 8.58 8.75 8.72 8.68 9.093 4.51 

220 9.01 9.12 9.15 9.09 9.526 4.54 

230 9.48 9.69 9.62 9.60 9.959 3.64 

240 9.9 10.03 10.02 9.98 10.392 3.93 

250 10.39 10.34 10.34 10.36 10.825 4.33 

260 10.78 10.81 10.82 10.80 11.258 4.04 

270 11.12 11.02 11.21 11.12 11.691 4.91 

280 11.46 11.54 11.6 11.53 12.124 4.87 

290 11.96 12 12.02 11.99 12.557 4.49 

300 12.34 12.52 12.65 12.50 12.99 3.75 

310 12.75 13 13.06 12.94 13.423 3.62 

320 13.11 13.35 13.35 13.27 13.856 4.23 

330 13.64 13.81 13.81 13.75 14.289 3.75 

340 14.1 14.12 14.2 14.14 14.722 3.95 

350 14.51 14.61 14.62 14.58 15.155 3.79 

360 14.9 15.02 15.02 14.98 15.588 3.90 

370 15.21 15.43 15.43 15.36 16.021 4.15 

380 15.7 15.94 15.93 15.86 16.454 3.63 

390 16.05 16.34 16.26 16.22 16.887 3.97 

Tests performed on the same tensioning device (16 tf) but with stroke 

500mm, were made by the same methodology as in the previous case, the 

values obtained being presented in Table 2.


Table 2 

Experimental results obtained for losses of 16 tf tensioning device with 500 mm stroke 

Pressure 

[bar] 

Fexp [tf] 

1 2 3 

Fînc 

[tf] 

Average Fexp 

[tf] 

Relative loss 

in device 

[%] 

60 2.43 2.4 2.42 2.598 2.417 6.98 

70 2.8 2.85 2.82 3.031 2.823 6.85 

80 3.22 3.23 3.23 3.464 3.227 6.85 

90 3.68 3.67 3.71 3.897 3.687 5.40 

100 4.08 4.09 4.12 4.330 4.097 5.39 

110 4.4 4.41 4.45 4.763 4.420 7.20 

120 4.7 4.85 4.89 5.196 4.813 7.36 

130 5.2 5.25 5.23 5.629 5.227 7.15 

140 5.76 5.73 5.73 6.062 5.740 5.31 

150 6.02 6.03 6.03 6.495 6.027 7.21 

160 6.53 6.51 6.53 6.928 6.523 5.84 

170 6.92 7.02 7.12 7.361 7.020 4.63 

180 7.46 7.46 7.47 7.794 7.463 4.24 

190 7.85 7.9 7.83 8.227 7.860 4.46 

200 8.25 8.24 8.47 8.660 8.320 3.93 

210 8.58 8.62 8.61 9.093 8.603 5.39 

220 8.92 8.9 8.94 9.526 8.920 6.36 

230 9.28 9.35 9.29 9.959 9.307 6.55 

240 9.81 9.76 9.81 10.392 9.793 5.76 

250 10.23 10.21 10.25 10.825 10.230 5.50 

260 10.61 10.65 10.54 11.258 10.600 5.84 

270 11.23 11.36 11.36 11.691 11.317 3.20 

280 11.48 11.73 11.71 12.124 11.640 3.99 

290 12.03 11.92 12.03 12.557 11.993 4.49 

300 12.46 12.51 12.4 12.990 12.457 4.11 

310 12.92 12.91 12.86 13.423 12.897 3.92 

320 13.3 13.2 13.28 13.856 13.260 4.30 

330 13.65 13.86 13.72 14.289 13.743 3.82 

340 14.02 14 14.03 14.722 14.017 4.79 

350 14.4 14.64 14.64 15.155 14.560 3.93 

360 14.71 14.84 14.85 15.588 14.800 5.06 

370 15.21 15.4 15.34 16.021 15.317 4.40 

380 15.58 15.52 15.75 16.454 15.617 5.09 

390 16.01 16.01 16.03 16.887 16.017 5.15 

Using the data in Table 1 and Table 2 was realized graph of the relative 

dispersion of the friction losses amounts values for the two tensioning devices 

investigated (Fig. 3).


Relative loss in device [%] 

10,00% 

9,00% 

8,00% 

7,00% 

6,00% 

5,00% 

4,00% 

3,00% 

2,00% 

1,00% 

0,00% 

60 80 100 120 140 160 180 200 220 240 260 280 300 320 340 360 380 

Pressure [bar] 

Fig. 3 – Relative losses in tensioning devices with pulling force 16 tf, 

for 200 mm and 500 mm stroke. 


Losses by tensioning 

device with pulling force 

16 tF and stroke 200 mm 

Losses tensioning device 

with pulling force 16 tF 

and stroke 500 

From Table 1 and Table 2, as well as the dispersion graph of the errors 

(Fig. 3), we see that the relative losses of force by friction within the device 

varies between 3.16% and 8.72%, the maximum known in literature for 

tensioning devices with similar characteristics being 10% - 15%. 

The studied devices performances are comparable to those of other 

devices made by other manufacturers, which validates the use of this device in 

the stand for tension reinforcement of pre-stressed concrete. 

REFERENCES 

ChiriŃă C., Zetu D., Grama A., Afrăsinei M., Device for Tensioning of Strands of 

Prestressed Reinforced Concrete Structures. Bul. Inst. Polit. Iaşi, LV (LIX), 1, s. 

ConstrucŃii de maşini, 65-70 (2009). 

*** Studiere şi proiectare de modele experimentale - Dezvoltarea unui sistem de 

echipamente tehnologice hidraulice de forŃă, inovative, pentru modernizarea 

pretensionării armăturilor la structurile de beton precomprimat. Hydramold, 

Contract de finanŃare nr. 151, Tensrelax, Etapa I, 2008. 

*** http://www.suco-pressureswitches.com/druckueberwachung/drucktransmitter/ 

drucktransmitter.php 

*** http://www.aplisens.pl/en/produkty/pc.html 

*** http://www.hydramold.com/produse.php?cat=SistemeMasurareDigitala. 

CERCETĂRI PRIVIND PIERDERILE PRIN FRECARE ÎN 

DISPOZITIVELE DE TENSIONARE LA MERSUL ÎN SARCINĂ 

(Rezumat) 

Determinarea mărimii pierderilor prin frecare la mersul în sarcină al 

dispozitivelor de tensionare se realizează cu ajutorul calculului forŃei teoretice de 

încărcare în dispozitiv şi a forŃei reale de tragere dezvoltate de dispozitivul de 

tensionare.


Scopul acestei cercetări este de a determina mărimea pierderilor prin frecare la 

mersul în sarcină al dispozitivelor de tensionare utilizate şi compararea acestora cu 

valorile date în literatura de specialitate, privind dispozitivele de tensionare realizate de 

firmele de profil. 

Pentru determinarea pe cale teoretică a pierderilor de forŃă prin frecare la 

dispozitivele de tensionare monofilare, se va pleca de la relaŃia (1) care dă valoarea 

teoretică a forŃei de încărcare. 

Pentru determinarea pe cale experimentală a pierderilor de forŃă prin frecare s-au 

utilizat două dispozitive de tensionare monofilare realizate în cadrul Departamentului de 

Ingineria Sistemelor AcŃionate Hidraulic şi Pneumatic (DISAHP) din cadrul Catedrei de 

Maşini Unelte şi Scule al FacultăŃii de ConstrucŃii de Maşini şi Management Industial 

de la Universitatea Tehnică “Gheorghe Asachi” din Iaşi, care pot tensiona tendoane la 

forŃă de 16 tf, unul având cursa de lucru de 200mm şi al doilea de 500 mm; ambele 

dispozitive au suprafaŃa pistonului cilindrului care realizează tensionarea de: 43,3 cm 2 . 

Pentru fiecare treaptă de presiune s-au efectuat câte trei măsurători obŃinându-se 

seturi de câte trei valori ale forŃei de tragere. Aceste valori, împreună cu forŃa de încărcare 

teoretică corespunzătoare sunt prezentate în tabelele din lucrare. Tot aici, sunt prezentate 

numeric şi mediile aritmetice ale seturilor de câte 3 valori ale forŃei experimentale 

măsurate precum şi pierderile relative prin frecare în dispozitiv. 

Din tabele, precum şi din graficul de dispersie al erorilor, se observă că pierderea 

relativă de forŃa prin frecare din dispozitiv variază între limitele 3,16% şi 8,72%, 

valoarea maximă cunoscută în literatura de specialitate, pentru dispozitivele de 

tensionare cu caracteristici similare fiind de 10%...15%. 

Cele prezentate au condus la concluzia că dispozitivele studiate realizează 

performanŃe comparabile cu dispozitive realizate de alŃi producători, ceea ce validează 

utilizarea acestui dispozitiv în standul de tensionare a armăturilor din betonul 

precomprimat.


Publicat de 



SecŃia 


EXPERIMENTAL RESEARCH REGARDING THE DYNAMIC 

BEHAVIOR OF LINEAR HYDRAULIC SERVO-SYSTEMS 

BY 

CORNELIU CRISTESCU ∗ , PETRIN DRUMEA, 

CĂTĂLIN DUMITRESCU and DRAGOŞ ION GUŞĂ 



Hydraulics and Pneumatics Research Institute / 

INOE 2000 – IHP, Bucureşti 

Abstract. The paper presents the results of experimental research on the 

dynamic behavior of linear hydraulic motors and linear positioning servosystems 

carried out in INOE 2000-IHP, in the framework of the NUCLEU 

Program. The experimental investigations were conducted on an experimental 

stand with data acquisition and computer processing. The article presents some 

experimental graphic results obtained in the research, results that are of real 

scientific interest, which have a practical value through the using of them in the 

design activities of the fluid power components and equipments. 

Key words: linear hydraulic servo-systems, dynamic behavior, 

experimental research, test stand, data acquisition. 


In the structure of hydraulic drive systems, in addition to the equipment 

for generating, conditioning, control and distribution of the hydraulic energy, 

there are hydraulic operative elements (motors) that make transformation / 

conversion of hydraulic energy into mechanical energy and perform mechanical 

work required by the drive system (Velescu, 2003). Therefore, knowing the 

dynamic behavior of hydraulic operative elements, as operative parts of the 

working machines, is of particular interest to ensure performance of hydraulic 

drive equipment and systems. The main operative elements used within 

∗ Corresponding author: e-mail: cristescu.ihp@fluidas.ro

52 Corneliu Cristescu et al. 

hydraulic control and actuation systems are classified in two categories namely: 

linear hydraulic motors and rotary hydraulic motors (Marin & Marin, 1987). 

Linear hydraulic motors, which are the subject of research presented, are 

operative hydraulic elements that perform a linear motion at the working 

mechanism of equipment and machinery. These operative elements have as 

their characteristic the rectilinear motion and they are currently known as 

hydraulic cylinders, or hydraulic actuators (Fig. 1). 

Fig. 1 – REXROTH Linear Hydraulic Motor (MHL). 

Knowing the dynamic behavior of linear hydraulic motors (MHL), early 

since the design stage, involves conducting theoretical research and, also, 

experimental researches, which have a substantial role to confirming the 

theoretical results (Oprean et al., 1989). Based on experimental measurements, 

the experimental research can determine the actual performance of the dynamic 

behavior of the hydraulic operative elements. Experimental testing of the linear 

hydraulic motors, in order to investigate the factors that influence the dynamic 

behavior of linear hydraulic motors (MHL), there has been necessary to design 

and construct an experimental test stand, which allowed conducting 

experimental research in good condition. The test stand was developed inside of 

the Laboratory of Servo-Control Equipments from INOE 2000-IHP. 

In design and implementation of test stand for dynamic behavior linear 

hydraulic motors the intention was to maximally use the existing facilities, to 

which there have been added other equipment, instruments, components and 

devices specially purchased, to minimize the costs of such an action. In this 

respect, there was made extensive use of the facilities existing in the institute. 

2. Design of Dynamic Test Stand 

In design of the experimental test stand for dynamic behavior of linear 

hydraulic motors, there has been considered the equipment existing in the 

Laboratory of Servo-Control Equipments, in INOE 2000-IHP, taking as a base 

the structure of an existing stand for testing seals of hydraulic cylinders. It is 

equipped with a hydraulic cylinder with bilateral rod, which is actually a linear 

hydraulic motor, mounted vertically, which may be required to lift various 

weights. Supply of the linear hydraulic motor is made through a servo valve, 

from a hydraulic pressure group, with manually adjustable flow and possibility 

for flow measurement. The hydro-mechanical diagram of test stand for dynamic 

behavior of a linear hydraulic motor is shown in Fig. 2.


These experimental stands are a functional group of components, 

equipment, devices and proper instrumentation, aiming at allowing the 

experiments to be conducted, by which to highlight the factors influencing the 

dynamic behavior of linear hydraulic motors. 

Fig. 2 – Hydro-mechanical diagram of the stand for dynamic tests. 

The main components of the dynamic test stand, according to hydromechanical 

diagram shown in Fig. 2, are the next ones: linear hydraulic motor, 

MHL; stroke transducer, TS; mass /weight to be lifted, M/G; force transducer, 

TF, pressure transducers, TP1 and TP2; servo valve SV; hydro-pneumatic 

accumulators, AC1 and AC2; pressure unit RU, comprising typical parts, 

mounted on a tank Rz; data acquisition board, DAQ; computer PC. 

In principle, as shown in Fig. 2, the hydro-mechanical diagram of the test 

stand for dynamic behavior of linear hydraulic motor includes three major 

subassemblies, namely: the pressure unit, the hydro-mechanical system, which 

contains the linear hydraulic motor being tested, and the data acquisition system 

with computer, sensors and transducers (Calinoiu, 2009). The pressure unit, RU, 

provides adjustable oil flow and it has all the elements specific to usual pressure 

blocks. 

The hydraulic operative system with linear motion consists of linear 

hydraulic motor MHL and servo valve SV. The system is equipped with 

transducers needed to capture the evolution of parameters of interest: built-in 

stroke transducer TS, force transducer TF and pressure transducers PT1 and 

PT2. The hydraulic operative system with linear motion, consisting of linear


hydraulic motor and servo valve, is actually the subject tested, for the purpose 

of ascertaining the dynamic behavior of linear hydraulic operative elements. 

The data acquisition system consists mainly of the data acquisition board 

DAQ, computer PC and stroke transducer TS, force transducer TF and pressure 

transducers PT1 and PT2, and it works on the basis of data acquisition and 

processing software. Charging the operative system is performed by placing on 

the motor rod, over the force transducer, some parts with masses with different, 

but known, weights, M/G. To conduct experimental research on dynamic 

behavior of linear hydraulic elements MHL, there was chosen, as research 

object, an electro-hydro-mechanical servo system comprising a real linear 

hydraulic motor with bilateral rod, MHL, controlled by an electro hydraulic 

servo valve SV, manufactured by MOOG Company. 

3. Physical Implementation of Dynamic Test Stand 

After design, the test stand has been physically implemented by mounting 

its components according to the hydro-mechanical diagram in Fig. 2 and located 

in the Laboratory of Servo-Control Equipments, as it can be seen in Fig. 3. 

Fig. 3 – Dynamic test stand for linear hydraulic motors. 

Of particular interest is how are located and mounted the transducers 

required to absorb variations of the main dynamic parameters, on which 

depends, ultimately, the acquisition accuracy of evolution of parameters 

(Călinoiu, 2009). The control servo-valve of the hydraulic linear servo-system is 

shown in Fig. 4 and the Fig. 5 shows the pressure transducer mounted on the top 

of the linear hydraulic motor. In Fig. 6 is shown the flow transducer mounted on 

the pump discharge circuit and in Fig. 7 can be seen the pressure transducer 

with local display and the gauge mounted on the discharge circuit of the MHL. 

Fig. 8 shows the force transducer, which is located under the weights as 

the load charging items of MH and in Fig. 9 is shown the stroke transducer 

incorporated in the structure of the hydraulic motor MHL (Călinoiu, 2009).

Fig. 4 – Servo Valve. 

Fig. 6 – Flow transducer. 

Fig. 8 – Force transducer. 


4. Conducting Experiments 

Fig. 5 – Pressure transducer. 

Fig. 7 – Pressure transducer and gauge. 

Fig. 9 – Stroke transducer. 

Conducting experiments on the dynamic behavior of linear hydraulic 

motors was based on the testing software outlined at the beginning of the tests. 

For research of dynamic behavior, there must be known variations over 

time of dynamic parameters of interest namely: variation of stroke, speed and 

acceleration, pressure variation in the two circuits of the linear hydraulic motor 

and variation of inertia force. 

Therefore, after preparation and implementation of all technical 

conditions necessary for the operation of this experimental stand, it proceeds as 

follows: there are placed, successively, different known masses M, of weight G, 

on the rod of the linear hydraulic motor, as its load/charge; there is performed 

actuation of the linear hydraulic motor for one, two or three up and down, 

consecutive cycles; there is measured the variation of parameters of interest by 

acquiring and registering their evolution over time; finally, there are analyzed 

the values and graphical evolution of dynamic parameters of interest.


According with Testing Program, the tests were conducted for 2 steps of 

flow: 12.5 l/min and 27 l/min and for 3 inertial masses: 5kg, 12kg and 17kg. 

The hydraulic operative system with linear motion is an assembly of the 

linear hydraulic motor and its control servo valve. 

Technical data regarding the linear hydraulic motor MHL: type: bilateral 

rod cylinder, diameter of the cylinder: 105.4 mm, diameter of the rod: 70 mm, 

working stroke: 160 mm, working pressure: 250 bar. 

Technical data regarding the servo valve SV: type : MOOG, series: 760, 

for pressure: 70 bar, rated diameter: 6 mm, maximum flow: 40 l/min. 

5. Experimental Results 

After conducting experiments on the dynamic behavior of linear 

operative hydraulic components and systems, along time range, in accordance 

with the testing program, there were obtained a series of graphical and tabular 

results regarding variation in time of the main dynamic parameters: variation of 

inertia force on the rod of MHL, Fi=Fi(t); variation of stroke of the servo 

system rod, x=x(t), variation of velocity of the servo system rod, v=v(t); 

variation of pressure p1, at the input of MHL, p1=p1(t); variation of pressure 

p2, at the output of MHL, p2=p2(t); variation of temperature at the input of 

MHL, T = T(t). 

In Fig. 10, one can see the graphical variations of dynamic parameters 

over a complete cycle down-up-down, on stroke of 160 mm., for the flow step 

of 27 l/min (corresponding the theoretical velocity of 90 mm/s) and for the 

attached masse on rod of 17 kg. In the Table 1, are presented some selected 

numerical values of these dynamic parameters, only for a part of lifting stroke. 

Fig. 10 – Graphical variations of parameters of interest.


Table 1 

Selective numerical values for lifting stroke 

t[s] F[N] P1[bar] z[mm] v[mm/s] P2[bar] t[C 0 ] 

1.64 -18.46 35.36 98.464 74.40 31.93 29.44 

1.66 -12.53 35.32 99.856 72.80 31.82 29.38 

1.68 1.404 35.40 101.36 73.60 32.06 29.39 

1.70 -5.34 35.43 102.83 72.80 32.09 29.39 

1.72 -6.28 35.39 104.28 73.60 32.04 29.42 

Experimental results are in concordance with numerical calculus and the 

theoretical results obtained by mathematical modeling and computer simulation. 


1. The experimental research of the dynamic behavior of the linear servosystems 

was made by using the modern method of analysis and synthesis, 

including theoretical and experimental research. The experimental research has 

validated the constructive solutions for the experimental device and the 

experimental testing method. 

2. The diagram of force variation reveals low values on a complete piston 

stroke; this is due to the small masses attached on the cylinder rod. 

3. The variation of the piston stroke, on a complete cycle, is linear, with 

no sudden variations, due to using of servo-valve. 

4. Speed variation is constant, as there are no large jumps at start, also 

due the using of servo-valve. 

5. Pressure values, on the two piston working circuits, are very close, also 

due to the small masses attached on the cylinder rod. 

6. Analysis of experimental data enables identification of sensitive 

behavior parameters and, through changing their value; it can optimize the 

dynamic behavior of components and/or hydraulic acting systems (Fluid 

Power). 

Acknowledgements. Experimental research presented has been achieved in the 

framework of the institutional research program NUCLEU, funded by the National 

Authority for Scientific Research in Romania-ANCS. 

REFERENCES 

Velescu C., Aparate şi echipamente hidrostatice proporŃionale. Ed. Mirton, Timişoara, 

2003. 

Marin V., Marin Al. Sisteme hidraulice automate–Constructie reglare exploatare. 

Ed. Tehnică, Bucureşti, 1987.


Oprean, A., Ispas C., Ciobanu E., Dorin Al., Medar S., Olaru A, 

Prodan D., AcŃionări şi automatizări hidraulice, Modelare, simulare, 

încercare. Ed. Tehnică, Bucureşti, 1989. 

Călinoiu C., Senzori şi traductoare. Vol. I, Ed. Tehnică, Bucureşti, 2009. 

CERCETĂRI EXPERIMENTALE PRIVIND COMPORTAREA DINAMICĂ A 

SERVO-SISTEMELOR HIDRAULICE LINIARE 

(Rezumat) 

Articolul prezintă unele rezultate experimentale obŃinute în cadrul unei cercetări 

complexe, atât teoretice, cât şi experimentale, desfăşurate în institutul INOE 2000-IHP, 

pe un proiect de crecetare insituŃională din programul de cercetare NUCLEU al ANCS. 

În acest articol se prezintă, în mod special, standul de testare realizat, echiparea 

acestuia cu senzori şi traductoare şi cu sistem de achiziŃie date, procedura de 

experimentare, precum şi principalele marimi variate în timpul expeirmentărilor. 

Analiza datelor experimentale oferă posibilitatea identificării parametrilor cu 

sensibilitate comportamentală, iar prin modificarea lor, se poate optimiza comportarea 

dinamică a componentelor şi/sau sistemelor hidraulice de acŃionare (Fluid Power).


Publicat de 



SecŃia 


EFFECTIVE METHODS OF COST BREAKDOWN 

FOR DIRECT COSTS, TOOLS AND PRODUCT COSTS 

BY 

FLORINA-CRISTINA FILIP * 

Transilvania University, Braşov 

Department of Economic Engineering and Production Systems 

Received: May 6, 2011 

Accepted for publication: July 23, 2011 

Abstract. Various customers require a detailed breakdown of the product 

costs in the quotation and delivery phase, depending on the project costs involved. 

Some customers have different procedures and consequently use their own forms. 

In principle, detailed product cost illustrations for customers are subject to 

restrictions and may only be forwarded to customers once this has been agreed 

with the relevant departments. The content of the product cost breakdown is 

defined in the escalation stages and forwarded on. All the basic information for all 

stages about tooling adjustment costs (testing costs), tooling costs, design costs 

and type-specific measuring equipment costs is the same. The aim of this 

procedure is to ensure that the information presented to the customer is of uniform 

content. This is particularly important if the customer submits an enquiry for a 

product to different sales regions on an uncoordinated basis. The purpose of this 

paper is to regulate the handling and content of detailed product cost illustrations 

based on the cost value (standard quotation cost value) for customers. Through 

the use of standard forms, detailed product costs are only passed on to the 

customer on a restricted basis. Product cost data may only be handed over to the 

customer when all other possibilities have been rejected and it is unavoidable. 

Key words: administration cost, standard, price, quality, profit. 


A strong manager must understand how costs are captured and assigned 

to goods and services and in addition to alternative methods of costing, a good 

manager will need to understand different theories or concepts about costing. 

* e-mail: florinacristinafilip@yahoo.com

60 Florina Cristina Filip 

Costing is such an extensive part of the management accounting function that 

many people refer to management accountants as cost accountants. But, cost 

accounting is only a subset of managerial accounting applications. 

The price of any product is one of its most important attributes: a products 

price may differentiate it from competing products. The direct costs associated 

with the manufacture of a product are those costs that are directly associated 

with its manufacture. The most obvious direct costs are direct material and 

direct labour costs. Other direct costs always occur when manufacturing a 

product, such as the cost of the energy used to produce the product, the cost of 

supervising the staff producing the product, etc. Whether these direct costs are 

included in the product costs as directs or indirect depends on the circumstances 

in a company. 

For instance, if the electricity supply to a factory is monitored as its point 

of entry into the factory, it may not be possible to measure the energy used by a 

particular production line. The cost of energy used in manufacturing a product 

could then be apportioned to it according to some company rules, or procedures. 

However, if the cost of energy is not significant, it may be regarded as an 

indirect cost. Direct costs are sometimes referred to as variable costs because 

they vary with the level of production of a product. In theory, if we do not 

produce any products of a given type its direct costs are zero. 

A direct cost is a cost that is directly associated with changes in production 

volume. This usually restricts the definition of direct costs to direct materials 

and direct labour (and a strong case can be made for not using direct labour, 

since this costs tends to be present even when production volumes vary). For 

example, the materials used to create a product are a direct cost, whereas the 

machine used to convert the materials into a finished product is not a direct cost, 

because it is still going to be sitting on the factory floor, irrespective of any 

changes in production volume. Thus, direct costing assumes that fixed costs are 

period costs, and so should be recognized as expenses during the period when 

they occur. 

Exist many situations in which direct costing should not be used and in 

which it will yield incorrect information. Its single largest problem is that it 

completely ignores all indirect costs, which make up the bulk of all costs 

incurred by today’s companies. This is a real problem when dealing with longterm 

costing and pricing decisions, since direct costing will likely yield results 

that do not achieve long-term profitability. 

2. Cost Breakdown for Tool Costs/Special Direct Costs 

2.1. Level 1 – Cost Breakdown at Sub-Assembly Level 

At Level 1, detailed tool costs for component parts are not given as 

standard. Rather, at sub-assembly level, it is the part-specific tooling cost, 

design cost and tooling adjustment cost if necessary material cost that are 

forwarded to the customer. This information can be obtained from the local cost


computation and calculating system. If the customer is not satisfied with the 

standard answer, it is necessary to check whether specific information can 

potentially be given in accordance with Level 2. 

2.2. Level 2 – Cost Breakdown at Component Level 

If the inspection process produces a decision to respond in accordance 

with Level 2, a cost breakdown is performed at component level by standard. 

This decision may have a contractual foundation or may be the result of an 

additional request for information from the customer following Level 1. 

Information on the tooling cost, the design cost and tooling adjustment cost if 

necessary material cost can be obtained from the local cost computation and 

calculating system. 

2.3. Level 3 – Detailed Special Direct Costs 

If the customer does not accept the information from Level 2, with 

detailed special direct costs must be sent to the customer. The contents of this 

form are defined by the work planning department and the responsible key user 

in network of value management competence for checking and forwarding the 

form to sales. The detailed special direct costs include the parts designation, the 

tool type and the outlay in hours as well as the hourly rate for the tooling costs, 

design costs and the tooling adjustment costs for internal form. From this form 

is generated a PDF-Form which can be forwarded to the customer about sales 

department. The costs for measuring equipment can also be indicated. 

2.4. Level 4 – Transfer of Data into Customer’s Own Form 

If the customer is requesting a breakdown of the product costs and 

expects to receive this information in a sheet which has been supplied for this 

specific purpose (customer-specific form), it is still necessary to follow the 

procedure and corresponding steps described under point 2.3. The results must 

be entered into the detailed special direct costs form. If required, the 

presentation is made by the responsible key user in network of value 

management competence after consultation with the customer. This procedure is 

to ensure that the contents of the statements made to the customer are the same. 

This is especially important if the inquiry for a product from customers is made 

to various sales regions at the same time without any coordination. 

2.5. Level 5 - Photographs of Tools 

Supplementing tooling documentation by including photographs is only 

permitted in exceptional cases and after consultation with the responsible key 

user in network of value management competence if customer visits can be 

prevent.


Fig.1 – Inquiry processing with cost breakdowns for tooling.


To assist in monitoring productive efficiency and cost control, 

managerial accountants may develop standards. These standards represent 

benchmarks against which actual productive activity is compared. Importantly, 

standards can be developed for labour costs and efficiency, materials cost and 

utilization and more general assessments of the overall deployment of facilities 

and equipment (the overhead) . 

72% 

4% 

24% 

Labour 

Overhead 

Part Costs 

Fig. 2 – Typically product costs breakdown (Shipulski, 2009). 

To meet demand, a manager may prudently authorize significant overtime. 

This overtime may result in higher than expected wage rates and hours. As a 

result, a variance analysis could result in certain unfavourable variances. 

However, this added cost was incurred because of higher customer demand and 

was perhaps a good business decision. 

4. Conclusion 

1. Costing can occur under various methods and theories, and a manager 

must understand when and how these methods are best utilized to facilitate the 

decisions that must be made. 

2. Direct costs are attributable to the production of the goods sold by a 

company. This amount includes the cost of the materials used in creating the 

good along with the direct labour costs used to produce the good. It excludes 

indirect expenses such as distribution costs and sales force costs. Cost of goods 

sold appears on the income statement and can be deducted from revenue to 

calculate a company's gross margin. 

3. The possible mark-ups and mark-downs of the calculated special direct 

costs are passed on by sales to the creator of the cost breakdown due to factors 

relating market policy. 

4. Detailed product cost illustrations for customers are subject to 

restrictions and may only be forwarded to customers once this has been 

agreed with the relevant departments. 

Acknowledgements. This paper is supported by the Sectoral Operational 

Programme Human Resources Development (SOP HRD), financed from the European 

Social Fund and by the Romanian Government under the contract number 

POSDRU/88/1.5/S/59321. 

REFERENCES 

Briciu S., Căpuşneanu S., Effective Cost Analysis Tools Of The Activity-Based Costing 

(Abc) Method. Annales Universitatis Apulensis, Series Oeconomica, XII, 1, 25- 

35 (2010).


Drury C., Cost and Management Accounting. Thomson Learning, London, 2006. 

Quesada-Pineda H., The ABCs of Cost Allocation in the Wood Products Industry: 

Applications in the Furniture Industry. Virginia Polytechnic Institute and State 

University, College of Agriculture and Life Sciences, Communications and 

Marketing, 2010. 

Shipulski M., Product Design – The most Powerful (and Missing) Element of Lean. 

http://www.shipulski.com/2009/12/01/product-design-the-most-powerful-andmissing-element-of-lean, 

2009 

*** Introduction to Managerial Accounting. http://www.principlesofaccounting.com/ 

chapter%2017.htm 

*** Industrial Management Unit 6: Product Costing & Pricing. http://labspace.open. 

ac.uk/mod/resource/view.php?id=361239 

*** Direct Costing. http://www.accountingtools.com/direct-costing 

*** Cost of Goods Sold – COGS. http://www.investopedia.com/terms/c/cogs.asp 

METODE EFICIENTE DE REPARTIZARE A COSTURILOR DIRECTE, A 

COSTURILOR CU FORłA DE MUNCĂ ŞI A COSTURILOR DE REGIE 

(Rezumat) 

Pentru a produce bunuri, fiecare companie de fabricaŃie necesită alocarea de 

materii prime, forŃa de muncă şi cheltuieli de regie în scopul de a determina costurile 

finale de producŃie. Pentru cea mai mare parte a companiilor industriale, costurile de 

fabricaŃie sunt cuprinse într-un interval de 60-70% din preŃul de vânzare final al unui 

produs. Prin urmare, nevoia de sisteme eficiente de repartizare a costurilor este esenŃială 

pentru a controla costurile de producŃie. Pentru unii manageri, este important să ştie care 

parte din costuri sunt considerate fixe sau variabile. O companie care produce bunuri 

trebuie să aibe în vedere aspectele legate de achiziŃionarea şi prelucrarea materiilor 

prime necesare pentru realizarea unui produs finit, intrucat produsele rezultate în urma 

procesului de fabricaŃie şi costurile de producŃie, reprezintă însumarea dintre costurile 

cu materiile prime directe, costurile cu forŃa de muncă şi costurile de regie.


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ANALYSIS BY FINITE ELEMENT METHOD OF ASSEMBLY 

WEDGE GRIPS - MANTLE CORBEL 

BY 

ANDREI GRAMA *1 , CONSTANTIN CHIRIłĂ 1 , 

DUMITRU ZETU 1 and MIHAI AXINTE 2 



”Gheorghe Asachi” Tehnical University of Iaşi, 


2 Department of Materials Science 

Abstract. To achieve the required pulling and tensioning tendons in 

prestressed concrete structures it use locking systems. Locking systems must be 

made of materials with appropriate strength so as not to destroy their active 

surface due the pulling forces having large and very large values. This is 

achieved by using alloy steel, heat treated on active surfaces and by 

dimensioning so as to provide necessary forces and unlocking of wedge grips at 

the end of tension cycle. 

Key words: wedge grips, clamping device, tensioning device, 

pretensioning, finite elements analysis, virtual model. 


Locking systems are devices with jams/friction wedges used to secure the 

ends of the thus wire and keep them tensioned after removing of the tensioning 

device, up to the strength of concrete and blocking tensioning device on the 

active contact surface between tendons and the wedge grips. 

Locking systems must be made of materials with appropriate strength so 

as they do not make damages due the pulling forces which have large and very 

large values and in abutment the strained tendons to be kept locking up after 

hardening of the cast concrete and relief. This is done on the one hand by using 

alloy steels, heat treated to superficial hardness of 60…65 HRC on the other 

hand, constructionmust be sized so as to provide necessary friction force of 

* Corresponding author: e-mail: andreiasi79@yahoo.com

66 Andrei Grama et al. 

blocking systems (at tensioning) and unlocking of wedge grips at the end of 

tensioning. 

Constructive solution of locking system of tensioning device is shown in 

Fig. 1. 

Fig. 1 – Wedge grips in the device for tensioning of tendon (ChiriŃă et al. 2009). 

Theoretical and experimental researches conducted on locking systems 

are designed to establish the type of material and their optimal dimensions and 

shape. To this purpose theoretical research will be made by finite element 

analysis. 

Finite element analysis modeling involves the following steps: establish 

constructive solution for blocking systems and their components; establishing of 

virtual model of locking elements; establishing of surface elements; setting 

parameters considered in the analysis; data processing; interpretation of 

numerical analysis results. 

For analysis by this method was used CATIA (Computer Aided Three 

Dimensional Interactive Application) that is a suite of multiplatform 

commercial software CAD/CAM/CAE French company Dassault Systemes 

developed. (Ghionea, 2007) 

In this paper, is presented finite element analysis for the assembly wedge 

grips – mantle corbel. 

In Fig. 2 is shown how is the decomposition of surface contact forces 

between wedge grips – mantle corbel. 

Fk + Fbl 

Fig. 2 – Mode decomposition of the contact force F on the surface of a wedge grips – 

mantle corbel (Axinte&Grama, 2009): 1 − mantle corbel, 2 − wedge grips F − force 

developed by hydraulic cylinder piston drive, Fbl − force who remain in reinforcement 

after it was blocking by tensioning in the abutment, FK − tensioning force. 

F


2. Establishing Virtual Model, Data Processing and Research Results 

Assemble contact elements with pretensions pull operation strand of 

tensioning device of 16 tf in the virtual model comprises: set of 3 wedge grips, 

mantle corbel, wire 

To obtain the contact element assembly for pull operation, will be done 

the following steps: 

i) choose of assembly components; 

ii) establish their constraints. 

It is considered as a fixed constraint and constraint wedge grips contact 

surfaces + mantle corbel (2x conical surfaces wedge grips x 1 = 2 surface) and 

surface contact constraint between the wedge grips and wire (3x surface wedge 

grips on wire x 1 surfaces wedge grips = 3 surfaces) 

Fig. 3 shows the studying assemblies, where all the stages of constraint 

on the operation of pulling contact with pretensions of wire are: Start → 

Mechanical Design → Assembly Design → Contact Constrain, existing 

elements tensioning device 16 tf. 

The model is subject to optimization, where in addition to the set of 

wedge grips teeth, and the effects of contact between the wedge grips and 

strand, only for wedge grips made of OLC 45 and mantle corbel made of 

40Cr10 (Table 1). 

Fig. 3 – Virtual model of studied assembly. 

Table 1 

Materials used for assembly and mechanical properties 

of elements introduced in the modeling 

Element assembly Material Mechanical resistance 

Wedge grips OLC 45 STAS 880/88 4.8 x 10 8 N/m 2 

Mantle corbel 40Cr10 STAS 791-88 5.6 x 10 8 N/m 2 

Wire Wire 7 4 STAS 15.1 x 10 8 N/m 2


In this case, two assemblies were used, the difference between them 

being that the contact area between teeth and tendon pulling wedge grips 

provides a fillet to the wedge grips teeth was observed difference between the 

two assemblies created. 

To generate finite element model, launches CATIA Analysis & 

Simulation package and select Structural Analysis Module Generation. Then 

select option Static Analysis of the New Analysis Case window, involving static 

analysis of the structure in terms of constraints and loads independent of time. 

The tree structure’s specified requirements are observed the following undertree 

specification structures. 

i) Links to identify the path to save files with the final results, to identify 

the path to save intermediate results files, and that to return to proper 

specifications solid model for analysis (Product1. CATProduct). 

ii) Finite element model specifications: Elements and Nodes, Properties.1 

and the State Case. The activation by double pressing the left mouse button, 

specification OCTREE Thetraedron Mesh.1: part1 or green symbol associated 

with the finite element type, automatically set type tetrahedron, OCTREE. 

Tetrahedron window appears in finite element order is selected (linear), and 

modified finite element model of global dimensions and maximum deviation 

from the real model for each item, according to Table 2. 

Table 2 

Finite element dimensions and the maximum deviation from the real model 

Element assembly Size Absolute sag 

Wedge grips 15 2 

Mantle corbel 20 5 

Wire 15 2 

A finer mesh was made on the surface of interest, which is the contact 

area between teeth and tendon (see Table 3). 

Local Mesh Size was achieved by Local Mesh Size icon Model Manager 

Bar and specifies the size of mesh element faces choosing areas of interest one 

by one. 

Table 3 

Finite element dimension and the maximum deviation from the real model 

Element assembly Size 

Wedge grips 2 

Wire 2 

Static Case with specifications Restraints.1, Loads.1, Static Case 

Solution.1 and Sensor.1 indicating sets of constraints, loads, if the synthesis 

solution and analysis results.


a) Introduction of displacement restrictions.Radial direction and axial 

displacement of the mantle corbel is locking. First we introduced the clamp tool 

-“clamp”, restricting displacement mantle corbel by radial direction. 

Axial displacement direction was constrained wedge grips on conical 

surfaces with locking mantle corbel by blocking constraint surface. 

The restriction on displacement in the axial direction of strand was 

constrained on the surfaces of contact between it and pulling teeth wedge grips 

in 3 places by: 

b) Loads modeling. It applies evenly distributed force between the three 

teeth in contact at the end of the wedge grip and strand. Force pulling is 16 tf 

and by every one of the teeth of wedge grips is exerted one force FK / no. wedge 

grips / no. teeth = 16/03/55 tf. In field of forces that is exerted on the virtual 

model we were introduce the module force 97N on negative direction of Y axis 

(perpendicular axis to the surface of strand). 

Total, the force was applied by a number of 3 areas represent 1 tooth. 

Data processing of studied model occur as follows: 

It launches the solver calculation of this program. After application of 

forces and achieving virtual model calculations the results are post processing 

using the following tools. 

1. Viewing the status of deformed: twisted images are all used to view the 

finite element mesh in the deformed configuration of the system, as a result of 

environmental action; 

2. Field of viewing equivalent stress (Von Mises) superimposed on the 

undistorted structure; 

3. Viewing field of displacements. 

Fig. 4 – Field of viewing equivalent stress (Von Mises) 

that appear in all deformed if we have no fillet on the top tooth. 

The tool Image Extrema→Extrema Creation can highlight maximum 

stress occurrence place, and if they exceed allowable values we can take locally 

or globally to optimize virtual model. In Fig. 4 may view the tensions that arise 

as we have no jurisdiction on the top of the tooth.


From we see the analysis that only ends peaks occur which leads to the 

conclusion that occurs in this area of tension concentrators. This requires 

modification of the rib end zone geometry virtual model. In Fig. 5, we can see 

the equivalent stress (Von Mises) at the tooth wedge grip (all over) where we 

have fillet on the top of the tooth. 

After changing the virtual model by applying a connected fillet (0.4 mm) 

to the strand of tooth displacement and stress field analysis shows that the 

element with maximum stress (S = 6.56628 008 N/m 2 ) is also on the extremities 

ribs. However this new value is acceptable and it is under the admissible value. 

In Fig. 6 we can see the action of equivalent stress (Von Mises) in tendon and 

wedge grips. 

Fig. 5 – Field of view equivalent stress (Von Mises) at the wedge grip 

teeth if we do not have range on top of the tooth. 

Fig. 6 – Field of view equivalent stress (Von Mises) that appear together 

in the tendon and wedge grip, the tooth has a fillet of 0.4 mm on top. 

Comparing the model shown in Fig. 6 we can see the analysis that we do 

not have peaks that occur on teeth or on their ends which leads to the conclusion 

that not appear in that area the tension concentrators. In this way we had 

optimize the geometric shape of tooth wedge grip. In Fig. 7 we can see 

equivalent tensions (Von Mises) at the tooth wedge grip (all over) if we had 

fillet on the tooth.


Fig. 7 – Field of view equivalent stress (Von Mises) at the wedge grip 

teeth when the tooth has a radius of 0.4 mm on top. 


1. In material terms that mean the manufactured wedge grips can be 

observed and the best material suited for them is 13CrNi35 STAS 791 -66 

because strength yield is much closer to the maximum stress of wedge grip. In 

this way we had optimize depending on the quality of the best material suited 

for material to achieve wedge grips. 

2. Using the model of finite element analysis, was optimized the 

geometric shape at the end of tooth contact area between the wedge grips and 

tendon. If we apply a connection fillet of 0,4 mm based by tooth wedge grip, the 

tensions in contact between tendon and wedge grips are much small (Fig. 6 and 

Fig. 7). Also, reducing the dimension of wedge grips length from 90mm to 

70mm (see Fig. 8), because the length of 90mm wedge grip there is locking on 

the strand. (Clamping force is greater than the force required for pooling) (see 

Fig. 8). 

Fig. 8 – Model wedge grips optimized.


REFERENCES 

*** Durabilitatea elementelor şi structurilor de beton precomprimat. INCERC 

Bucureşti – Filiala Cluj-Napoca. 

Axinte M., Grama A., Finite Elements Analysis on Forces And Tensions That Act on the 

Wedge Grips for Drawing Single Wire. Bul. Inst. Polit. Iaşi, LV(LIX), 2, s. 

ŞtiinŃa Materialelor, 55-68 (2009). 

ChiriŃă C., Zetu D., Grama A., Afrăsinei M., Device for Tensioning of Stranda of 

Prestressed Reinforced Concrete Structures. Bul. Inst. Polit. Iaşi, LV(LIX), 1, s. 

ConstrucŃii de maşini, 65-71 (2009). 

Ghionea I., , Proiectare asistată în Catia V5, elemente teoretice şi aplicaŃii. Ed. Bren, 


Grama A., Zetu D., ChiriŃă C., Mathematical Model of Forces That Act on the Wedge 

Grips for Drawing Single Wire. Bul Inst. Polit. Iaşi, LVI(LX), 2b, s. ConstrucŃii 

de maşini, 125-134 (2010). 

Manea M., Etanşări fără contact – AplicaŃii. Bacău, 2008. 

ANALIZA PRIN METODA ELEMENTULUI FINIT A ANSAMBLULUI 

BACURI DE TRAGERE – BUCŞĂ PORT-BACURI 

(Rezumat) 

Pentru a realiza tragerea şi tensionarea tendoanelor necesare structurilor din 

beton precomprimat se utilizează sisteme de blocare. Sistemele de blocare trebuie să fie 

realizate din materiale cu duritate corespunzătoare astfel încât să nu se distrugă partea 

lor activă datorită forŃelor de tragere mari şi foarte mari. Acest lucru se realizează prin 

utilizarea unor oŃeluri aliate, tratate termic pe suprafeŃele active ale acestora şi prin 

dimensionarea în aşa fel încât să asigure forŃele necesare în blocaje şi desfacerea 

bacurilor de tragere la sfârşitul ciclului de tensionare. 

Cercetările teoretice şi experimentale efectuate asupra sistemelor de blocare au ca 

scop stabilirea tipului de material precum şi a dimensiunilor şi formei optime ale 

acestora. În acest scop vor fi efectuate cercetări teoretice prin analiza cu element finit. 

Analiza prin modelare cu element finit presupune parcurgerea următoarelor 

etape: stabilirea soluŃiei constructive pentru sistemele de blocare şi elementele acestora, 

stabilirea modelului virtual al elementelor de blocare, stabilirea elementelor de 

suprafaŃă, stabilirea parametrilor avuŃi în vedere în efectuarea analizei, procesarea 

datelor, interpretarea rezultatelor obŃinute din analiza numerică. 

Pentru efectuarea analizei prin această metodă s-a folosit CATIA (Computer 

Aided Three Dimensional Interactive Application) care este o suită software comercială 

multiplatformă CAD/CAM/CAE dezvoltată de compania franceză Dassault Systèmes. 

Se prezintă analiza prin metoda elementului finit pentru ansamblul bacuri de 

blocare – bucşă portbacuri. 

Pe modelul de bacuri de blocare s-a evidenŃiat modul de descompunere al 

forŃelor pe suprafaŃa de contact bucşă – bacuri de blocare, acesta fiind prezentată în 

lucrare. Asamblul elementelor de contact la opera ia de tragere prin pretensionare a 

toronului din dispozitivul de tensionare de 16 tF în cadrul modelului virtual cuprinde: 

set de 3 bacuri de tragere, bucşa port bacuri de blocare, toron.


Modelul supus studiului are în vedere optimizarea, pe lângă forma dinŃilor setului 

de bacuri de tragere, şi efectele contactului dintre zona activă a bacurilor de tragere şi 

toron, numai pentru bacurile de tragere confecŃionate din OLC 45 şi bucşă 40Cr10. 

În urma analizei prin metoda elementului finit, constatăm umătoarele concluzii: 

Din punct de vedere al materialului din care se confecŃionează bacurile de 

tragere, se poate observa că materialul cel mai potrivit pentru acestea este 13CrNi35 

STAS 791 -66 datorită limitei la curgere mult mai apropiată de tensiunea maxima din 

dintele bacului de tragere. În acest mod am optimizat în funcŃie de calitatea materialului 

cea mai potrivită marcă de material pentru realizarea bacurilor de tragere. 

Prin cele două modele de analiză cu element finit, am optimizat geometric forma 

dintelui la capătul zonei de contact dintre bacurile de tragere şi tendon. S-a observat că 

aplicând o teşire la extremităŃile dinŃilor (capetelor) de 1x45 0 , tensiunea maximă pe 

zona teşită scade sub valoarea efortului normal maxim. S-a observat că dacă aplicăm o 

rază de racordare de 0,4 mm pe baza dintelui bacului de blocare, tensiunile la nivelul 

contactului dintre tendon şi bacul de blocare sunt mult mai mici.


Publicat de 



SecŃia 


ABOUT THE TUNABLE PROPERTIES OF THE YIG FILMS 

WITH APPLICATIONS IN MICROWAVE DEVICES 

MANUFACTURING 

BY 

DANIELA IONESCU *1 , ION BOGDAN 1 and GABRIELA APREOTESEI 2 

Received: June 27, 2011 



1 Department of Telecommunications 

2 Department of Physics 

Abstract. The tunable properties of the yttrium iron garnet (YIG) films 

with applications at microwave devices were investigated in this paper. For 

improving the manufactured systems, two directions are to be followed in 

research regarding the YIG special properties: the frequency tuning capabilities 

and the tunable negative permeability of the YIG structures. Study was 

performed by structural simulation methods with HFSS 13.0 by Ansoft. A dc 

applied magnetic field (3 ÷ 6.5 kOe) was set in order to obtain the ferromagnetic 

resonance modification of the material. By simulations, tunability was achieved 

in the 12 - 20 GHz frequency range. The tunable electromagnetic properties of 

the YIG films were also studied in ac electromagnetic fields (12 - 30 GHz), on 

the high-frequency side of the resonance. The field-driven response of the 

labyrinthine magnetic domain walls is harmonic, determining a similar evolution 

of the magnetization. For a manufactured metamaterial including the YIG film 

and a metallic grid array, simulated with HFSS, the negative magnetic 

permeability was determined in a controlled frequency domain. This domain was 

identified as 13 - 20 GHz, depending on the metamaterial structure and 

controlled by the dc/ac driving field. The study is dedicated to the performance 

optimization of the YIG devices in microwave range. 

Key words: advanced modeling techniques, simulation in manufacturing 

systems, optimal design. 

1 Corresponding author: e-mail: danaity@yahoo.com

76 Daniela Ionescu et al. 


The tunable properties of the yttrium iron garnet (YIG) films with 

applications at microwave devices were investigated in this paper. The YIG 

material having a high Q factor in microwave frequencies is used at the YIGtuned 

oscillators, drivers, tunable YIG filter for wideband, multi-function YIG 

components. The elastic domain walls of YIG have opened the way for specific 

applications like the magnetic bubble domain-type memories, field modulating 

structures and metamaterials. 

For improving the YIG manufactured systems, two directions are to be 

followed in the research regarding the YIG special properties: the frequency 

tuning capabilities, the tunable negative permeability of the YIG structures. 

2. Frequency Tuning Capabilities 

2.1. The YIG Films Structure Details 

Our study was performed by structural simulation methods, using the 

HFSS 13.0 program (Ansoft Technologies). The polycrystalline films of YIG 

were reproduced at microscopic level, considering the garnet structure, the 

particularities of the unit cell and the intimate interactions between the 

microcomponents. 

Fig. 1 – The YIG la3d(Oh 10 ) space group (Steven Dutch, Natural and Applied Sciences) 

(left). Intensity of magnetization map, corresponding to the magnetic domain structure 

in the vicinity of a YIG film surface (Del Mar Photonics nano-imaging gallery) (right). 

The iron garnet (YIG, Y3Fe2(FeO4)3) is known like a ferrimagnet with a 

sharp ferrimagnetic resonance. Its cubic structure (la3d(Oh 10 ) space group, Fig. 

1) represent a three-dimensional framework: groups of independent, distorted 

FeO4 tetrahedrons linked, by sharing corners, to distorted FeO6 octahedrons. 

The Fe 3+ ions in the two coordination sites exhibit different spins and determine


the magnetic behavior of the material with a unit cell containing 128 ions and a 

lattice constant of 12.376 Å. The YIG present a high resistivity of 10 16 Ω/m and 

an electric permittivity εeff around 15 (Buschow, 2005). 

If an external magnetic field is applied parallel to the easy magnetization 

axis, in the vicinity of the surface of a YIG film cut perpendicular to the easy 

axis, the magnetic domain structure reorganizes at field variation. The domain 

configuration changes due to the elasticity of the domain walls, which is 

analogous to a liquid surface tension (Fig. 1). Domain wall displacements and 

domain rotations occur, determining a continuously tunable negative 

permeability at film level. 

2.2. Resonance Shifting for the Metamaterial Samples 

The ferromagnetic resonance frequency (FMR) for pure YIG can be 

estimated with Kittel's formula (Marques et al., 2008): 

( )( ) 

f ≅ γ H + H H + H + M 

(1) 

0 [ SI ] a a S . 

For a single crystal of YIG, the saturation magnetization, Ms = 0.18 T, the 

anisotropy field, Ha, is of ca. 70 Oe and the gyromagnetic ratio γ = 2.8 

GHz·kOe −1 (Buschow, 2005). The external bias magnetic field H was taken in 

range of 3…6.5 kOe. The almost linear FMR dependence on the bias field was 

confirmed by the simulation results. The polycrystalline structure of the YIG 

film has to be considered. 

Previous studies indicate that the FMR is dependent on: 

i) the static magnetic polarizing field; FMR varies almost linear with H0 

(He et al., 2007), (Zhao et al., 2007), (Kang et al., 2008); 

ii) the grain size (in the range of 12…20 nm); FMR is inversely 

dependent on grain size on a log-log scale (Gokarn et al., 1982); 

iii) the orientation of the magnetic field with respect to the crystalline 

axes (influences strongly the shapes of ferromagnetic resonance spectra) (Jalali 

et al., 2002); 

iv) temperature; the dependence is non-linear (Capolino, 2009); 

v) the impurities (nine substitution elements can influence the YIG 

structure: Si, Ti, Cr, Co, Ha, Sn, Ge, Mg, Ta) (Chiang et al., 2002); 

vi) a periodic layer structure (PLS) in a YIG film (Buschow, 2005), 

(Kanivets&Sarnatsky, 2010). 

The tunable electromagnetic properties of the YIG films were studied in 

ac electromagnetic fields (12 - 30 GHz), on the high-frequency side of the 

ferromagnetic resonance. Similar results were reported (DeFeo et al., 2006), 

based on the local ac dynamics of the labyrinthine magnetic domain phase in a 

YIG sample. The field-driven response of the domain walls is harmonic, 

determining a similar magnetization evolution, in the range of 0.18…0.25 T for 

the bulk YIG, respectively of 0.156…0.166 T for the YIG films (Capolino,


2009), (Gokarn et al., 1982). A similar pattern was considered in our 

simulation scenery. 

Our study was focused on the composed metamaterial structure 

consisting of a YIG film and a copper wires array, excited with a microwave 

field. The polarized YIG ensures the negative magnetic permeability, while the 

metallic array ensures the negative electric permittivity for specific frequency 

domains, depending on the polarizing field and structure parameters. 

The metamaterial samples were simulated inside a rectangular waveguide with 

the cross-section of 22.86 × 10.16 mm, excited on a TE10 mode (Fig. 2). In the 

metamaterial samples we have a YIG film of different thicknesses (typical 400 

µm) and copper wires of 0.3 mm wide and 1 mm spacing between their centers. 

A gadolinium gallium garnet (GGG, Gd3Ga5O12) substrate (paramagnetic) of 

1 mm thick was chosen because its lattice constant and thermal expansion 

coefficients match with those of YIG. 

Fig. 2 – The rectangular waveguide with the metamaterial sample. 

A dc magnetic field (3…6.5 kOe) was set in order to obtain the 

modification of the metamaterial resonance. The resonance frequency 

dependence on the bias field is illustrated in Fig. 3. The red curve corresponds 

to the simulation results and the blue dot curve was obtained by theoretical 

calculation of resonance using the Kittel formula. 

By simulations, the tunability of the resonance frequency was achieved 

in the 12…20 GHz range. The theoretical resonance frequency increases almost 

linear with the bias field, while the simulation results indicate a polynomial 

evolution with respect to H0. Considering the mechanism which generates the 

resonance, a simple linear dependence is no more justified. Punctual 

experimental results (He et al., 2007), (Marques et al., 2008), obtained for 

metamaterial structures based on YIG films, are placed in the vicinity of the 

polynomial curve and confirm the simulation results.


Fig. 3 – Dependence of the resonance frequency on the bias field, for the composed 

metamaterial sample consisting of the YIG film and the metallic grid array. 

For bias fields of ca. 3…4.8 kOe the frequency tuning capabilities of 

YIG. are to be considered for applications. For higher bias field, the slope of the 

fr versus H curve is lower and the energy consumed for tuning is higher. 

3. Tunable Negative Permeability 

3.1. Components of the Permeability Tensor 

For the considered metamaterial structure, where the YIG film is 

completed with a metallic periodic array, the local ac susceptibility variations 

determine a global negative magnetic permeability in a controlled frequency 

range. 

For the YIG film, the complex magnetic permeability can be written as 

(Marques et al., 2008) 

with 

fr 

10 9 fr sim [GHz] 

20 

19 

18 

17 

16 

15 

14 

13 

12 

fr_K 

11 

10 9 fr Kittel [GHz] 

fr theor [GHz] 

10 

9 

8 

7 

6 

3 3.5 4 4.5 5 5.5 6 6.5 

ɵ 

⎛ µ r −iµ 

κ 0⎞ 

⎜ ⎟ 

µ = µ 0 ⎜ 

iµ κ µ r 0 

⎟ 

, (2) 

⎜ 0 0 1⎟ 

⎝ ⎠ 

ωm ( ω0 − 2πiα1 f ) 

( − 2 i ) − 4 

µ r = 1+ 

ω π α f π f 

µ 

κ = 

0 1 

2 2 2 

( ) 2 2 2 

ω − 2πiα f − 4π 

f 

0 1 

H [kOe] H 

10 3 

2πω 

f 

m 

, (3) 

, (4)


where α1 is the damping coefficient, ω0 = γH0 is the resonance angular 

frequency, γ = the gyromagnetic ratio; H0 is the sum of the external magnetic 

field H and shape anisotropy field Ha along the easy magnetization axis, 

ωm = γMs is the characteristic frequency, Ms = the saturation magnetization, 

(Ms = 0.18 T, typical for single-crystal of YIG). 

The HFSS program was used to obtain the S-parameters corresponding to 

the field propagation through the metamaterial samples placed inside the 

rectangular waveguide. The real, respectively imaginary parts of the relative 

permeability were calculated, for different values of the magnetic field applied 

for magnetization. The following expression can be written for the permeability 

tensor components (Marques et al., 2008) 

B 

H 

≈ 

≈ 

µ r −iµ 

κ 0 

µ −iκ 

= = µ ˆ = µ 0 iµ κ µ r 0 , (5) 

iκ 

µ 

0 0 1 

µ = µ '+ i µ " ; κ = κ '+ i κ" 

. (6) 

The electromagnetic parameters were computed and the continuously 

tunable negative permeability was illustrated on graphs. 

3.2. Parametrical Curves for Negative Permeability 

Parametrical curves have been obtained, which indicate the frequency 

domain of negative values for the effective permeability of the incorporated 

YIG film. This domain was identified 

in the range of 12…30 GHz, 

depending on the metamaterial 

structure and controlled by the dc/ac 

driving field (see near). 

The first set of curves is given 

in Figs. 4 and 5 and has been obtained 

for different value of the magnetic 

polarizing field H. The permeability 

curves have been represented in the 

12…30 GHz domain, on the high-frequency side of the resonance, where the 

composed metamaterial is used for applications. 

The resonance shift with H and the shape of the curves for the 

permeability tensor components also varies. Deeper and wider permeability 

curves in the resonance vicinity are associated with lower bias field, 

corresponding also to a wider negative permeability domain. For intense the 

bias fields (up to 6 kOe), this domain is narrower, but placed at higher 

frequencies. 

The parametrical curves corresponding to the frequency evolution of the 

negative permeability tensor components for different YIG film thicknesses are 

0 

fres 

µ" 

12...30 GHz 

µ' 

f 

represented 

area


given in Fig. 6. The polarizing field was set at 5.5 kOe and the resonance 

frequency remains practically the same due to the fact that the film thickness 

does not influence the interaction mechanism between substance and the driven 

field. Instead of this, the film magnetization decreases for thicker films, the 

amount of energy necessary to order all deep domain moments being higher 

than for thinner films. In the same time, the domain of negative permeability 

values decreases significantly when the film thickness is grown from 100…850 

µm. 

Fig. 4 – Real parts of the magnetic permeability tensor components, µ', respectively κ', 

obtained by simulation for the YIG based metamaterial. The YIG film thickness is of 

400 µm. The resonance tunability was achieved in the 12…20 GHz range. 

Fig. 5 – Imaginary parts of the magnetic permeability tensor components, µ", 

respectively κ", obtained by simulation for the YIG based metamaterial. 

(The YIG film thickness is of 400 µm.) 

One remarks that the magnetic permeability curves are controlled by the 

YIG film thickness. If the film is thicker than a specific value, the real part of 

the permeability, µ', becomes positive over the whole frequency range (the 

dielectric effect of the ferrite overwhelms its magnetic properties). Our 

simulation results indicate that this phenomenon takes place in the YIG case for 

a film thickness of 860 µm, which represents the threshold value for existence 

of a propagation passband where both magnetic permeability and electric 

permittivity occur.


Fig. 6 – Real parts of the magnetic permeability tensor components, µ', respectively 

κ', obtained by simulation for the YIG based metamaterial, for a polarization field of 5.5 

kOe and different YIG film thicknesses. The frequency domain corresponding to the 

negative values of interest can be selected from graph for each curve. 


1. This study is dedicated to the performance optimization of the 

metamaterial structures based on YIG films, in microwave range. We have 

focused on the tunable properties of the YIG films. The garnet is a ferrimagnet 

with a sharp ferrimagnetic resonance, for which a bias field determines domains 

configuration changes due to the elasticity of the domain walls. As a result, 

resonance frequency shifts and a continuously tunable negative permeability 

occurs at film level. The considered metamaterial structure where the YIG 

works consists of a YIG film and a copper wires array, excited with a 

microwave field (12…30 GHz). 

2. By simulations, the tunability of the resonance frequency was 

achieved in the 12…20 GHz range, presenting a polynomial evolution with 

respect to the bias field, punctually confirmed by the reported experimental 

results. 

3. The optimum tuning range for bias field is of ca. 3…4.8 kOe, for 

which the material responds better to the polarization control. 

4. The negative values for the effective permeability of the incorporated 

YIG film occur in the frequency range of 13…20 GHz, depending on the 

metamaterial structure and controlled by the dc/ac driving field. 

5. The shape of the curves for the permeability tensor components 

depends on the bias field; a wider negative permeability domain corresponding 

to a lower bias field, but this domain is placed at higher frequencies for more 

intense bias fields (up to 6 kOe). 

6. The domain of negative permeability values decreases significantly 

when the film thickness is grown from 100 - 850 µm.


7. For a threshold value of the YIG film thickness of 860 µm inside the 

metamaterial, the propagation passband where both magnetic permeability and 

electric permittivity occur vanishes. 

Acknowledgements. This paper was supported by the project PERFORM-ERA 

"Postdoctoral Performance for Integration in the European Research Area" (ID-57649), 

financed by the European Social Fund and the Romanian Government. 

REFERENCES 

Buschow K.H.J., Concise Encyclopedia of Magnetic & Superconducting Materials. 

Elsevier, 2005. 

Capolino F., Applications of Metamaterials. Taylor and Francis, 2009. 

Chiang W.-C., Chern M.Y., Lin J.G., Huang C. Y., FMR Studies of 

Y3Fe5O12/Gd3Ga5O12 (YIG/GGG) Superlattices and YIG Thin Films. Journal of 

Magnetism and Magnetic Materials, 239, 1-3, 332-334 (2002). 

Gokarn S.G., Palkar V. R., Multani M.S., (1982), Sub-micron YIG Microstructure and 

Ferromagnetic Resonance Linewidth. Materials Research Bulletin, 17, 8, 957-962. 

He Y., He P., Dae Yoon S., Parimic P. V., Rachford F. J., Harris V. G., Vittoria C., 

Tunable NIM Using Yttrium Iron Garnet. Journal of Magnetism and Magnetic 

Materials, 313, 1, 187–191 (2007). 

Jalali A. A., Kahl S., Denysenkov V., Grishin A. M., Vanishing of Cubic 

Magnetocrystalline Anisotropy in Critical Angles Effect: Ferromagnetic 

Resonance Spectra. Phys. Rev. B, 66, 104419 (2002). 

Kang L., Zhao Q., Zhao H., Zhou J., Magnetically Tunable Negative Permeability 

Metamaterial Composed by Split Ring Resonators and Ferrite Rods. Optics 

Express, 16, 12, 8825-8834 (2008). 

Kanivets A. A., Sarnatsky V.M., Generation of High-Frequency Ultrasonic Oscillations 

by Thin Films of Yttrium Iron Garnet. XXII Session of the Russian Acoustical 

Society, Session of the Scientific Council of Russian Academy of Science on 

Acoustics, Moscow, 2010. 

Marqués R., Martin F., Sorolla M., Metamaterials with Negative Parameters, Theory, 

Design, and Microwave Applications. Ed. Wiley & Sons, 2008. 

Zhao H., Zhou J., Zhao Q., Li B., Kang L., Bai Y., Magnetotunable Left-handed 

Material Consisting of Yttrium Iron Garnet Film and Metallic Wires. Appl. Phys. 

Lett., 91, 13, 131107 (2007). 

ASUPRA PROPRIETĂłILOR CONTROLABILE ALE STRATURILOR 

SUBłIRI DE YIG CU APLICAłII LA FABRICAREA 

DISPOZITIVELOR DE MICROUNDE 

(Rezumat) 

Sunt studiate proprietăŃile controlabile ale straturilor subŃiri de granat de fier şi 

itriu (yttrium iron garnet, YIG), cu aplicaŃii la dispozitivele de microunde. Pentru 

implementarea practică a unor structuri compuse cât mai performante pe bază de straturi 

de YIG, cercetarea se canalizează pe două direcŃii în ceea ce priveşte proprietăŃile 

speciale ale acestui material: posibilităŃile de reglaj a frecvenŃei de rezonanŃă, respectiv


obŃinerea permeabilităŃii magnetice negative controlabile la metamateriale cu straturi de 

YIG. Studiul a fost realizat prin metoda simulării structurale cu ajutorul programului 

HFSS 13.0 de la Ansoft. Structura de tip metamaterial considerată constă dintr-un strat 

de YIG (100...850 µm) şi o reŃea metalică periodică din fire de cupru (de 0.3 mm, 

separate la 1 mm), testată în câmpul de microunde (12...20 GHz) al unui ghid 

dreptunghiular. Câmpul magnetic constant de prepolarizare al YIG-ului a fost setat la 

3...6.5 kOe. 

În urma simulărilor s-a obŃinut un domeniu de reglaj pentru frecvenŃa de 

rezonanŃă în intervalul 12...20 GHz, prezentând o evoluŃie polinomială în raport cu 

câmpul de prepolarizare, confirmată punctual de rezultatele experimentale raportate în 

literatură. Nivelul optim de reglaj pentru câmpul de prepolarizare este de cca. 3...4.8 

kOe, pentru care răspunsul materialului este consistent şi energia consumată pentru 

reglaj este mică. 

Componentele reale ale tensorului permeabilitate magnetică iau valori negative 

pentru YIG-ul incorporat în metamaterialul considerat în intervalul de frecvenŃe de 13... 

20 GHz, depinzând de parametrii geometrici ai structurii metamateriale şi de câmpurile 

de control cc/ca. Forma curbelor ce descriu evoluŃia cu frecvenŃa a acestor componente 

tensoriale depinde de câmpul de prepolarizare; un domeniul mai larg de permeabilitate 

negativă corespunde unui câmp de prepolarizare mai slab, dar acest domeniu se 

plasează la frecvente mai înalte pentru câmpuri mai intense (de până la 6 kOe). Grosimea 

stratului de YIG reprezintă un puternic parametru de control al domeniului de 

permeabilitate negativă a metamaterialului. LăŃimea acestui domeniu scade până la 

anulare odată cu creşterea grosimii stratului de la 100 la 860 µm (fenomenul prezintă 

efect de prag). Aceasta are drept consecinŃă anularea benzii de propagare permise, în 

care ambii parametrii electromagnetici ai metamaterialului: permitivitatea electrică 

(asigurată de reŃeaua metalică) şi permeabilitatea magnetică (asigurată de stratul de 

YIG) sunt negativi. Studiul este dedicat optimizării performanŃelor structurilor metamateriale 

pe bază de YIG, în domeniul microundelor.


Publicat de 


Tomul LVII (LXI), Fasc. 1, 2012 

SecŃia 


EXPERIMENTAL DETERMINATION OF 

AND ON TURNING STEELS 

BY 

ALIN LUCA * , MIRCEA COZMÎNCĂ and ANA MARIA MATEI 

Received: July 25, 2011 


”Gheorghe Asachi” Technical University of Iasi, 


Abstract. Establishing the roughness of cutting parts is very important 

because it affects the functioning of the work pieces. This influence reflects the 

operating characteristics of the piece such as wear resistance, corrosion, the 

coefficient of friction, the degree of tightness, etc. Surface roughness is 

characterized by different roughness parameters and the most important are the 

arithmetic average deviation of the profile Ra and ten-point mean roughness 

Rz.This paper aims to highlight the relationship between Ra and Rz by measuring 

the roughness of specimens subjected to the turning process. 

Key words: cutting, roughness, roughness parameters. 


General Consideration. In machining processes, it is necessary to obtain 

the desired surface roughness of the parts in the prescribed parameters in order 

to produce parts providing the required functioning. The most used parameters 

to describe the roughness of machined surfaces are Ra and Rz. The values of 

these two parameters are chosen from SR ISO 4287 due to the conditions of 

turning surface that has to be met within the product. Turning is a widely used 

manufacturing process in which a knife with a cutting edge removes the 

undesired material from a moving cylindrical workpiece. The feed rate of the 

cutting tool is subject to movements, parallel to the axis of rotation of the 

workpiece. Turning takes place on a lathe which provides the power needed to 

* Corresponding author: e-mail: alin_luca83@yahoo.com

86 Alin Luca et al. 

process the workpiece at a certain cutting speed, feed rate and depth of cut. 

Because numerous studies have shown that the influence of the depth of cut has 

a very small influence on the surface roughness (Fnides, 2009), (Sood et al., 

2000), (Doniavi et al., 2007), (Cakir, 2009) this parameter will not be taken into 

consideration . Instead, the main attack angle, k affects significantly the surface 

roughness (Constantinescu, 1998), (Dragu et al., 1982). Therefore, it is 

necessary to determine the influence of these three parameters, cutting speed, 

feed rate and main attack angle on turning surfaces. 

2. Planning and Performing the Experiments 

In practice, the requirements of the drawings and specific checks aim 

often for the arithmetic mean roughness, Ra, or for the ten-point mean 

roughness, Rz, Ra can be chosen from Table 1. 

Ra 

[µm] 

0.012 

0.025 

Table 1 

Rational choice of Ra according to the surface usability 

Surface characteristics Examples 

High contact tensions and very low 

wear 

High precision spindle shafts of tool 

machine 

0.05 Very low wear Bearers of bearings 

0.1 

0.2 

0.4 

0.8 

1.6 

3.2 

6.3 

12.5 

25 

50 

100 

Surfaces with high precision and high 

tension 

Surfaces subject to wear and high 

precision 

Surface subjected to mean speeds and 

pressures. Surfaces of alignment 

Reduced wear at low speeds and 

tensions 

Guiding and centering surfaces with 

periodic movements 

Contact surfaces without moving 

Visible exterior surfaces 

Rough contact surfaces without 

moving. Free surfaces of holes 

Seals. The rolling elements of 

bearings 

Front seals. Shafts and bearings 

Motion screw. Sliding surfaces 

Centering surfaces 

Areas under the felt seal. Shevered 

teeth or rectified. Bolted joints subject 

to vibration 

Keying, armatures, fixed hub-shaft 

fits, discs 

Removable fixed fits. Channels for Vbelts. 

Threads. Milled teeth 

Surfaces with appearence issues. 

Front surfaces of the shafts 

Rough surfaces, raw, peeled Castings, forgings, rolled, molded, 

pressed 

As for Rz, an approximate correspondence between the values of these 

two parameters is given by ratio (1):


Rz= 4...5Ra . (1) 

To study this statement it has been made ten try-outs to verify the 

statement made by Eq. (1) and to see how this ratio varies. Tests were 

conducted on specimens of OL50 (STAS 500/2-80) with a length l = 250 mm 

and a diameter D = 47.5 mm. Specimens were subjected to turning on a SN 250 

lathe in dry conditions. The material was chosen because it is wide usage in 

manufacturing. To measure the two parameters, Ra and Rz a Taylor Hobson - 

Surtronic 25 (2d) roughness tester was used. As shown in Table 2, three and 

four values for each parameter studied were chosen. 

Table 2 

Distribution of analyzed parameters 

Analyzed parameters 

v, [m/min] f, [mm/rot] k, [degrees] 

90 0.08 0 

120 0.12 30 

180 0.16 60 

75 

Parameter values were chosen after studying the matter of speciality. To 

see how these parameters affects the Eq. (1) were made 10 tests. 

3. Results and Discussion 

In Tables 3…5 are shown the experimental results of Ra and Rz and the 

ratio Rz/Ra for the studied parameters. 

Table 3 

Setting up the experiment and the experimental results of the influence 

of cutting speed on Ra and Rz and on the ratio Rz/Ra 

v, m/min Ra Rz Rz/Ra 

90 4.74 20.60 4.35 

120 4.59 18.80 4.10 

180 4.60 18.70 4.07 

To highlight the influence of each parameter on turning surface 

roughness were drawn Figs. 1…3. 

Table 4 

Setting up the experiment and the experimental results of the influence of feed 

rate on Ra and Rz and on the ratio Rz/Ra 

f, mm/rev Ra Rz Rz/Ra 

0.08 1.70 8.26 4.86 

0.12 4.63 18.95 4.10 

0.16 4.84 22.45 4.64


Fig. 1 – Main effects plot of cutting speed on Rz/Ra ratio. 

From Fig. 1 it can be seen that the ratio value decreases with the 

increasing of speed. 

Fig. 2 – Main effects plot of feed rate on Rz/Ra ratio. 

From Fig. 2 it can be seen that the ratio value drops significantly when 

the feed rate f = 0.12 mm/rev. 

Table 5 

Setting up the experiment and the experimental results of the influence of main 

attack angle on Ra and Rz and on the ratio Rz/Ra 

K, degrees Ra Rz Rz/Ra 

0 3.33 16.40 4.92 

30 4.48 18.40 4.11 

60 5.95 25.85 4.35 

75 6.40 28.50 4.45 

From Fig. 3 it can be seen as the higher reported value is obtained at k=0 ° 

while the smaller reported value is obtained at k =30 ° .


Fig. 3 – Main effects plot of main attack angle on Rz/Ra ratio. 


1. The statement it is verified experimentally, the values of the Rz/Ra ratio 

are between 4.07 and 4.92. 

2. Smaller values of Rz/Ra ratio were obtain when k=30 ° , f=0.12 mm/rot 

and v = 120 m/min, higher value of the ratio were obtained when the parameters 

studied were smaller. 

3. The value of the ratio increases inversely with the value of Ra, namely 

at lower values of Ra (obtained at finishing) the report moves toward maximum. 

Acknowledgements. This paper was realised with the support of 

CUANTUMDOC “Doctoral Scholarships for research and innovation performance” 

project, financed by the European Social Found and Romanian Government. 

REFERENCES 

Cakir C. M., Ensarioglu C., Demirayak I., Mathematical Modeling of Surface 

Roughness for Evaluating the Effects of Cutting Parameters and Coating 

Material. Journal of Materials Processing Technology, 209, 102-109 (2009). 

Constantinescu C., Teoria aşchierii în mecanica fină-îndrumar pentru lucrări practice. 

Ed. „Gh. Asachi” Iaşi, 1998. 

Doniavi A. et al., Empirical Modeling of Surface Roughness in Turning Process of 1060 

Steel Using Factorial Design Methodology. Journal of Applied Sciences, 7 (17), 

2509-2513 (2007). 

Dragu D. et al., ToleranŃe şi măsurători tehnice. Ed. Didactică şi Pedagogică, Bucureşti, 

1982. 

Fnides B, Surface Roughness Model in Turning Hardened Hot Work Steel Using Mixed 

Ceramic Tool. Mechanika, 3 (77) (2009). 

*** http://www.omtr.pub.ro/didactic/ssim.pdf. 

*** http://www.penet.ucoz.com/Cap10.doc. 

Sood R., Guo C., Malkin S., Turning of Hardened Steels. Journal of Manufacturing 

Processes, 2, 3 (2000).


*** OŃeluri de uz general pentru construcŃii. Mărci. STAS 500/2-80. 

STABILIREA RELAłIEI DINTRE Ra ŞI Rz LA 

STRUNJIREA OłELULUI OL50 

(Rezumat) 

Se urmăreşte variaŃia raportului Rz/Ra, la strunjirea oŃelului OL50. Pentru a 

verifica în ce măsură afirmaŃia conform căreia Rz= 4...5Ra este adevărată, s-au efectuat o 

serie de 10 încercări pe o epruvetă avand diametrul d = 47,52 mm, supusă operaŃiei de 

strunjire, pe un strung SN250, variind principalii factori de influenŃă asupra rugozităŃii, 

respectiv viteza de aşchiere v, avansul s şi unghiul de atac principal k. Valorile 

parametrilor rugozităŃii au fost masuraŃi cu ajutorul rugozimetrului Taylor Hobson - 

Surtronic 25. În urma experimentării s-a observat că afirmaŃia anterioară se verifică 

experimental.


Publicat de 



SecŃia 


FORCE SENSOR IN A WIM HYDROSTATIC SYSTEM 

BY 

IRINA MARDARE and IRINA TIłA * 

“Gheorghe Asachi”Technical University of Iaşi 

Department of Fluid Mechanics, Hydraulic Machines and Drives 



Abstract. Weigh-in-motion of road vehicles is essential for the management 

of freight traffic road infrastructure and maintenance. Various types of weigh in 

motion sensors have been developed. Existing WIM systems still have limited 

accuracy and/or excessively high cost, and their durability in many circumstances 

was not proven. We propose a hydraulic system with a force sensor. It is simple, 

robust and relatively cheap. In this paper is presented such a system. On these 

bases it was developed an experimental setup and in the same time the 

SimHydraulics model so one can use only the simulation in the design process in 

order to obtain maximum efficiency. 

Key words: hydraulic system, force sensor, weigh-in-motion, pressure 

measurement. 

1. Preliminaries 

A considerable demand has emerged in recent years for more accurate 

and reliable weigh-in-motion (WIM) systems and sensors in order to provide 

road authorities and managers with up to date and online measurements of axle 

and vehicle weights. WIM of Road Vehicles is essential for the management of 

freight traffic road infrastructure design and maintenance and the monitoring of 

vehicle and axle loads. Various types of WIM sensors have been developed: 

banding plates (McCall&Vodrazka 1997), piezoceramic cables and bars 

(Caprez et al., 2000), piezoquartz bars, capacitive strips and mats (McNulty & 

O'Brien, 2003), optical fiber, strain gauge or load cell scales (Chang et al., 

* Corresponding author: e-mail: iddtita@yahoo.com

92 Irina Mardare and Irina TiŃa 

2000), fiber Bragg grating (Zhang et al., 2008), bridge weigh-in-motion 

technology (Liljencrantz et al., 2007) etc. Advances in sensor and electronic 

technology have resulted in operation systems since the end of the 1980’s. 

Such systems (banding plates, piezoceramic cables and bars, piezoquartz 

bars, capacitive strips and mats) have been installed and are in operation in 

some countries. Existing WIM systems still have limited accuracy and/or 

excessively high cost, and their durability in many circumstances was not 

proven. Other WIM systems are for now research object or prototype. 

We propose a hydraulic WIM system which is very simple, has only one 

sensing point, and does not require electronic built-in-the road sensors. The 

hydraulic system is supposed to be simple, robust and relatively cheap 

2. The Diagram of the Hydraulic System 

In this paper is proposed a structure for the oil-filled force sensor system 

which is presented in Fig. 1. 

Fig. 1 – The structure of the force sensor. 

The oil filled chamber is deformable only on its upper wall; lateral walls 

are thought to be rigid. The volume of the chamber is V. This chamber acts like 

a hydraulic force cell. Under the action of force F the volume changes with ∆V. 

This change of volume produces pressure p1. The pipe is connected to the 

ℓ it is placed a fixed orifice (diameter d0). The cross 

chamber. At the distance 1 

section A0 of the orifice is chosen as it acts like a “low-pass” filter. The pressure 

p2 before the orifice is smaller than p1 with linear pressure loss ∆pλ. 

Accumulator Ac compensates slow pressure variations due to the temperature 

changes and is place at the distance ℓ 2 downstream the orifice. 

3. The Simscape Functional Diagram 

The Simscape functional diagram will simulate the functional elements of 

the system presented in Fig.1. 

The functional diagram for modelling the system is shown in Fig. 2.

Fig. 2 – The Simscape functional diagram. 

Sine Wave 

S PS 

Simulink-PS 

Converter 

MTR1 

S 

C 

C 

R V 

P 

Ideal Translational 

Motion Sensor 


Velocity Source 

R 

f(x)=0 

Solver 

Configuration 

P 

A 

Translational 

Hydro-Mechanical 

Converter 

R 

MTR2 

Hydraulic Piston 

Chamber 

PS S 

PS-Simulink 

Converter4 

PS S 

PS-Simulink 

Converter2 

A 

C 

Scope4 

Scope2 

MTR3 

C2 C1 

A B 

A B 

Pipe+ 

B 

A 

Inert P PS S 

A 

B 

P 

Hydraulic Pressure 

Sensor 

Custom Hydraulic 

Fluid 

Hydraulic 

Pipeline1 

Fluid 

Inertia 

Fixed Orifice 

A B 


Sensor3 

PS-Simulink 

Converter5 

PS S 

PS-Simulink 

Converter 

Hydraulic 

Pipeline3 

Fluid 

Inertia1 

Hydraulic 

Reference1 

A B 

A B 

Hydraulic 

Reference4 

Scope 

A 

B 

P 


Sensor1 

Hydraulic 

Pipeline2 

A B 

A 

B 

P 


Sensor2 

Hydraulic 

Reference2 

PS S 

PS-Simulink 

Converter1 

Fluid 

Inertia2 

A B 

PS S 

PS-Simulink 

Converter3 

Scope1 

Hydraulic 

Reference3 

Gas-Charged 

Accumulator 

Scope3 

Scope5 

Bul. Inst. Polit. Iaşi, t. LVIII (LXII), f. 1, 2012 93


The diagram in Fig. 2 makes possible to set different values for 

constructive or functional parameters. One can use it for design of such systems. 

The diagram uses functional elements from Simscape library and especially 

from SimHydraulics one. In this diagram one can see the blocks for the force 

sensor, for the pipes considering friction losses and inertial losses, and the 

pressure or flow sensors. 

There are some elements there are not among those in the library and it 

was necessary to use a new association of elements in order to obtain them. 

The diagram offers also some points with Scopes for one to diagnose the 

behaviour of each functional element or association of functional elements. 

The input which is the force may be a constant, may be step, sinusoidal or 

any other type of signal. 

The Simscape diagram presented in this paper is relatively easy to use 

and requires experience in hydrostatic systems in order to assign the parameters 

for functional elements. 

In a previous paper is shown the Simulink block diagram for the same 

system (TiŃa & Mardare, 2007). 

4. Experimental Equipment 

Experimental equipment is shown in Fig. 3. The experimental setup 

includes two pressure sensors, pressure gauge and accumulator. The sinusoidal 

variation of the force is obtained using a piston actuated with an electric motor 

and a cam which sets the stroke of the piston. 

Fig. 3 – Experimental equipment. 

5. Conclusions and Further Research 

1. A hydrostatic force sensor may be used in a WIM system. 

2. For the proposed structure of the system (Fig.1), it is possible to realise 

the Simscape functional diagram. MATLAB programming language is suitable 

for hydrostatic systems and the Simscape method, using functional elements, is 

relatively easy to use.


3. For the force sensor included in a WIM system the dynamic 

characteristics are essential. This is the reason for the authors will continue the 

research using the functional diagram presented in this paper for the dynamic 

analysis of the system. 

4. For further researches the Control System Toolbox procedure is next to 

be used in order to compare the three methods: the Simulink method, the 

Simscape method and the Control System Toolbox procedure for the case of a 

WIM hydrostatic system. 

Acknowledgements. The present work has been supported from the Grant 

(CNCSIS) PNII, 2703/22-111/2008. 

REFERENCES 

Caprez M., Doupal E., Jacob B., O'Connor A., O'Brien E., Test of WIM Sensors and 

Systems on an Urban Road. International Journal of Heavy Vehicle Systems 

(IJHVS), 7, 2-3, 111-135 (2000). 

Chang W., Sverdlova N., Sonmez U., Streit, D., Vehicle Based Weigh-in-motion System. 

International Journal of Heavy Vehicle Systems (IJHVS), 7, 2-3, 205-218 (2000). 

Liljencrantz A., Karoumi R., Olofsson P., Implementing Bridge Weigh-in-motion for 

Railway Traffic. Computers & Structures, 85, 1-2, 80-88 (2007). 

McCall B., Vodrazka W., States’ Successful Practices Weigh-in-Motion Handbook. 

U.S. Department of Transportation, Federal Highway Administration: Available 

from: http://www.ctre.iastate.edu/research/wim%5Fpdf/ accessed: 2007-04-10, 

1997. 

McNulty P., O'Brien E.J., Testing of Bridge Weigh-In-Motion System in a Sub-Arctic 

Climate. Journal of Testing and Evaluation, 31, 6, Paper ID: JTE11686_316 

(2003). 

TiŃa I., Mardare I., Weigh-in-Motion Hydraulic System. Annals of DAAAM 

International for 2007 & Proceedings of the 18 th International DAAAM 

Symposium, 18, 2007, pp. 759-760. 

Zhang H., Wei Z., A Portable Multi-function WIM Sensor System Based on Fiber Bragg 

Grating Technology. 19-th International Conference on Optical Fiber Sensor, 

7004, 1-2, 456-457 (2008). 

SENZOR DE FORłĂ INCLUS ÎNTR-UN SISTEM WIM HIDROSTATIC 

(Rezumat) 

Pornind de la cazurile în care sunt necesari senzorii de forŃă, este avută în vedere 

utilizarea acestora la sistemele de cântărire in mişcare (WIM). Aceste sisteme sunt în 

atenŃia cercetătorilor, mai ales datorită aplicării lor la cântărirea autovehiculelor grele pe 

autostrazi. În lucrare este prezentat un sistem WIM realizat ca un circuit hidrostatic. 

Este detaliată structura propusă pentru un asemenea sistem. Pentru sistemul prezentat 

este realizată schema funcŃională Simscape. Aceasta permite analiza dinamică a 

sistemului. Pentru sistemele WIM, studiul comportării dinamice este esenŃial, motiv 

pentru care unul dintre obiectivele cercetărilor viitoare este şi acela al aplicării metodei


Control System din pachetul MATLAB şi realizarea unui studiu comparativ între metoda 

Simulink (TiŃa & Mardare, 2007), metoda Simscape din această lucrare şi metoda 

Control System. Validarea prin corecŃii, eventuale, de model se va face prin 

confruntarea cu caracteristicile ridicate pe standul experimental.


Publicat de 



SecŃia 


RESEARCHES CONCERNING THE UNIFORMIZATION OF 

CUTTING FORCES IN FACE MILLING 

BY 

ANA-MARIA MATEI ∗ , MIRCEACOZMÎNCĂ and ALIN LUCA 



”Gheorghe Asachi”Technical University of Iaşi, 


Abstract. This paper presents some theoretical results regarding the 

evaluation of cutting force’s components in face milling. Face milling force 

depends on the every variant of face milling process (number of teeth which 

simultaneously cut) and the relative position between cutting teeth and the 

material being cut (cut-down milling and cut-up milling). Regarding to this, 

determination of the cutting force components in face milling, FZ , FX , FY, is 

based on the forces for a single tooth, the cutting tooth position beside the XYZ 

coordinates system of the tool and the number of teeth that simultaneously cut. In 

order to standardize the relationships, a calculus of some practical situations is 

proposed. 

Key words: cutting, face milling, forces, constants. 


The valuation models of cutting force in face milling previously proposed 

(Cozmîncă et al., 2009 a), (Matei et al., 2010), (Matei et al., 2011) consider the 

influences of the specific elements acting on a tooth, through the relationships 

used to evaluate the cutting force in turning whose validity has been proven 

experimentally, and the influences of the specific elements in face milling: the 

possible variant of face milling, the number of teeth that of teeth that 

simultaneously cut and the cutting tooth position beside the XYZ coordinates 

system of the tool (cut-up and cut-down milling). 

∗ Corresponding author: e-mail: anca_reea@yahoo.com

98 Ana-Maria Matei et al. 

By reproducing the five individual types of face milling in the ZX plane 

of the cutter, it results the variation limits of the contact angle values, 

respectively the values of the number of teeth which simultaneously cut, zs, and 

from this point, directly, the differences between FZ, FX and FY for each face 

milling variant. The highest values for the force components FZ, FX and FY are 

obtained in complete face milling (Ψ5 = 180°, zs = z/2), and the lowest are 

obtained in asymmetrical milling with Ψ2 < 90° and zs ≥ 2 (Cozmîncă, 1995). 

Fig. 1 – The values of Ψ and zs for the five types of face milling. 

For the other types of face milling the values of force’s components F Z, 

F X and F Y depend on the surface width (t, mm) and the diameter D, namely on 

the number of cutter teeth, z. 

This paper presents some numerical tests of the models developed until 

this point and some new theoretical models for the evaluation of cutting force 

components in face milling with a higher degree of generality and easier to use 

by designers. 

2. Theoretical Researches Concerning the Face Milling 

Forces Uniformization 

2.1. Numerical Tests concerning the Application of the Evaluation 

Models of Face Milling Forces 

Considering the theoretical models for the evaluation of face milling 

forces developed in the previous papers, we propose some numerical 

determinations in order to verify their application in real situations. Thus, we 

considered three possible cases, namely the processing of an workpiece made 

from OLC60 using three face milling tools with the next features: F1 = face


milling cutter with D = 63 mm, z = 5, φ = 72°; F2 = face milling cutter with D = 

=100 mm, z = 10, φ = 36° and F3 = face milling cutter with D = 140 mm, z = 

=14, φ = 25.7°. 

Assuming that every mill is processing the workpiece in all 5 variants of 

face milling, for each of these cases, we calculated the number of teeth that 

simultaneously cut depending on every variant of milling, and then we 

calculated the values of cutting force components in face milling using the 

relationships from the previous papers (Matei et al., 2011). 

First of all an evaluation of the forces with an average load from a tooth 

level is necessary and then an evaluation of the resulting force of the cutter with 

z s teeth that simultaneously cut. In the paper (Matei et al., 2011) we have shown 

that to calculate the force acting on a tooth we must consider the particularities 

of the face milling process, especially the chip thickness variation, since these 

relations are used then to calculate the force for z s teeth that simultaneously cut. 

To run the theoretical tests for the new valuation models of face milling 

force, first we propose a calculation of the average value of the force acting on a 

tooth. Thus we developed an example for calculating the force components 

acting on a tooth, considering that the working parameters (t, s, v) and the 

geometric parameters (γ, α, λ, K) are as close as possible to those used in face 

milling (Cozmîncă, 1995), (Cozmîncă et al., 2009 b).. The results are presented 

in the Table 1. 

Working 

conditions, 

constant 

t = 4 mm 

γ = 5° 

α = 12° 

λ = 5° 

K = 60° 

Cd = 1.66 

n = 1 

µ = 0.5 

k1 = 1.0397, k2 = 

=0.4110, k3 = 

0.0435, k4 = 

0.0754 

C1=1.0357, C2 = 

=0.4167, C3 = 

=0.1272 

Table 1 

The average values for the force’s components acting on 

a tooth – processing with F1 and F2 

Face milling 

variant 

Asymmetrical 

face milling, Ψ = 

=90°, t = D/2 

Cutter 

used 

Feed, s 

[mm/rot] FZ [N] FX [N] FY[N] 

F1 0.097 33.354 13.419 4.096 

F2 0.085 29.227 11.759 3.590 

Complete (full) F1 0.114 39.199 15.771 4.814 

symmetrical face 

milling, t = D 

F2 0.085 29.227 11.759 3.590 

Asymmetrical F1 0.097 33.354 13.419 4.096 

face milling, Ψ < 

 

>90°, t > D/2 

F2 0.106 36.448 14.665 4.476 

Incomplete F1 0.12 41.262 16.601 5.068 

symmetrical 

face milling 

F2 0.106 36.448 14.665 4.476


For processing with the milling cutter F3 we have chosen the same 

working conditions described above excepting the cutting depth which is equal 

to 6 mm for this case. The results were represented in Table 2. 

Table 2 

The average values for the force’s components acting on a tooth – processing with F3 

Face milling variant 

Feed, s 

[mm/rot] FZ [N] FX [N] FY[N] 

Asymmetrical face milling, Ψ = 90°, t = D/2 0.084 43.325 17.431 5.321 

Complete (full) symmetrical face milling, t = D 0.084 43.325 17.431 5.321 

Asymmetrical face milling, Ψ < 90°, t < D/2 0.084 43.325 17.431 5.321 

Asymmetrical face milling, Ψ >90°, t > D/2 0.113 58.283 23.449 7.158 

Incomplete symmetrical face milling 0.113 58.283 23.449 7.158 

With the average values of the force components acting on a tooth, 

further we can obtain the values for the force components in face milling for all 

five possible variants. The results were centralized in Table 3. 

As a result of a comparative analysis of the values obtained, the following 

conclusions can be drawn: 

i) the values of cutting force varie proportionally to the number of teeth 

that simultaneously cut (if asymmetrical milling); 

ii) the values of cutting force are bigger in incomplete symmetrical face 

milling than in complete face milling; this variation is due to the chip thickness, 

so in the complete milling, the chip thickness varies from the minimum value 

a min = 0 up to the maximum a max = s d, while in incomplete milling the chip 

thickness varies from a certain thickness to the maximum; 

iii) the components values will not be in the well known ratio of turning, 

respectively, Fz > Fx > Fy, so in most of the cases, F X will have a higher value 

for cut – down milling for all five possible variants, and F Y will have a higher 

value in full symmetric milling; 

iv) there are some differences between processing with F1 and the other 

milling cutters because of the small number of cutting teeth that do not comply 

with the exactly number of teeth that simultaneously cut corresponding to each 

variant; 

v) theoretical models used to calculate the force are relatively complex 

and difficult to use. 

2.2. New Theoretical Models for the Evaluation of 

Cutting Force in Face Milling 

In order to unify the relationships for the evaluation of face milling force 

developed in the previous papers (Matei et al., 2011) we consider a simplifying 

assumption according to which the force components developed on a single


tooth, F z, F x and F y are equal for all those z s cutting teeth that simultaneously 

cut. Therefore the cutting force components take values according to the cross – 

sectional area of the chip, so we have F z = F zmed, F x = F xmed and F y = F ymed for 

an average thickness of the chip a = a med. 

Cutter 

used 

Table 3 

Numerical tests concerning the face milling force 

Cutting force components depending on the every variant of face milling and 

the forces developed on a single tooth level FZ, FX, FY 

Asymmetrical face milling with Ψ = 90° 

Cut – up face milling Cut – down face milling 

FZ FX FY FZ FX FY 

F1 44.321 (-) 37.209 6.145 37.209 44.321 6.145 

F2 43.386 -69.412 8.974 70.353 43.386 8.974 

F3 80.537 (-) 142.645 18.624 142.645 80.537 18.624 

Asymmetrical face milling with Ψ < 90° 



F1 44.321 (-) 37.209 6.145 37.209 44.321 6.145 

F2 37.269 (-) 69.037 8.210 69.037 37.269 8.210 

F3 80.537 (-) 142.645 18.624 142.645 80.537 18.624 

Asymmetrical face milling with Ψ > 90° 



F1 30.170 11.832 12.669 21.370 111.456 12.669 

F2 44.696 (-) 49.044 15.667 124.226 64.096 15.667 

F3 73.462 (-) 74.927 39.369 308.855 143.221 39.369 

Complete (full) symmetrical face milling 

FZ FX FY 

F1 27.196 (-) 15.041 14.443 

F2 17.580 (-) 14.544 17.948 

F3 22.325 (-) 24.587 37.247 

Incomplete symmetrical face milling 

FZ FX FY 

F1 13.568 (-) 29.430 12.669 

F2 45.061 (-) 49.044 15.667 

F3 73.462 (-) 74.927 39.369 

Analyzing the mathematical models developed in the previous papers one 

can observe the existence of some constants in the form of sums in their 

structure.


These constants get different nuances depending on the variant of milling 

and the cutting teeth position beside the XYZ coordinates system of the tool and 

the number of teeth that simultaneously cut, and we’ll call them C zteoretic and 

C xteoretic. They are used to calculate the values of the tangential component F z 

and the radial component F x, respectively. 

Table 4 presents the values of theoretical constants depending on the 

factors listed above. 

Replacing the constants in the structure of the mathematical models of 

previous papers and considering the assumptions described above, we obtain the 

Eqs. (1) to determine the cutting force components in face milling. 

The variant 

of face 

milling 

Asymmetrical 

face 

milling with 

Ψ = 90° and 

Ψ < 90° 

F = F C + F C . 

Z z z x x 

med teoretic med teoretic 

F = F C + F C . 

X x x z z 

med teoretic med teoretic 

F = F z . 

Y y s 

med 

Table 4 

The values of theoretical constants depending on 

the specific influencing factors at face milling 

The 

relative 

position of 

the tool 

and the 

material 

being cut 

Fac 

e 

mill 

ing 

forc 

es 

com 

ponent 

s 

FZ 

Cut – up 

face 

milling FX 

Cut – 

FZ 

zs 

∑ 

1 

C zteoretic 

⎛ 2π 

⎞ 

sin ⎜ zsi 

z 

⎟ 

⎝ ⎠ 

zs 

⎛ 2π 

⎞ 

−∑ cos⎜ 

zsi 

1 z 

⎟ 

⎝ ⎠ 

zs 

∑ 

1 

⎛ 2π 

⎞ 

cos⎜ 

zsi 

z 

⎟ 

⎝ ⎠ 

C xteoretic 

zs 

⎛ 2π 

⎞ 

−∑ cos⎜ 

zsi 

1 z 

⎟ 

⎝ ⎠ 

zs 

⎛ 2π 

⎞ 

−∑ sin ⎜ zsi 

1 z 

⎟ 

⎝ ⎠ 

zs 

∑ 

1 

⎛ 2π 

⎞ 

sin ⎜ zsi 

z 

⎟ 

⎝ ⎠ 

(1)

Asymmetrical 

face 

milling with 

Ψ > 90° 


down face 

milling FX 

Cut – up 

face 

milling 

Cut – 

down face 

milling 

Symmetrical complete 

and incomplete face 

milling 

FZ 

FX 

FZ 

FX 

FZ 

FX 

zs 

∑ 

1 

1 

⎛ 2π 

⎞ 

sin ⎜ zsi 

z 

⎟ 

⎝ ⎠ 

z/4 

2π 

∑sin( 

zsi 

) + 

1 z 

zs −( 

z/4) 

2π 

+ ∑ cos( zsi 

) 

z 

z/4 

2π 

− ∑cos( 

zsi 

) + 

z 

+ 

1 

zs −( 

z/ 

4) 

z/4 

∑ 

1 

+ 

∑ 

1 

zs −( 

z /4) 

z/4 

∑ 

1 

− 

1 

2π 

sin( zsi 

) 

z 

2π 

cos( zsi 

) + 

z 

∑ 

zs −( 

z/4) 

zs 

/2 

∑ 

1 

+ 

1 

2π 

sin( zsi 

) 

z 

2π 

sin( zsi 

) − 

z 

∑ 

zs 

/2 

1 

2π 

cos( zsi 

) 

z 

2π 

sin( zs 

) + 

i z 

∑ 

2π 

cos( zs 

) 

i z 

zs 

/ 2 

2π 

− ∑ cos( zs 

) + 

i z 

+ 

1 

zs 

/2 

∑ 

1 

2π 

sin( zs 

) 

i z 


zs 

⎛ 2π 

⎞ 

−∑ cos⎜ 

zsi 

1 z 

⎟ 

⎝ ⎠ 

z/4 

2π 

− ∑cos( 

zsi 

) + 

z 

+ 

1 

zs −( 

z/4) 

∑ 

1 

1 

zs −( 

z /4) 

1 

2π 

sin( zsi 

) 

z 

z/4 

2π 

−∑sin( zsi 

) − 

z 

− 

z/ 

4 

∑ 

1 

− 

∑ 

zs −( 

z / 4) 

1 

2π 

cos( zsi 

) 

z 

2π 

sin( zsi 

) − 

z 

1 

∑ 

zs −( 

z /4) 

1 

2π 

cos( zsi 

) 

z 

z /4 2π 

−∑cos( zs 

) − 

i z 

− 

∑ 

1 

zs 

/2 

1 

2π 

sin( zs 

) 

i z 

zs 

/2 

2π 

− ∑ cos( zsi 

) + 

z 

+ 

∑ 

1 

zs 

/2 

1 

2π 

sin( zsi 

) 

z 

zs 

/2 

2π 

−∑ sin( zsi 

) − 

z 

− 

∑ 

2π 

cos( zsi 

) 

z 

1. The values of the force’s components FZ and FX acting on the cutter 

depend through Cz theoretic and Cx theoretic on the face milling variant (symmetrical 

or asymmetrical), the type of face milling (cut-up or cut-down milling) and the 

number of teeth that simultaneously cut. 

2. Working conditions, respectively the cutting regime, tooth geometry, 

nature of the material being cut and the environment are present in FZ, FX and 

FY through the corresponding components acting on a cutting tooth.


3. For each mill and variant of face milling a number of teeth that 

simultaneously cut is assigned, respectively certain values are obtained for FZ 

and FX from the cutter level. The FY component takes values based on the value 

of Fy component on a single tooth and the number of teeth that simultaneously 

cut. 

4. The valuation models of face milling force presented in the end of this 

paper were developed in a more accessible structure, and arranging the 

constants in a table makes them relatively easy to use. 

Acknowledgements. This paper was realised with the support of EURODOC 

“Doctoral Scholarships for research performance at European level” project, financed 

by the European Social Found and Romanian Government. 

REFERENCES 

Cozmîncă M., Bazele aşchierii. Ed. ”Gheorghe Asachi”, Iaşi, 1995. 

Cozmîncă M. et al., About the Cutting Forces at Face Milling. Bul. Inst. Polit. Iaşi, 

LV(LIX), 2, s. ConstrucŃii de Maşini, (2009a). 

Cozmîncă M. et al., A New Model for Estimating the Force Components Fz, Fx and Fy 

when Cutting Metals with Single Tooth Tools. Bul. Inst. Polit. Iaşi, LV(LIX), 1, s. 

ConstrucŃii de Maşini, (2009b). 

Matei A. M., Cozmîncă M., Ibănescu R., Theoretical Models of Cutting Force 

Components at Face Milling. Bul. Inst. Polit. Iaşi, LVI(LX), 2b, s. ConstrucŃii de 

Maşini,. (2010) 

Matei A. M., Cozmîncă M., Luca A., Mathematical Models for the Evaluation of 

Cutting Force Components in Face Milling. Bul. Inst. Polit. Iaşi, LVII(LXI), 4, s. 

ConstrucŃii de Maşini, (2011) 

CERCETĂRI PRIVIND UNIFORMIZAREA FORłELOR DE AŞCHIERE 

LA FREZAREA FRONTALĂ 

(Rezumat) 

Valorile forŃelor FZ şi FX de la nivelul frezei frontale depind prin mărimile 

Cz theoretic si Cx theoretic de varianta de frezare (asimetrică sau simetrică), de tipul frezării 

frontale (în contra sau sensul avansului) şi de numărul de dinŃi care aşchiază simultan. 

CondiŃiile de lucru, respectiv regimul de aşchiere, geometria dintelui, natura 

materialului aşchiat şi mediul de aşchiere, sunt prezente în FZ, FX şi FY prin intermediul 

componentelor corespunzătoare de la nivelul unui dinte aşchietor. Pentru fiecare freză şi 

pentru orice variantă de frezare frontală se asigură un număr de dinŃi care aşchiază 

simultan, respectiv se obŃin anumite valori pentru forŃele FZ şi FZ de la nivelul frezei. In 

ceea ce priveşte componenta FY, aceasta capătă valori în funcŃie de valoarea 

componentei Fy de la nivelul unui dinte aşchietor şi de numărul de dinŃi care aşchiază 

simultan. Variantele de frezare nu influenŃeaza mărimea componentei FY de la nivelul 

frezei frontale.


Publicat de 



SecŃia 


A PRELIMINARY STUDY OF THE KEY FACTORS FOR 

SUSTAINING TOTAL QUALITY PR ACTICES 

BY 

MEHRAN DOULAT ABADI 1* and SHA’RI MOHD YUSOF 2 

University of Technology, Malaysia, 

1 Department of Industrial Engineering and Management 

2 Department of Manufacturing and Industrial Engineering 



Abstract. Quality management is broadly recognized around the world as a 

very important competitive priority for the long-term success of an organization. 

Over the last decades, a numerous quality management concepts have been 

applied and been practiced by various organizations to attain world-class statue. 

However, the recognition of the successes through using the quality management 

approach historically has been made by obtain national quality or business 

excellence awards. Achieving excellence is hard enough at the best of times; 

sustaining it is even harder. Organizations face many challenges and impediments 

in sustaining a quality excellence framework in the long run to drive business 

improvement. This paper will attempt to investigate and identify the key factors 

affecting the implementation of quality award model through literature review and 

thus to list these factors. Hence, understanding of the factors will help the 

organizations to ensure that these are dealt with appropriately in their journey 

towards excellence. 

Key words: Total quality management, sustainability, business excellence, 

success factors. 


Quality management is broadly recognized around the world as a very 

important competitive priority for the long-term success of an organization over 

several decades. However, the recognition of the successes through using the 

* Corresponding author: e-mail: dmehran2@live.utm.my

106 Mehran Doulat Abadi and Sha’ri Mohd Yusof 

quality management approach historically has been made by obtain national 

quality or excellence awards (Laszlo, 1996). The national quality or excellence 

awards programs are the next major quality management event following TQM 

(Vora, 2002), (McAdam et al., 1998). Therefore, for many organizations 

participating in national quality or excellence award program is a way to 

support their TQM practices towards achieving world-class statue (Yusof & 

Aspinwall, 2000), (Puay et al, 1998), (Adebanjo, 2001), (Grigg & Mann, 

2008a). However, to get significant benefit and the best result, organizations 

must maintaining the practices with the awards principles for the long-term on 

the competitive path and made it a part of their organizational culture. Studies 

by (Coulambidou & Dale, 1995) and (Angell & Corbett, 2009) support this 

view. Today, the national quality/excellence frameworks and their criteria have 

been commonly accepted by many organizations as a powerful tool for 

assessing an organization along the quality and excellence path (Meers& 

Samson, 2003). Therefore, given today’s business climate, it would be hard to 

find an organization to ignore the practice on quality management approaches. 

However, achieving excellence is hard enough at the best of times; sustaining it 

in today’s world of increasing global competition, rapid technological 

innovation, changing processes and frequent movement in economic, social and 

customer environments, is even harder. 

2. Literature Review 

After the successful of quality management practices in Japan several 

countries established national quality/excellence award programs to pursue 

excellence in an effective way and to recognize which organizations employed 

the best quality management practices. All awards principles are strongly 

grouped based on the core of key principles and major constitutes of TQM 

(Ghobadian & Woo, 1994), (Ghobadian & Gallear, 1997), (Thompson& 

Simmons, 1997), (Hendricks & Singhal, 1999), (Zairi, 2001), (Tan et al., 2003). 

Therefore, getting a national quality/excellence award is a confirmation for 

TQM implementation successfully (Hendricks & Singhal ,1996); (Ghobadian & 

Gallear, 2001) and (Eriksson, 2004). Improvements as assessed against the 

quality/excellence framework will lead to long-term business success. As a 

result, there has been a trend in organizations to use quality-based initiatives as 

a source of competitive advantage. 

The literature suggests that the success of an organization by using the 

quality-based initiatives does not depend on individual quality tools and 

techniques, but it much depends on a range of general management practices, 

including: top management commitment and support, establishment of trust and 

communication, employee empowerment and motivation, common metrics 

across the organization, a stepwise problem solving approach, and standardised 

analysis using quality tools (Choo et al., 2007), (Easton & Jarrell, 1999), 

(Ehigie & McAndrew, 2005), (Hodgetts et al.,1999), (Powell, 1995), (Soltani et


al., 2005), (Venkateswarlu & Nilakant, 2005). Business excellence or 

organizational excellence and all round growth have been the spicy subject of 

discussion among management academics at the recent years. The first major 

study on the concept of ‘business excellence’ as a topic of academic research 

was undertaken by (Peters & Waterman, 1982) in the famous book, In Search of 

Excellence–Lessons from America’s Best-run Companies. This investigation, 

designed to understand what some leading US organizations were doing to 

succeed in the face of the Japanese-led quality revolution. They provide many 

suggestions for success criteria behind business excellence concept for 

‘excellent organizations’. 

2.1. TQM versus Business Excellence 

First, TQM as the fourth level of quality management (Van der Wiele et 

al., 1997) has been one of the major intuitive for improved productivity for 

almost over three decades ago. Although TQM is much older than that, the 

‘total quality movement’ really picked up steam in the late 1970s and early 

1980s when several large American corporations adopted the techniques that 

enabled the Japanese to be so successful. There is clear evidence that many 

organizations view TQM as the basis for excellence (Adebanjo, 2005). Business 

excellence is the goal of every modern organization and can be defined as the 

next step after TQM, for the success of enterprise on the competitive path (Vora, 

2002), (McAdam et al., 1998).The use of excellence models is popular for the 

same reasons that TQM became unpopular. (Adebanjo, 2005).The term of 

Organizational Excellence or Business Excellence is generally associated with 

the European Fundamental for Quality Management (EFQM) excellence model. 

EFQM to provide a model that ideally represents the business excellence 

philosophy that can be applied in practice to all organizations irrespective of 

country, size, sector or stage along their journey to excellence (Dommartin, 

2000). 

Table 1 shows a review of the philosophy, principles, process , 

performance, and problem indicators of both reveals that business excellence 

was and still is fundamentally based on the quality management concept and 

practices. On the other hand, the business excellence model provides a clear 

road sign for organizations to follow towards excellence, we may also note that 

both TQM and business excellence stress primarily the importance of 

continuous improvement. 

2.2. Key success Factors of Total Quality 

This section presents a review of the key common success factors or 

constructs of organizational excellence developed and utilized by researchers in 

previous studies. Because of limited resources, it is always not feasible for 

organizations to devote their efforts to concurrently address all the success


factors. Key common success factors or contributing variables or critical factors 

or enablers, in this study can be viewed as those things that must go right in 

order to ensure the successful implementation of quality management concepts 

such as Business Excellence (BX). In this paper, we tried to investigate key 

common success factors for BX based on extant literature review. The 

investigation of the key common success factors for successful the practices of 

BX are presented in Table 2. 

Table 1 

TQM versus Business Excellence 

Concepts(5Ps) Total Quality Management (TQM) Business Excellence (EFQM Model) 

Philosophy To combine people and quality 

techniques to achieve continuous 

improvement in the quality of the 

product and hence in all aspects of the 

operation (Harriss, 1995). 

Principles Customer focused, leadership, 

involvement people, process 

approaches, continual improvement, 

and supplier relationship. 

Process SPC Statistical Process Control) P-D- 

S-A (Plan, Do, Study, Act) 

performance Continuous improvement of the 

organization, Customer satisfaction, 

and employee development. 

Problem TQM is conceptual and philosophical. 

Its strong Ideological and culture 

perspective cannot be easily 

developed in companies (Salengun & 

Fazel, 2000) 

To assist organization to participate 

in improvement activities leading 

ultimately to excellence results and 

driving force for sustainable 

excellence (EFQM 2010). 

Result orientation, Customer focus, 

leadership and constancy of purpose, 

management by process and facts, 

people development and 

involvement, continuous learning, 

innovation and improvement, 

partnership development, and public 

responsibility. 

RADAR (Results, Approach, 

Deployment, Assessment, and 

Review). 

Customer results, people results, and 

society results. Key performance 

results. 

Business excellence needs to avoid 

evolving into a purely scoring, short- 

term oriented mechanism, losing the 

fundamentals of the quality focus. 

Various studies have been carried out for the identification of those 

factors of successful BX practices, from three different areas: contributions 

from quality gurus, formal evaluation BX models and empirical research. Out of 

the 38 different critical factors developed by the researchers, 11 were found to 

be the most popular critical factors for TQM, meanwhile 7 were found to be the 

most popular critical factors for BX. They are all critical factors for TQM and 

BX, ranked from the highest to the lowest level of popularity: top management 

commitment/support; product/service design; supplier quality management 

(some researchers used different terms such as vendor quality management; 

supplier chain management; supplier quality assurance; cooperative supplier 

relations; supplier management); process management (includes process


knowledge, process control and improvement, process analysis and 

improvement, process focus); human resource management (includes employee 

relations, employee empowerment, employee involvement, employee 

participation, employee management, providing assurance to employees, human 

resource development, employment continuity, work force commitment); use of 

tool and technique; work environment and culture; continuous improvement; 

customers focus (customer driven processes, customer orientation, and customer 

satisfaction/involvement); leadership; training and education. 

From the review, it can be summarized that critical factors for successful 

TQM implementation can be classified into three group of researchers: Group 

researchers I: ‘soft’ factors and ‘hard’ factors (Ahire et al., 1996), (Thiagaragan 

et al., 2001), (Lau & Idris, 2002), (Tari & Sabater, 2004), (Rahman & Bullock, 

2005), (Vouzas & Psychogios, 2007), (Fotopoulos & Psomas, 2009); Group 

researchers II: TQM factors can be divided into strategic factors, tactical 

factors, and operational factors (Salaheldin, 2009); and Group researchers III 

comprises the TQM framework into organizing, systems and techniques, 

measurement and feedback, and culture and people (Chin et al. 2002). 

The most important factors in the successful implementation of BX are 

full management support and commitment and giving the correct training to the 

right people at the right time (McQuater et al. 1995), (Bunney & Dale, 1997). 

Managers must understand the importance of their commitment in order to 

spread the use of these tools and techniques and to improve the TQM level & 

TQM results (Tari & Sabater, 2004). However, tools alone cannot provide 

results by themselves. They must be developed to reflect the companies’ culture 

and management vision (Govers, 2001). Tools and techniques also can be used 

to reinforce recommendations made to managers (McQuater et al. 1995). The 

key to improvement is to focus on the improvement objectives and 

recommendations, and use tools and techniques as an aid for that purpose. 

The most important factors in the successful implementation of BX are 

full management support and commitment and giving the correct training to the 

right people at the right time (McQuater et al., 1995), (Bunney & Dale, 1997). 

Managers must understand the importance of their commitment in order to 

spread the use of these tools and techniques and to improve the TQM level and 

TQM results (Tari & Sabater, 2004). However, tools alone cannot provide 

results by themselves. They must be developed to reflect the companies’ culture 

and management vision (Govers, 2001). Tools and techniques also can be used 

to reinforce recommendations made to managers (McQuater et al. 1995). The 

key to improvement is to focus on the improvement objectives and 

recommendations, and use tools and techniques as an aid for that purpose. 

As can be seen from the review, the CSFs for BX sustainability are very 

similar to the CSFs for TQM implementation due to its close. The proposed 

critical factors for effective implementation of BX sustainability are 

summarized in Table 3.


Table 2 

Key success factors of business excellence practices developed and utilized by researchers 

Freq 

Key Common Business Excellence Sustainability (BES) 

No. 

. 

Success Factors 

1 Top 

management 

commitment/su 

pport 

2 People 

management 

3 Middle 

management 

involvement 

4 Training and 

education 

5 Reward and 

recognition 

6 Teamwork and 

cooperation 

7 Quality policy 

and strategic 

planning 

8 Communicating 

for quality 

relation 

9 Supplier 

management 

10 Accredited 

quality 

management 

systems 

11 Organizing for 

quality 

12 Managing by 

process 

+ + + + + + + + + + + 11 

+ + + + + + + + + 9 

+ + + + + + + + + 9 

+ + + + + + + 7 

+ + + + + + + 7 

+ + + + + + + + + 9 

+ + + + + + + 7 

+ + + + + 5 

+ + + + 4 

+ + + + 4 

+ + + + 4 

+ + + + 4 

13 Benchmarking + + + 3 

14 Self-assessment + + + 3 

15 Cost of quality + + 2 

16 Quality control 

+ + + + + + 7 

techniques 

17 Measuring 

customer wants 

and satisfaction 

+ + + + + + + 7 

4. Discussions and Future Research 

A review of the literature shows that, TQM is rather than a mere set of 

factors, a network of interdependent components, a management system 

consisting of critical factors, techniques and tools (Hellsten & Klefsjo, 2000).


Table 3 

The proposed key factors for effective BES 

Key factors (criteria) Sub-factors (sub-criteria) 

Management 

responsibility 

Strategic quality planning/quality policy; the role of divisional top 

management; top management commitment/support; internal 

Resource 

stakeholders’ involvement (middle management involvement) 

Technology-and production related resources; financial-related resources; 

management information and communication-related resources 

People management Employee involvement/empowerment; education; and training; 

Quality in design and 

teamwork and cooperation; work environment culture 

Process management/operating procedures; role of quality department; 

process 

product design; process analysis and improvement; applied quality tools 

and techniques 

Measurement, Quality measurement, feedback and benchmarking; continuous 

analysis & feedback improvement; performance measurement: external and internal; quality 

data and reporting; communication to improve quality; recognition and 

rewards; quality systems 

Supplier 

Supplier quality management/supplier chain management; contact with 

management supplier and professional associates 

Customer focus Customer 

processes 

involvement/satisfaction/orientation; customer driven 

These techniques and tools are vital to support and develop the quality 

improvement process (Bunney & Dale, 1997), (Hellsten & Klefsjo, 2000), 

(Curry & Kadasah, 2002), (Tari, 2005). BX in general and its principles and 

criteria in particular are, alongside critical factors, another important component 

of TQM, which emphasizes their importance for the improvement of quality of 

business and results. (Tari & Sabater, 2004) suggested that firms must develop 

both the hard and the soft parts of TQM in order to succeed. With the passage 

of time and with changing customer’s needs and expectations the word of 

‘quality’ has been replaced by ‘excellence’. As part of this quality progress, 

business or organizational excellence has become a recent goal of quality 

management movement (Fang Zhao, 2004). Therefore, business excellence 

practice can be named as the fifth level of quality management and next step 

after TQM, for the success of an organization (McAdam et al., 1998), (Vora, 

2002) in this never-ending journey. However, there has been a lack of empirical 

and published research and any comprehensive studies reported in the literature 

focusing on and revealing factors affecting implementation of BX principles at 

management level of winner organizations. 

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Survey Results. Total Quality Management, 11( 4/5 & 6), 448-462 (2000). 

Yusof S. M., Aspinwall E. M., Critical Success Factors for TQM Implementation in 

Small and Medium Enterprises. The TQM Magazine, 10, 803-809 (1999). 

Zairi M., TQM: What Is Wrong with the Terminology? The TQM Magazine, 6(4), 6 -8 

(1994). 

Zhang Z., Waszink A., Wijngaard J., An Instrument for Measuring TQM 

Implementation for Chinese Manufacturing Companies. International Journal of 

Quality and Reliability Management, 17(7), 730-755 (2000). 

Zhang,Z., Implementation of TQM: An Empirical Study of Chinese Manufacturing 

Firms. Universiteit Gronigen, Netherlands, 2001. 

STUDIU PRELIMINAR ASUPRA FACTORILOR CHEIE CARE SUSłIN 

PRINCIPIILE MANAGEMENTULUI CALITĂłII TOTALE 

(Rezumat) 

Managementul total este unamim recunoscut în întreaga lume ca o proritate 

esenŃială în competiŃia pentru asigurarea succesului pe termen lung al unei organizaŃii. 

În ultimele decenii, numeroase concepte ale managementului calităŃii au fost aplicate de 

diferite organizaŃii în vederea obŃinerii statutului de ”world class”. Cu toate acestea, 

recunoaşterea succesului prin aplicarea managementului calităŃii a fost întotdeauna 

realizat prin obŃinerea de premii pentru calitate naŃională sau excelenŃă în afaceri.


Atingerea excelenŃei este extrem de grea în majoritatea cazurilor, dar şi mai grea este 

păstrarea acesteia în timp. OrganizaŃiile trebuie să facă faŃă la numeroase încercări şi 

impedimente pentru menŃinerea excelenŃei în calitate în lungul drum al conducerii cu 

succes a afacerii. Articolul de faŃă intenŃionează să investigheze şi să identifice factorii 

cheie care afectează modelului calităŃii prin prisma literaturii de specialitate. Ca urmare, 

înŃelegerea acestor factori de influenŃă se va constitui într-un instrument pentru diferite- 

le organizaŃii în drumul către obŃinerea şi menŃinerea excelenŃei în calitate.


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SYNTHESIS ON THE ASSESSMENT OF 

CHIPS CONTRACTION COEFFICIENT CD 

BY 

MARIUS MILEA ∗ and MIRCEA COZMÎNCĂ 



”Gheorghe Asachi ” Technical University of Iaşi, 


Abstract. The assessment of plastic deformation of chips can be done based 

on theoretical and experimental considerations. This paper presents the two ways 

of assessing the chips contraction coefficient, namely the theoretical and 

experimental methods. The results obtained from both methods will be used to 

develop a mathematical model for chips contraction coefficient. 

Key words: chips contraction coefficient, theoretical methods, experimental 

methods. 


The chips contraction coefficient represents the capacity of plastic 

deformation of different metals during the cutting process. The assessment of 

plastic deformation is based on both theoretical and experimental 

considerations, which are used to develop a mathematical model for Cd. In this 

mathematical model, Cd is dependent on the parameters of the cutting process. 

2. The Assessment of Cd 

2.1. Analytical and Experimental Methods 

2.1.1. Analytical methods consist in the mathematical calculus of the 

elements that characterize the plastic deformation of the cut metal in various 

conditions. In order for this method to be applied it is necessary to know the 

∗ Corresponding author: e-mail: milea_marius@yahoo.com

118 Marius Milea and Mircea Cozmîncă 

yield curves, the structural characteristics of the materials and the distribution of 

the efforts in the deformation zone. The Ernst and Merchant approach 

introduces the concept of the single shear plane and the angle it makes with the 

surface generated referred to as the shear angle. It has become a classic 

approach in metal cutting and has been applied in analyzing the cutting of 

different materials even when shearing cannot occur at all. Zorev suggests a 

model for the cutting of ductile materials in agreement with the theory of 

plasticity (Astakhov, 2003). Summarizing the analytical models, it can be said 

that each cutting approach or model reflects a particular aspect of metal cutting 

practice. No model can cover all various cutting conditions that can be found 

during the real process. 

2.1.2. Experimental methods directly measure the elements of plastic 

deformation. V. Astakhov suggests two experimental methods for the 

determination of the chip contraction coefficient. The simplest method is to 

measure the chip thickness and calculate Cd as 

t2 

Cd 

= , (1) 

t1 

where t2 is the chip thickness and t1 is the uncut chip thickness. This method 

cannot be always used because of the chip saw-toothed free surface or its 

smallness. The second method is the weighing of the chips. After determining 

the length L, the width dw and the weight Gch , the chip thickness is calculated 

as 

t 

2 

Gch 

= , (2) 

d Lρ g 

w1 w 

where ρw is the density of the work material and g=9.81m/s 2 is the gravity 

constant. These methods, either theoretical or experimental, can lead to errors in 

assessing the chips contraction coefficient. To eliminate these errors, a 

theoretical-experimental model of assessing Cd is suggested, based on both 

theoretical considerations, but mostly on experimental results. 

2.2. The Development of a New Model for the Assessment of Cd 

As we know, theoretical equations for the chips contraction coefficient do 

not include all the parameters with significant influence, reason why we suggest 

a new model which takes into consideration the sense and the level of influence 

of each parameter, that is HB , v, s, K, γ and λ.. 

A series of authors suggested the model below for the assessment of Cd, 

which includes the six parameters mentioned above, 

n n 

δ 5ω 6 

Cd = C . (3) 

n1 n2 n3 

n 

HB v s K 4


The influence levels n1…n6 will be determined experimentally. The 

complementary angles δ=90−γ and ω=90−λ are used in order to avoid the 

negative and null values of γ, λ angles. The authors of this model also suggested 

the use of experimental diagrams in double-logarithmic coordinates to assess 

the influence levels n1,…,n6. 

After determining the influence levels, a second step is necessary, namely 

the determination of the C constant with Eq. (4), where v0, s0, K0, δ0, ω0 and HB0 

are the values for the independent variables that provide the same value for Cdo 

(experimental data tabels or diagrams in double-logarithmic coordinates can be 

used) 

d 0 

n1 0 

n2 0 

n3 

0 

n4 

0 

n5 n6 

δ0 ω0 

C HB v δ K 

C = . (4) 

After determining both the influence levels and the C constant, other 

experimental data are necessary to take into consideration the interdependencies 

and the values obtained so far may need correction. Thus, after assessing the 

interdependencies, Eq. (3) may become 

n5m n6m 

δ ω 

Cd = Cm . 

(5) 

n1m n2m n3m n4m 

HB v s K 

The experimental data, the corrected values of C constant and of the 

influence levels will be used to verify Eq. (5). The values obtained with Eq. (5) 

will be compared with experimental data. Depending on the similarity between 

the two sets of values, a new model for assessing Cd can be obtained by 

correcting Eq. (5) and introducing an intermediary parameter with 

n n n n 

1m 2m 3m 4m 

HB v s K 

Cm = C 

. 

(6) 

d exp n5m n6m 

δ ω 

In Eq. (6), the parameters HB, v, s, K, ω, δ have the values for which 

Cdexp. has been obtained experimentally. 

Finally, Eq. (7) is obtained, in which nm, nv, ns, nK, nλ, nγ represent the 

influence levels of the cut material, the cutting speed, the feed and the 

constructive angles of the tool, γ, λ, K. 

nγ nλ 

δ ω 

Cd = Cm . (7) 

nm nv ns n 

HB v s K K 

The calculus model (7) suggested for the assessment of Cd must be 

adjusted according to the specific parameters of each cutting process and cut 

material (ductile or less ductile-fragile).

120 Marius Milea and Mircea Cozmîncă 


1. The new suggested model is useful for the development of a new 

model for the assessment of the cutting forces depending on the chips 

contraction coefficient and is complementary to other existing models. 

2. This model includes six parameters of paramount significance, namely, 

the cut material hardness, the cutting speed, the cutting feed and the 

constructive angles of the tool. 

3. In the future, the values obtained with this new model of assessing Cd 

could be used for classifying the metallic materials from the point of view of 

their cutting workability. 

REFERENCES 

Astakhov V., Shvets, S.V, The assessment of Plastic Deformation in Metal Cutting. 

Journal of Materials Processing Technology, 146(2), 193-202 (2004).. 

Croitoru I., Segal R., Cozmîncă M.,Model for Predicting Cutting Forces. Bul. Inst. 

Polit. Iaşi, , XLIV(XLVIII), Supliment I, s. V, 189-193 (1998). 

Cozmîncă I., Cozmîncă M., Ibănescu R., About a New Method for Cutting Forces 

Evaluation. Bul. Inst. Polit. Iaşi, , LIV(LX), 2a, s. V, 125-129 (2010). 

Cozmîncă I., Ibănescu R., Rădulescu M., Ungureanu C., Voicu C., , Experimental 

Results Regarding the Chips Contraction at Steel Turning. Bul. Inst. Polit. Iaşi, 

LVI (LX), 1, s. V, 13-18 (2010). 

SINTEZĂ ASUPRA EVALUĂRII COEFICIENTULUI 

DE DEFORMARE PLASTICĂ CD 

(Rezumat) 

Se prezintă pe scurt cele mai importante modele teoretice şi experimentale de 

evaluare a coeficientului de deformare plastică Cd. Modelul propus de cercetătorii în 

domeniu include şase factori cu influenŃă semnificativă, respectiv, viteza principală de 

aşchiere, avansul de aşchiere, duritatea materialului aşchiat şi unghiurile constructive 

ale sculei. S-a încercat dezvoltarea acestui model, evaluându-se interdependenŃele dintre 

aceşti parametri şi obŃinându-se o nouă relaŃie care va trebui verificată experimental.


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INFLUENCE OF CHORD VARIATION ON THE 

PERFORMANCE OF A KINETIC MINITURBINE 

BY 

EUGEN-VLAD NĂSTASE ∗ and DORU CĂLĂRAŞU 

“Gheorghe Asachi” Technical University of Iaşi, 


Received: July 10 2011 


Abstract. In this paper we present an analisys of miniturbine performance 

when the value of chord is variable. This study is conducted for a given situation 

in which is known the average flow speed and number and the geometry of blades 

for an kinetic miniturbine. 

Key words: water current turbines, kinetic miniturbine. 


There are basically two methods of extracting energy from water. The 

conventional method is to place a barrage across an estuary with a large tidal 

range to create a static head or pressure difference, and operate a low head 

hydro-electric power plant with intermittent, reversing flow. The less wellknown 

method of extracting energy from tidal and other flows is to convert the 

kinetic energy of moving water directly to mechanical shaft power without 

otherwise interrupting the natural flow. The aim of this paper is to study 

influence of chord variation on the performance of a kinetic miniturbine. 

2. Mathematical Model 

The performance of the miniturbine (Zanette et al., 2007), (Jureczko et 

al., 2005), (Lanzafame & Messina, 2007) is theoretically predicted by 

considering the coefficient of power defined with equation: 

∗ Corresponding author: e-mail: nastase_eugenvlad@yahoo.com

122 Eugen-Vlad Năstase and Doru Călăra u 

3 

K 

p 

dPT 

= , 

3 

(1) 

ρπrdrV∞ where, ρπrdrV∞ is the power available from a stream of water, and dP T is the 

power extracted by miniturbine. 

Fig. 1 – Blade cross-section at radius r. 

Can be estimated (from Fig. 1) the tangential component 

dT = dR sin δ − dR cosδ 

, (2) 

u P R 

1 2 

P 2 P 

dR = ρC W cdr (3) 

1 2 

R 2 R 

dR = ρC W cdr . (4) 

For evaluate the torque on the miniturbine shaft we are using the relationship: 

ρ 2 sin( δ − ε) 

dM = nrdTu = n rcd rW CP 

. 

(5) 

2 cos ε 

Finally the coefficient of power is 

2 2 2 ( ∞ ) 

dP ωdM cnω V + ω r sin( δ − ε) 

K = = = C . (6) 

cos ε 

T 

p 3 

ρπrdrV∞ 3 

ρπrdrV∞ 3 

2πV∞ 

P 

In the above equations: ρ is the density of water, V ∞ is the free stream 

velocity of the current, n is the number of blades, c is the chord, CP is lift 

coefficient, CR is drag coefficient. The consequence of this relationship is that 

power and hence energy capture are highly sensitive to chord, number of blades, 

velocity, lift coefficient.


3. Results 

Calculations were made for two types of blades: type swept (Fig. 4a), 

with chord constant along the blade and type sword (Fig. 4b), with chord 

variable along the blade. 

Fig. 2 – Variation of lift and drag coefficient. 

Fig. 3 – Variation of finess and power coefficient. 

Fig.4 – Efficiency for variation of chord. 

Legend of representation: 

Value of chord is constant along the blade: c=0,034(m), c=0,044 (m), 

and c=0,054 (m). 

Value of chord is variable along the blade: c=0,034 (m), c= 

0,044(m), 

and c=0,054 (m).

124 Eugen-Vlad Năstase and Doru Călăra u 


1. The analysis performed is found that better performance is obtained for 

the case of variable chord along the profile (Fig. 4b). 

4). 

2. For the two cases considered the best value for chord is c=0.034m (Fig. 

3. This type of kinetic miniturbine have the advantage of not requiring 

storage dam, is simple construction and maintenance costs are relatively small. 

REFERENCES 

Jureczko M., Pawlak M., Mezyk A., Optimisation of Wind Turbine Blades. Journal of 

Materials Processing Technology, 463-471 (2005). 

Lanzafame R., Messina M., Fluid Dynamics Wind Turbine Design: Critical Analysis. 

Optimization and Application of BEM Theory. Renewable Energy, 32, 2291-2305 

(2007). 

Zanette J., Imbault D., Tourabi A., Fluid-structure Interaction and Design of Water 

Current Turbines. Grenoble, France, 2007. 

INFLUENłA VARIAłIEI CORZII ASUPRA 

PERFORMANłELOR UNEI MINITURBINE CINETICE 

(Rezumat) 

În această lucrare se prezintă analiza performanŃelor unei miniturbine pentru 

două situaŃii. Primul caz este cel al unei miniturbine cu pale care au coarda constantă în 

lungul anvergurii. Al doilea caz este cel al miniturbinei cu pale având coarda variabilă 

în lungul anvergurii. Acest studiu este realizat pentru o locaŃie cunoscută, când se 

cunoaşte viteza medie de curgere a apei, geometria şi numărul de pale.


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THEORETICAL RESEARCH REGARDING THE BLADES 

NUMBER INFLUENCE OF THE MINITURBINE EFFICIENCY 

BY 

EUGEN-VLAD NĂSTASE ∗ and DORU CĂLĂRAŞU 





Abstract. The aim of the paper is to is to achieve a theoretical study on the 

influence of the number of blades on the performance of miniturbine. The study 

is conducted to a location with average speed of water flow known. 

Key words: renewable energy, kinetic turbine, energy conversion. 


Renewable energy technologies offer the promise of non-polluting 

alternatives to fossil and nuclear-fueled power plants to meet growing demand 

for electrical energy (Batten et al., 2006), (Lago et al., 2010), (***, 2011). 

Water current turbines are defined as systems that convert hydro kinetic energy 

from flowing waters into electricity, mechanical power, or other forms of 

energy. This is a free flowing turbine in which the water runoffs are given by 

the difference in altitude of the river in the stretch considered. The same 

principle is applied in energy conversion systems used in ocean currents and 

tide motors (Liu et al., 2011). The great benefit in which this type of 

arrangement is that there is no need to build dams or dikes to supply water to 

the turbine and consequently, results in a low environmental impact. 

2. General Considerations 

For efficiency analysis of kinetic miniturbine we can use the following 

energetic parameter (Batten et al., 2006), (Myers & Bahaj, 2006): 

∗ Corresponding author: e-mail: nastase_eugenvlad@yahoo.com

126 Eugen-Vlad Năstase and Doru Călăraşu 

8 πr(1 − k) tan δ sin δ 

ERR = 

, 

⎛ µ ( k + 1) ⎞ 

( k + 1) ⎜1+ ⎟ 

⎝ λ( h + 1) ⎠ 

where: r − radius of a cross section, k,h-coeficients of induced velocity, δ − 

angle between flow velocity direction and chord, λ − tip speed ratio, µ − 

solidity. On the other hand, for ERR we have 

C pnc 

ERR = . 

r 

Using Eq (2) we can represent the dependence between number of blades and 

lift coefficient, what we can see in Fig. 1. 

Fig.1 – Dependence between lift coefficient and blades number. 

For theoretical research, after analyzing the dependence between the 

number of blades and lift coefficient (Fig. 1), we consider three situation: 

miniturbine with three, four and five blades (Fig. 2). 

Fig.2 – Models of kinetic miniturbine. 

For these three cases we calculated and represented variation along the 

blade for the following parameters: lift coefficient (Fig. 3), tangential force (Fig. 

4), torque to the turbine shaft (Fig. 5), and power coefficient (Fig. 6). 

(1) 

(2)


3. Results of Research 

Legend of representation: miniturbine with three blades; 

miniturbine with four blades; miniturbine with five blades; 

Fig. 3 – Lift coefficient variation. 

Fig. 4 – Tangential force. 

Fig. 5 – Torque variation. 

Fig. 6 – Power coefficient along the blade.

128 Eugen-Vlad Năstase and Doru Călăraşu 


1. Analyzing the results it is found that for this situation is more efficient 

turbine with three blades. 

2. Hidrokinetic energy offers the promise of non-polluting alternatives to 

fossil and nuclear-fueled power plants. 

REFERENCES 

Batten W. M. J., Bahaj A. S., Molland A. F., Chaplin J. R., Hydrodynamics of Marine 

Current Turbines. Renewable Energy, 31, 249-256 (2006). 

Lago L. I., Ponta F. L., Chen L., Advances and Trends in Hydrokinetic Turbine Systems. 

Energy for Sustainable Development, 14, 287-296 (2010). 

Liu H. W., Ma S., Li W., Gu H. G., Lin Y. G., Sun X. J., A Review on the Development 

of Tidal Current Energy in China. Renewable and Sustainable Energy Reviews, 

15, 1141-1146 (2011). 

Myers L., Bahaj A. S., Power Output performance Characteristics of a Horizontal Axis 

Marine Current Turbine. Renewable Energy, 31, 197-208 (2006); http://www. 

esru.strath.ac.uk/EandE/Web_sites/05-06/marine_renewables/home/ 

projsummary. htm (available at 18.03.2011). 

CERCETĂRI TEORETICE PRIVIND INFLUENłA NUMĂRULUI DE PALE 

ASUPRA EFICIENłEI UNEI MINITURBINE 

(Rezumat) 

Lucrarea are drept scop realizarea unor studii teoretice privind influnŃa 

numărului de pale asupra eficienŃei unei miniturbine. Miniturbina studiată este 

proiectată pentru încercări de laborator. Ca urmare sunt cunoscute dimensiunile 

rotorului, viteza medie de curgere a apei în canal şi geometria profilului palei. În aceste 

condiŃii se constantă că performanŃele cele mai bune se obŃin în cazul miniturbinei cu 

trei pale.


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ABOUT MODELLING THE MOVEMENT CONTROL OF 

THE ELECTROMECHANIC LINEAR ACTUATOR 

BY 

VASILE NĂSUI ∗ 

North University, Baia Mare, 

Department of Engineering System and Technological Management 


Accepted for publication: June 23, 2011 

Abstract. This article refers to the conception and the practical 

achievement of some systems of linear movement of electro-mechanical 

actuator type used to position precisely of some pre-working equipment. 

These systems have to own dynamic characteristic high precision and high 

viability. The important functions of the actuator is the movement control 

achieved by an appropriate electronic. The electro-mechanic acting can be 

developed by using appropriate control destinations such as the loading 

equipment the actuators have to fulfill strict and stable criteria in any 

situation. The actuators having multiple applications both industrial and of 

these products. 

Key words: fuzzy logistic, linear actuator, controller, advanced 

control. 


The actuator have a complex structure, the mechanical part being 

composed of certain elements which ensure a high cinematic and dynamic 

precision, and the commanding part being represented by a computer-led 

system, based on appropriate software. Like conventional linear actuator it 

also comprised a motor rotation and a motion screw mechanism, with 

∗ e-mail: nasui@ubm.ro

130 Vasile Năsui 

balls or rollers. One type is the electromechanical actuator, which converts 

the torque of an electric rotary motor into linear mechanical thrust. The 

motor rotates the drive screw by a synchronous function is to provide 

thrust and positioning in machines used for production or testing 

(Borangiu, 2003). 

The electromechanical linear actuators are designed to provide 

precision, efficiency, accuracy, and repeatability in effecting and 

controlling movement. A typical linear electro-mechanic actuators is 

sketched in Fig. 1 (Năsui, 2006). 

Controller 

Fig. 1 – The overall scheme of linear actuator system. 

A position at some point along the screw is commanded by the user 

and the motor turns the screw until the nut reaches that position. These 

provide the optimal solution for the timing belt drive, worm gear drive, or 

a coupling direct inline drive, connected from the motor shaft to the screw. 

An actuator's construction of quantitative and qualitative mechanical 

transmissions due to a wide range of usage, high efficiency, cinematic, 

dynamic capacities and high precision. 

The acting directions are shown in order to improve the parameters 

such as increasing speeds and portent capacity which generate the 

necessity of the analysis and the solution to some problems concerning the 

dynamics of the systems (Nasui, 2006). 

The remainder of the paper is organized as follows. In section two 

the capabilities of the electromechanical linear actuators are presented. A 

short description of model geometry is given in section three. Some results 

of the virtual simulation are presented in section four. 

Finally, in the last section the main conclusions from this study are 

drawn and the perspectives for future research are outlined. Fully 

reviewing your application can prevent mistakes, ensuring optimal system 

performance. 

2. The Control Algorithm 

The main task is the exact position towards the control surface and 

its quick matching of the two factors simultaneously which is difficult to


achiever in practice. The compromise between speed and precision has to 

be established from the beginning by applying a control algorithm. 

The working regime is non-linear and non stationary owing to the 

cutting forces. The acting of the actuators can not be represented as a 

linear element and the classical analysis methods can not be used. Besides 

this there are safely and viability aspect to take into account. 

The control algorithm has to prevent the situations where the engine 

is the critical functioning area in extreme dynamic conditions. The acting 

dynamic is very important. Fast swifts where the maximum working limits 

are exceeded should to avoid. Quick changes in the load from the dynamic 

point of view show similar modifications of parameters at the PD 

regulator for this programmed. Because of the many non-linearity in the 

system, the programmed should have variable parameters which are hard 

to achieve. 

The algorithm that ensures the dynamic proprieties for the regulator 

which do not reach the saturation point can protect the mechanism 

structure and the durability of the acting engine. In this situation it is 

necessary for the experts produce heuristic regulators which use the 

dependences between the engine and the acting system. These heuristic 

regulations have very good static and dynamic properties according to the 

saturation rate stage compensators and algorithms are necessary for stages 

over the saturation rate (Borangiu, 2003), (Năsui, 2006). 

Additionally this process should be as fast as possible. Obtaining a 

good quality of these two factors simultaneously is difficult in particular 

case and practically impossible in the general case. Compromising 

between precision and speed is necessary. The sufficient precision has to 

be established in the beginning and should be applied control algorithm 

maintaining it. The electromechanical actuator without any load is a 

nonlinear element (Nasui, 2006). 

Actuator cannot be represented as a linear element and we cannot 

use classical methods of analyses for it. There are also aspects of 

reliability and safety. A various and heavy load of the actuator, frequent 

changes of the control signal (especially direction) can cause quick 

amortization of electrical motor and the power driver. Control algorithm 

should prevent situations where the motor is load to much for a long time. 

The control is based on a microprocessor, converter and power 

output block. The reversed link of control is achieved by potentiometer 

supplied by the control circuit present in the electro-mechanic unit by the


actuator type. A typical the closed loop position control system is sketched 

in Fig. 2. 

Σ Kp Kv Σ 

Σ Actuator ∫ 

∫ Kv 

Fig. 2 – The closed loop position control systems. 

3. Programmable Motion Control Systems 

An actuator's function is to provide thrust and positioning in 

machines used for production or testing. The motion control system’s 

purpose is to control any one, or combination, of the following 

parameters: position, velocity, acceleration, torque (Nasui, 2006). 

Many motion control systems are integrated into a larger system. 

Various computer-based devices, such as programmable controllers, 

stand-alone industrial computers, or mainframe computers serve to link 

and coordinate the motion control function with other functions. Thus, a 

more integrated motion control system would appear as shown below: the 

assembly of the process of developing new products, covering the 

conception aspects, manufacture and the link between them. The essential 

component of many automatic control systems is actuator. The application 

of a specific command causes a corresponding signal at the input through 

action input transducer. The result is an unbounded increase in controlled 

variable and loss of control by the command source. 

The general form of the block diagram of a feedback control system 

is shown in the Fig. 3. In order to develop and optimization pattern 

according to very strict engineering requirement it is necessary the 

introduction of a number of performance criterion and the formulation of 

some appropriate objective functions (Năsui, 2006). 

This results a characteristic of functioning specific to each measure 

of translation unit according to the dimension and the step of the moving 

screw with which it is equipped. These restraints refer the achievement to 

the possibility of achieving technical performances referring to the 

parameters of functional geometrical precision.

Computer 


position 

directio 

speed 

Controller 

Fig. 3 – Established actuator block scheme. 

These are essential in the case of using positioning systems from 

robots, machine-tools computer controlled, specific to transitory 

conditions with frequent speeding and braking, starting and stopping at a. 

fixed point. 

They have to be reliable and work stably in every situation. The 

actuator controller should be reliable itself and independent from other 

equipment failures. The electromechanical unit is controlled throughout a 

power output block. The controller only sets power control and direction 

lines to a required state and the power output realizes powering of motor. 


power control 

direction 

load 

mechanical 

power 

Electro 

mechanism 

1. The result of the research we can value immediately any system 

of linear movement because it has as a basis the newest techniques of 

modeling in the field. These have as an objective the developing of new 

system of electro-mechanic linear actuator type as well as perfecting the 

existent ones in a very short time and in highly energetic and economical 

conditions. 

2. This research is part of the modern preoccupations regarding the 

improvement of new systems of linear acting and of numerical modeling 

using algorithms and programs of numerical computation and virtual 

instrumentation. The sequence of special performances obtained through 

this regulatory system to the actuators makes possible their application to 

the machine tools with numerical control and especially at those of the last 

generation with parallel kinematics as well as the robot industry. 

3. Control algorithm permit to deflect surface with maximum 

power in wide range. Unfortunately speed is limited by electromechanical 

construction and over saturation we can only compensate phase lag.


Amplitude of signal is suppressed. Work of controller has been checked 

in many possible situations 

4. The control algorithm allows the positioning regulation at 

maximum powers and very large limits. The movement speed is limited 

by the mechanic construction and cannot compensate wholly the 

saturation stage/point. 

5. The conception and its manufacture assisted on the computer has 

as application field the assembly of the process of developing new 

products, covering the conception aspects, manufacture and the link 

between them. 

6. The laboratory test done a data acquiring stand shoved that the 

control system of the movement achieved by the controller can control the 

positioning precision with established static precisions. Tests have been 

done in the maximum admitted loading conditions and minimum and 

maximum extreme working extreme speeds. 

Acknowledgements. The author makes the best of the researches done 

within the grant regarding the development of the actuators within the flexible 

systems of reworking in the laboratory of The Basis of Experimental Research of 

the North University of Baia Mare, Engineering Faculty. The author to thanks the 

support and contribution of Eugen PAY from UNBM in the definition of 

requirements and selection of design options of this electro-mechanic actuator 

linear. 

REFERENCES 

Banks J. et al., Discrete Event System Simulation. Prentice Hall Inc., SUA, 2001. 

Borangiu Th., Advanced Robot Motion Control. Ed. Academiei Romane, 2003. 

Ispas C., Predencea N., Ghionea A. et al., Maşini-unelte. Ed. Tehnică, Bucureşti, 

1998. 

Maties V., Mandru D., Tatar O. Actuatori in mecatronică. Ed. Mediamira, Cluj- 

Napoca, 2000. 

Mohora C., Cotet E., Patrascu G., Simularea sistemelor de producŃie. Ed. 

Academiei Romane, 2001. 

Montgomery D.C., Design of Analysis of Experiments, 4 th Ed., John Wiley & 

Sons, New-York, 1996. 

Năsui V., (2006), Actuatori liniari electromecanici. Ed. Risoprint, Cluj Napoca, 

1996. 

Nãsui V., Stand de proba pentru mecanisme liniare. Brevet de Inventie RO 

122562 B1. 

Popa A., Controlul digital al sistemelor mecatronice. Ed. Orizonturi Universitare, 

Timişoara, 2002.


Olaru A., Dynamic of the industrial robots. Ed. Bren, Bucureşti, 2001. 

Zetu D. & Carata E., Sisteme flexibile de fabricaŃie. Ed. Junimea, Iaşi, 1998. 

*** Data Acquisition Devices type ADUC 112, Analog Device. 

MODELAREA CONTROLULUI MIŞCĂRII LA 

ACTUATORII LINIARI ELECTROMECANICI 

(Rezumat) 

Lucrarea are ca scop realizarea unei noi abordări a concepŃiei şi realizării practice 

a unor sisteme de mişcare liniară de tip actuator liniar electromecanic, utilizate pentru 

poziŃionarea precisă a unor echipamente de prelucrare. Aceste sisteme trebuie să aibă 

caracteristici dinamice ridicate, precizie si fiabilitate înaltă. Se foloseşte un nou sistem de 

transformare a mişcării de rotaŃie în miscare de translaŃie şi invers, care facilitează 

achiziŃionarea datelor de control a mişcării. Problema pusă se referă la o reconsiderare a 

modalităŃilor de control a mişcării configurând noi aplicaŃii experimentale. FuncŃia importantă 

la actuatoarele liniare electromecanice, în special cele industriale este controlul 

mişcării realizată de către un echipament electronic, care trebuie să îndeplineascăa criterii 

stricte şi stabile în orice situaŃie.


Publicat de 



SecŃia 


FINITE ELEMENT MODELING OF A KNEE AND HIP 

REHABILITATION EQUIPMENT 

BY 

IOANA PETRE ∗1 , TUDOR DEACONESCU 1 , 

ANDREA DEACONESCU 1 and DAN PETRE 2 

“Transilvania” University of Braşov, 

1 Department of Economic Engineering and Production Systems 

2 Department of Materials Engineering and Welding 



Abstract. In this paper it is presented the finite element modeling of 

rehabilitation equipment for the knee and hip affections using CATIA V5R19 

Software. It is presented the equipment to be analyzed, the material, type of 

analysis, the constraints and the loads applied. After analyze is made, the soft 

shows the results, which are interpreted. The results of the analysis, Von Missed 

stresses and displacements show the resistance of the equipment under loadings 

conditions. 

Key words: rehabilitation equipment, Finite Element Modeling, CATIA. 


The paper presents finite element modeling of an ankle and knee 

rehabilitation equipment. The rehabilitation equipment proposed for the analysis 

allows performing isokinetic exercises in order to offer continue passive motion 

for recovery of the patients with affections of the hip and knee joints (Petre & 

Deaconescu, 2009). 

Finite element modeling (FEM) involves several steps, shown in Fig. 1, 

which are followed for acquiring the results presented below. Preprocessing 

includes several steps: geometric domain modeling, material modeling, 

generation finite element structure, constraints modeling, loads modeling, 

∗ Corresponding author: e-mail: babes_ioana@yahoo.com

138 Ioana Petre et al. 

checking finite element model. Post processing includes: viewing and studying 

results and optimization the model. 

PREPROCESSING 

SOLVING FINITE 

ELEMENT 

MODEL 

Fig.1 – Finite element modeling steps. 

Those steps are followed in the analysis presented in this article. The 

model is designed in CATIA V5R19 and the structural analysis of rehabilitation 

equipment is realized using CATIA Generative Structural Analysis workbench. 

The results of the analysis, Von Missed stresses and displacements show the 

resistance of the equipment under loadings conditions. 

2. Finite Element Analysis of Rehabilitation Equipment 

Equipment drawing is realized by the help of the following modules: 

Sketcher module – for describing the plane profiles which are the base for 

tridimensional elements generation, Pad module – for describing the 

tridimensional elements, and Assembly Design module – for describing the 

assemblies and subassemblies. 

Designed equipment is considered to have a material (aluminum) with the 

following physical properties, which are important during the analysis: Young 

modulus (7 × 10 10 N/m 2 ), Poisson ratio (0.346), density (2710 kg/m 3 ), the 

coefficient of thermal expansion (2.36 ×10 -5 K), Yield strength (9.5 × 10 7 N/m 2 ). 

Finite element analysis of knee and hip rehabilitation equipment was 

performed using software CATIA V5R19, with Generation Structural Analysis 

module. The type of analysis is static analysis (Static Case) which is performed 

considering some constraints and independent of time loads. 

Generating finite element structure involves the model meshing and 

introduction the finite element properties. Meshing model is achieved through a 

network, composed of nodes and elements. Table 1 shows the main features of 

the finite element model. 

Table 1 

Principal characteristics of the finite element model 

Entity Finite elements Description 

Nodes 53283 

Elements 26388 Tetrahedron 

POST 

PROCESSING 

The constraints are applied between adjacent elements and between 

elements and the fixed part. Applied restrictions are: Clamp restriction, applied 

to the base; Slider restriction, applied to the slider. 

On the surface of the bars is applied a distributed force of 300 N value 

applied as follows: 40% of the force value on the bar corresponding to the calf


and 60% of the force value on the bar corresponding to the thigh, oriented 

perpendicular to the surface, in the Y axis direction. Other forces applied: 

pulling force of the muscle − 350N, gravity force − 10N. 

The symbols of the mesh, of constraints and restrictions, and of the 

applied forces are shown on the analyzed model, as seen in Fig. 2. 

Fig. 2 – Equipment to be analyzed. 

After verifying the finite element model, the next stage is the analysis 

phase. Solving the model is carried out automatically by the software, with the 

Compute command. Post-processing stage involves visualization of the results 

and then optimization of the model. 

The following figures present the results of static analysis. 

Fig. 3 – Deformation. 

After analyzing the proposed equipment for operation in real time, it can 

be observed that there is no deformation (Fig. 3), so the equipment does not 

deform under applied forces. Fig. 4 allows viewing equivalent von Mises stress 

fields (Von Mises Stresses Visualizing). 

The highest tensions can be observed in the central part of the bars, in the 

rod clevises of frame gripping, in the clamping pins of the slipping rolls and in 

the slider spring.


Equivalent tensions von Misses theory is determined by the maximum 

breaking strain energy, so the equivalent tension is determined by the relation: 

1 

2 2 2 2 2 2 

σe = ( σx − σ y ) + ( σ y − σz ) + ( σz − σx ) + 6( τxy + τ yz + τxz 

) , (1) 

2 

where σ is normal stresses and τ tangential stresses (Lateş, 2002). 

In the image is presented the color palette that accompanies the result. 

The lowest tension values are at the bottom of the palette, and the maximum 

values at the top of it, but the dialog box contains explicit values, as follows: 

Min: 6.64759e-005 N/m 2 and Max: 5.69198e+006 N/m 2 . 

Fig. 4 – Von Mises stress. 

Since the yield strength of the material is 9.5 × 10 7 N / m 2 , it can be 

concluded that the model will resist at the applied distributed force. 

Displacements of the analyzed model are presented in Fig. 5. 

The maximum displacement of the nodes was found to be 37mm. This 

maximum is located on the slider and on the cross-bar from the knee, and is 

shown in red. 

It is necessary for the slider to realize a larger displacement, so the 

equipment will be actuated with a pneumatic muscle, connected by the slider 

with a pulleys mechanism. The pulleys will double the stroke realized by the 

pneumatic muscle, at the needed value. 

Fig. 5 – Equipment translational displacement.



1. Finite element analysis is a powerful tool that allows engineers to 

quickly analyze and refine a design. It can be applied to problems involving 

vibrations, heat transfer, fluid flow, and many other areas (Petre et.al, 2011). 

2. The analysis above shows the behavior of the proposed equipment 

under the applied forces. The results highlighted that the Von Misses stress and 

displacement contours of the equipment are safe for the imposed loadings. 

3. This research will be continued with a more in depth analysis and 

testing of the real equipment in laboratory. 




POSDRU/88/1.5/S/59321. 

REFERENCES 

Petre I. Deaconescu T., Isokinetic Equipment Designed For Therapeutic Exercises. 

Proceedings of International Conference on Economic Engineering and 

Manufacturing Systems, ICEEMS 2009, Braşov, 2009. 

Lateş M. T., Metoda Elementelor Finite. AplicaŃii. Ed. Univ. Transilvania din Braşov, 

2008. 

Petre I., Deaconescu T., Deaconescu A., Petre D., Finite Element Analysis of Pneumatic 

Muscle. in 15th International Conference on Modern Technologies, Quality and 

Innovation - MODTECH 2011: New Face of TMCR, Vadul lui Vodă, Chişinău, 

2011, pp. 861-864. 

*** Generative Part Structural Analysis. Available at: http://www.catia.com.pl/tutorial/ 

generative_part_structural_analysis.pdf; accessed 12.07. 2011. 

*** http://catiadoc.free.fr/pdf/EN-Dassault-Systems_Generative_ Assembly_ Structural 

_ Analysis.pdf; Accessed 12.07.2011. 

MODELAREA CU ELEMENT FINIT A UNUI 

ECHIPAMENT DE REABILITARE 

(Rezumat) 

Lucrarea prezintă modelarea cu element finit a unui echipament de reabilitare a 

afecŃiunilor genunchiului şi gleznei. Echipamentul propus spre analiză permite efectuarea 

unor exerciŃii de mişcare pasivă continuă, fiind acŃionat cu un muşchi pneumatic. 

Modelarea cu elemente finite (MEF) presupune parcurgerea etapelor: preprocesare, 

rezolvarea modelului cu elemente finite şi postprocesarea. Preprocesarea cuprinde 

mai multe etape: modelarea domeniului geometric, modelarea materialului, generarea 

structurii cu elemente finite, modelarea constrângerilor, modelarea încărcărilor, 

verificarea modelului cu elemente finite. Postprocesarea cuprinde: vizualizarea şi 

studiul rezultatelor şi optimizarea modelului.


În modelarea echipamentului de reabilitare sunt urmărite aceste etape, prezentate 

mai sus. Astfel, în etapa de preprocesare are loc realizarea echipamentului, stabilirea 

materialului, a constrângerilor, a încărcărilor şi verificarea modelului, apoi se realizează 

analiza modelului cu element finit, iar în ultima etapă are loc vizualizarea şi interpretarea 

rezultatelor şi optimizarea modelului. 

Softul folosit este CATIA V5R19, cu modulele: Sketcher, Pad şi Assembly 

Design - pentru realizarea modelului, precum şi Generation Structural Analysis pentru 

analizarea acestuia. 

În urma analizei echipamentului propus se observă că nu există deformare, deci, 

cadrul echipamentului nu se deformează sub acŃiunea forŃelor aplicate. Tensiunile cele 

mai ridicate se observă în partea centrală a barelor, în furcile de prindere a barelor de 

cadru, ştifturile de prindere a rolelor de alunecare şi în arcul ce Ńine culisorul. Având in 

vedere ca valoarea maximă a tensiunilor von Misses este de 5.69198×10 6 N/m 2 , iar 

rezistenta admisibila a materialului este de 9,5×10 7 N/m 2 , se poate trage concluzia că 

modelul piesei va rezista forŃei distribuite aplicate. Deplasarea maximă, pentru condiŃiile 

impuse este de 37mm, şi este efectuată de culisor şi de bara transversală corespunzătoare 

cuplei genunchiului. 

Analiza prezentată prezintă comportamentul echipamentului de reabilitare sub 

acŃiunea unor forŃe impuse, iar rezultatele dovedesc rezistenŃa acestuia şi siguranŃa 

oferită în condiŃiile impuse.


Publicat de 



SecŃia 


SOME CONSIDERATIONS REGARDING 

PNEUMATIC MUSCLE VOLUME 

BY 

IOANA PETRE ∗1 , TUDOR DEACONESCU 1 , 

ANDREA DEACONESCU 1 and DAN PETRE 2 

“Transilvania” University, Braşov, 

1 Department of Economic Engineering and Production Systems 

2 Department of Materials Engineering and Welding 



Abstract. A pneumatic artificial muscle is composed of an inner tube of 

variable length, made of a flexible material, typically neoprene. The tube is 

covered with a multilayer texture, made of nylon with strengthening and 

protecting role from the environment influences. Under the action of compressed 

air it increases its diameter and decreases its lengths. A pneumatic muscle is used 

successfully in different areas, especially in robotics. The paper presents some 

calculus models for volume of the pneumatic muscle, considered in relaxed and 

contracted state. 

Key words: pneumatic muscle, volume, relaxed state, contracted state. 


An artificial pneumatic muscle has in composition an inner membrane 

made of a flexible material covered by a helical net of inextensible fibers made 

of nylon with strengthening and protecting role from the environment 

influences. Under the action of compressed air it increases its diameter and 

decreases its lengths, working as a spring (Deaconescu, 2009). 

A pneumatic muscle is used successfully in different areas, because of 

their advantages, like the passive damping, good power-weight ratio and usage 

in rough environments (Hildebrandt et al., 2002). The relaxed state is 

∗ Corresponding author: e-mail: babes_ioana@yahoo.com


determined by the size of the original tube and by the protective membrane 

characteristics. Range contraction-expansion depends on the lower limit of the 

angle and therefore on the density and thickness of braid fibers. 

An interesting aspect in the study of behavior of a pneumatic muscle in 

functioning is its volume evolution under the action of compressed air. The 

following article presents a way of calculating the pneumatic muscle volume. 

In relaxed state, the muscle is considered as a cylinder, and the volume equation 

results from this. In contracted state, the muscle has a cylinder part, but at the 

ends, there are two spherical sections (Albienz et al., 2005). 

2. Pneumatic Muscle Volume Calculus 

The pneumatic muscle volume can be calculated in relaxed state or in 

contracted state. It starts from the assumptions that: 

In relaxed state, the muscle can be considered as a cylinder of radius r0= 

=d0/2. 

In contracted state, the muscle has cylinder form in the central area, with 

radius r = d/2 and length l, and at the ends has two junction areas as a spherical 

cap (Fig.1) (Dragan, 2007). 

Fig. 1 – Pneumatic muscle in relaxed and contracted state. 

In the relaxed state, muscle volume has the formula: 

2 

2 d0 

0 0 0 0 

V = πr l = π l 

4 

(1) 

Muscle volume in the contracted state can be calculated as follows 

(Dragan, 2007) 

V = V + V , (2) 

m 0 

2 

2 c 

d 

V0 = π ( l − 2h) 

, 

4 

(3) 

πh 2 2 2 

Vc = ( 3r0 + 3r 

+ h ) , 

6 

(4) 

where Vm – pneumatic muscle volume, Vc – spherical area volume, V0 – 

cylindrical volume. 

As known, a pneumatic artificial muscle is composed of an inner tube of 

variable length covered with a multilayer texture. The figure below illustrates 

how a wire wound on a flexible tube.


Fig. 2 – Correlation between muscle radius r, muscle length l, angle between 

fibers α0 and constant length of the s. 

Fig. 2 shows the correlation between the radius of the muscle, its length 

and fiber angle. Due to the fact that the length of the fibers is constant, it can be 

written the following relations: (Albienz et al., 2005) 

l cosα 

= , 

(5) 

l cosα 

From Eq. (6) results 

Using Eq. (5) 

0 0 

r sin α 

= . (6) 

r sin α 

0 0 

2 2 

1 − cos α 1 − cos α 

r = r = r 

0 0 

sin α 2 

0 1 − cos α 0 

l cos α 0 

cos α = 

l 

0 

. 

. (7) 

It is defined relative muscle contraction ε as 

z 

ε = 

l0 l0 − l 

= . 

l0 

(8) 

Hence contracted muscle length 

l = l0 

( 1 − ε ) . 

With those notations, Eq. (7) becomes 

(9) 

From Fig. 1 can be written: 

r = r 

0 2 

2 2 

1 − (1 − ε) cos α 

1 − cos 

2 2 

d d 0 

2 

h − = r − 

α 

0 

0 

(10) 

= 

4 4 

2 

r . 

0 

(11) 

From the above relations results 

2 2 2 2 

2 1 − (1 − ε) cos α0 2 1 − (1 − ε) cos α0 

0 2 0 0 

2 

1 − cos α0 1 − cos α0 

h = r − r = r 

− 1. 

Eq. (4) becomes 

2 2 2 2 

2 1 − (1 − ε) cos α ⎛ 

0 1 − (1 − ε) cos α ⎞ 

0 

V0 = πr ⎜ 0 l 

2 0(1 − ε) − 2r0 −1 

⎟. 

2 

1− cos α ⎜ 0 1− cos α ⎟ 

⎝ 0 ⎠ 

(12) 

(13)


Eq. (5) becomes 

⎧ ⎛ ⎞⎫ 

3 2 2 2 2 

πr0 1 − (1 − ε) cos α0 ⎪ 1 − (1 − ε) cos α0 

⎪ 

Vc 

= − 1 1 2 

3 2 ⎨ + ⎜ ⎟ 2 ⎬ , (14) 

1− cos α ⎜ 

0 1− cos α ⎟ 

⎪⎩ ⎝ 0 ⎠⎪⎭ 

where l0 is muscle length at rest, d0 and α0 is the diameter, respectively, the rake 

angle of braid fiber in relaxed state. 

The following relations based on Eq. (2) express the total volume of the 

contracted muscle, depending on the relative contraction. 

Vm ( ε) = V0 ( ε) + 2 Vc ( ε) 

(15) 

Entering into Eq. (2) the above equations results 

2 2 2 2 2 

π d ⎡ 0 1 − (1 − ε ) cos α ⎛ 

0 1 − (1 − ε ) cos α ⎞ 

0 

Vm ( ε ) = ⎢ 

⎜l 2 0(1 − ε ) − d0 

− 1⎟ 

+ 

2 

4 ⎣ 1− cos α ⎜ 0 1− cos α ⎟ 

⎝ 0 ⎠ 

2 2 2 2 

2d0 1 − (1 − ε ) cos α ⎛ 0 1 1 − (1 − ε ) cos α ⎞⎤ 

0 

+ − 1 

2 ⎜ + 2 ⎟⎥ 

3 1− cos α0 ⎝ 2 1− cos α0 

⎠⎦⎥ 

(16) 

Fig. 3 presents the graphic variation of the contracted muscle volume and 

its relative contraction. 

Fig. 3 – Vm = f(ε). 

The volume of the muscle in relaxed state can be calculated using the 

length of each fiber of the membrane and the number of windings which make 

the fiber around the tube. 

The relation for calculating a cylinder volume is 

Vm = πr h . (17) 

With the dependencies shown in Fig. 2, the muscle radius depends on its 

length 

2 2 

2 s − l 

r = . 

2 2 

4π 

n 

(18) 

From Eq. (18) results 

2 

2 2 2 2 

s = 4π 

r n + l . (19) 

Using Eq. (8) and Eq. (9), Eq. (19) becomes


2 2 

2 2 2 1 − (1 − ε) cos α0 

2 2 

0 2 0 

1− cos α0 

s = 4 π n r + l (1 − ε) 

. (20) 

Let g = thickness of fiber winding, and n = lo/g = number of windings 

made by the fiber around the tube. Eq. (20) becomes 

2 

2 2 

4π 

2 1 − (1 − ε) cos α0 

2 

0 2 0 

(1 ) 

2 

g 1− cos α0 

s = l r + − ε 

. (21) 

For a thickness g ranging from 0.5 mm and 1 mm can be plotted 

graphically the value of s (Fig. 4). For values of s ranging between 1.18·10 5 and 

2.25·10 5 mm, total volume of the muscle varies according to graph from Fig.5. 

Fig. 4 – s = f(g) Fig. 5 – Vm = f(s) 

Considering that the length of each fiber is constant (s = const) and n = 

number of windings made by the fiber around the tube, Eq. (17) becomes: 

2 2 

π ⎛ s − l ⎞ 

Vm = ⎜ l 

4 2 2 ⎟ . 

⎜ π n ⎟ 

⎝ ⎠ 


(22) 

1. The paper presented different calculus methods for pneumatic muscle 

volume. It has been considered two states of the muscle, according to the 

compressed air alimented – the relaxed state and the contracted state of the 

muscle. 

2 It has been plotted the main results - the variation of the contracted 

muscle volume and its relative contraction, the value of length fiber for given 

ranges of thickness and the value of the total volume depending on fibers 

length. 




POSDRU/88/1.5/S/59321.


REFERENCES 

Albienz J., Dillmann R., Kerscher T., Zollner J. M., Dynamic Modelling of Fluidic 

Muscles using Quick-Release. Forschungszentrum Informatik, Germany, 2005. 

Dragan L., Theoretical and Experimental Aspects Regarding the Geometrical 

Parameters of the Pneumatic Muscles. Proc. of the International Conference of 

the Carpathian Euro-region Specialists in Industrial Systems, 8th Ed., 12-14 

May, North University of Baia Mare, 2010. 

Deaconescu T., Deaconescu A., Pneumatic Muscle Actuated Isokinetic Equipment for 

the Rehabilitation of Patients with Disabilities of the Bearing Joints. Proc. of the 

International MultiConference of Engineers and Computer Scientists, Hong 

Kong, IAENG Hong Kong, 2009. 

Hildebrandt A., Sawodny O., Neumann R., Hartmann A., A Flatness Based Design for 

Tracking Control of Pneumatic Muscle Actuators. Seventh International 

Conference on Control, Automation, Robotics And Vision (ICARCV’O2), 

Singapore, 2002. 

CONSIDERAłII PRIVIND CALCULUL VOLUMULUI 

MUŞCHIULUI PNEUMATIC 

(Rezumat) 

Un muşchi artificial pneumatic este alcătuit dintr-un tub interior de lungimi 

variabile, realizat dintr-un material elastic, de obicei neopren. Tubul este învelit cu o 

Ńesătură formată din mai multe straturi, realizată din nylon, pentru a-i da rezistenŃă şi 

pentru a-l proteja de influenŃele din mediul de lucru. Sub acŃiunea aerului comprimat, 

muşchiul pneumatic îşi măreşte diametrul şi îşi micşorează lungimea. 

Un aspect de interes în studiul comportamentului în funcŃionare a unui muşchi 

pneumatic este acela al modului în care evoluează volumul acestuia sub acŃiunea aerului 

comprimat. În lucrarea de faŃă se prezintă o modalitate de calcul a volumului muşchiului 

pneumatic. Rezultatele obŃinute se referă la variaŃia volumului muşchiului în stare 

contractată în raport cu contracŃia relativă a acestuia, variaŃia lungimii unei fibre în 

raport cu grosimea acesteia şi variaŃia volumului total al muşchiului pentru diferite lungimi 

ale fibrelor.


Publicat de 



SecŃia 


MODERN TECHNIQUES FOR EXPERIMENTATION OF 

ADJUSTABLE HYDROSTATIC PUMPS 

BY 

TEODOR COSTINEL POPESCU ∗ and IOAN LEPĂDATU 



Hydraulics and Pneumatics Research Institute, 

INOE 2000 – IHP, Bucureşti 

Abstract. This paper aims at dissemination of experimental research work 

conducted at IHP Bucharest, in order to determine the actual performance of 

hydrostatic adjustable pumps. To this end laboratory tests were conducted on a 

radial piston pump MOOG type, which can automatically vary the flow rate by 

adjusting the eccentricity of the stator ring. The paper presents the experimental 

modern means used for this purpose, the type of experimental tests, results and 

their interpretation. 

Key words: hydrostatic adjustable pump, experimental tests. 


On the adjustable hydrostatic pump tests were conducted under static and 

dynamic mode (Popescu et al, 2010), (Lepădatu, 2010). Technique used for the 

dynamic tests was based on flow measurement by the indirect method, which 

consists of measuring the differential pressure across a diaphragm mounted on 

the pump discharge pipe. In this way, the two pressure transducers mounted in 

the downstream and upstream of the diaphragm measure almost instantaneously 

pressures and thus, very fast, is known the flow rate value. 

2. Presentation of the Test Stand 

Functional schematic diagram of the stand, Fig. 1, contains: radial piston 

pump tested PPR, driven by the AC electric motor ME, with power of 37 kW 

∗ Corresponding author: e-mail: popescu.ihp@fluidas.ro

150 Teodor Costinel Popescu and Ioan Lepădatu 

and 1465 rev/min; electropump EP, providing overcharging of aspiration of the 

pump tested; the filter F, for protection to contamination of the aspirated oil; 

diaphragm DM and pressure transducers TP1, TP2, which measure the flow 

discharged by the pump; transducer TPr for measuring the discharge pressure of 

the pump tested, proportional pressure valve SPP with a role in protection and 

simulation of the load; throttle DR, for load adjustment at the pump tested; the 

electric control hydraulic distributor D which in position "0" allows adjustment 

of valve SPP, in position "b" allows the pump discharge, and in position "a" 

allows load adjustment by means of throttle DR and calibration of the flow 

measuring diaphragm by means of the turbine flowmeter TD. 

Fig. 1 – Functional diagram of the test stand for the pump PPR. 

The stand also contains a data acquisition system SAD, type NATIONAL 

INSTRUMENTS, Bus-Powered M Series Multifunction DAQ for USB - 16-Bit, 

up to 400 kS/s, up to 32 Analog Inputs, Isolation, interface between pressure 

transducers, type DRUK SYSCOM 18, cod PTX 1400-400, Pn=400 bar, G1/4” 

(TP1, TP2, TPr), flow transducer type HYDAC, cod EVS 3100-1PTX 1400- 

400, 60l/min (TD), and temperature transducer type DRUK SYSCOM 18, cod 

PT 100-50...+400 O C, G1/2” (TT), on the one hand, command for adjustment of 

capacity of the pump tested Vc and acquisition of information about capacity 

achieved Vr, on the other hand, the command for pressure control of the tested 

pump discharge Pc and acquisition of information about the acieved discharge 

pressure Pr; a computer system PC, type NATIONAL INSTRUMENTS, cod


NI PXI-1031 with an installed data acquisition and processing software type NI 

LabVIEW; a stabilized voltage source STS; a random function generator GFA. 

For accuracy of measurements, noise of the acquired signals was filtered by 

means of a “low-passing” filter, with a cutting frequency of 100 Hz 

3. Calibration of Flow Measurement Diaphragm 

Under dynamic mode flow was measured with the measuring diaphragm 

located right on the discharge of the pump DM (acc. Fig. 1). For calibration of 

the diaphragm, the flow was measured accurately with a turbine flow meter 

"Turboquant". The calibration feature of the diaphragm was raised, namely the 

functional dependence of the flow, as measured by the turbine transducer, on 

the pressure drop at the diaphragm, measured with the pressure transducers TP1 

and TP2. Coordinates of the points on the interpolated calibration feature of the 

diaphragm are shown in Tables 1…3. 

Table 1 

Diaphragm calibration results: 0.000...1.723 bar; 0.000...19.393 l/min 

∆P, 

0.000 0.260 0.380 0.520 0.740 1.090 1.330 1.723 

BAR 

Q 

l/min 

0.000 6.856 8.594 10.633 12.757 15.547 17.228 19.393 

Table 2 


∆P, 

2.116 2.592 3.069 3.597 4.125 4.774 5.423 6.082 

BAR 

Q 

l/min 

21.554 23.860 25.936 28.169 30.262 32.252 34.624 36.783 

Table 3 


∆P, 

6.740 7.558 8.435 8.510 9.012 

BAR 

Q 

l/min 

38.916 41.495 43.955 44.113 49.987 

4. Experimental Measurements under Static Mode 

To determine the behavior of the pump under static mode there were 

applied control signals Uc within the range 0…10 V, ascending and descending, 

sinusoidal and ramp-shaped, with frequencies of 1 Hz, 0.7 Hz, 0.35 Hz and 0.1 

Hz. All measurements were made at pressure p = 20 bar and oil temperature 

Toil= 40 ± 5 0 C.


4.1. Response to Ascending / Descending Ramp Signal 

Measurements were made for control signals with maximum amplitude 

(10Vd.c.) and frequencies of 1 Hz, 0.7 Hz, 0.35 Hz and 0.1 Hz. For frequencies 

of 1 Hz results are shown in the diagrams of Fig. 2. 

Fig. 2 – Variation over time of flow Q, l/min and eccentricity e, mm, 

to ramp signal with frequency of 1 Hz. 

4.2. Response to Low Frequency Sinusoidal Signal 

Measurements were made for control signals with maximum amplitude 

(10Vd.c.) and frequencies of 1 Hz, 0.7 Hz, 0.35 Hz and 0.1 Hz. For frequencies 

of 0.1 Hz results are shown in the diagrams of Fig. 3. 

Fig. 3 –Variation over time of flow Q, l/min and eccentricity e, mm, 

to sinusoidal signal with frequency of 0.1 Hz. 

4.3. Pump Static Characteristic 

Dependency of the prescribed command on the flow achieved by the 

pump was determined for ascending/descending ramp-shaped control signal and 

frequencies between 0.1 Hz and 0.005 Hz. It appears that the hysteresis 

decreases with decreasing frequency, a phenomenon explained by the fact that 

the quantities progress being slower, on their decreasing variance there is more


time for the transitional regimes, relaxation of materials (especially in springs), 

etc. to diminish their effects. For the frequency of 0.005 Hz the static control 

characteristic of the pump is shown in Fig. 4. 

Fig. 4 – Pump static characteristic, flow Q, l/min depending on the prescribed 

command, voltage Uc, V, to ramp signal with frequency of 0.005 Hz. 

5. Experimental Measurements under Dynamic Mode 

5.1. Response to Step Signal of the Flow Discharged by Pump 

Determination of the response to step signal was performed to multiple 

amplitudes, respectively 100% (10 V), 75% (7.5 V) and 50% (5V) to see how 

response times depend on this parameter. At each of amplitudes the system was 

excited also with a train of step signals of different frequency, respectively 0.1 

Hz, 0.5 Hz and 1 Hz, to see if the response is maintained to frequency change. 

Response of the positioning system of the stator ring of the radial piston 

adjustable pump to step signals that have values less than 25% and 50% of the 

maximum value does not differ essentially, in aspect and performance (over 

adjustment, delay time, stabilization time) from the response to the maximum 

value signals, of 10Vd.c. 

Fig. 5 presents the variation over time of the pump flow to step control 

signal, with amplitude of 100% (10V) and frequency of 0.1 Hz. From this test 

results: delay time = 116 ms, stabilization time = 193 ms, over adjustment = 

zero. 

5.2. Response to Sinusoidal Signal of the Flow Discharged by Pump 

To determine the response of the pump adjusted flow there were applied 

control signals with different values of frequency (0.5 Hz, 6 Hz and 10 Hz) and 

amplitude (100%, 75%). Fig.6 presents the response to sinusoidal signal with 

amplitude of 100% (10Vd.c.) and frequency of 6 Hz.


Fig. 5 – Response over time of pump flow Q, l/min to step control signal, 

with amplitude of 100% (10V) and frequency of 0.1Hz. 

Fig. 6 – Response over time of pump flow Q, l/min to sinusoidal control signal, 

with amplitude of 100% (10V) and frequency of 6 Hz. 

Response of the positioning system to sinusoidal signals of various 

frequencies shows that amplitude of the response decreases, and the phase 

difference between "control" and "response" increases when the frequency of 

control signal decreases. Based on these responses resulted Bode diagram. 

5.3. Response in Frequency 

To determine the limit frequency at which the pump performs flow 

adjustability there was applied a sinusoidal control signal with maximum 

amplitude (10Vd.c.) and with ascending frequency, starting from 0.1 Hz to 20 

Hz for 10 sec. The evolution over time of the flow adjusted by the pump 

represents the response in frequency of the positioning system developed, 

shown in Fig. 7. 

It was noted that when increasing the frequency first minimum flow rates 

decrease and then the maximum ones; the phenomenon is due to the two springs 

of the small pistons (large and small spring), components of the positioning 

system of the pump stator ring, which adjusts the amount of eccentricity e, mm.


Fig. 7 – Response in frequency of pump flow Q, l/min to sinusoidal control signal, 

with amplitude of 100% (10V) and frequency of 0.1...20 Hz. 

5.4. Bode Diagram 

Dynamic performance of the radial piston pump, with mechatronic 

servomechanism for positioning of the stator ring, is highlighted by: attenuation 

of the response amplitude characteristic, depending on frequency of the control 

signal; characteristic of phase shift (delay) of response from command, 

depending on frequency of the control signal. This two characteristics form the 

Bode diagram, shown in Fig. 8. Amplitude and phase shift measurements were 

made for eight points of frequency: 0.1 Hz, 0.2 Hz, 0.5 Hz, 1 Hz, 2 Hz, 5 Hz, 10 

Hz, 20 Hz. From the Bode diagram it results that the amplitude attenuation of 3 

dB, i.e. a 30% reduction of the response amplitude, occurs at a frequency of 2.4 

Hz of the control signal. At this frequency the phase displacement is 35 0 . 

a 

b 

Fig. 8 – Bode diagram: 

attenuation, dB-Frequency, Hz − a; phase displacement, deg- Frequency, Hz − b. 


1. The article presents a minimal structure of a modern stand, on which 

can be carried out tests under static and dynamic mode for volumetric rotary 

machines, namely hydrostatic pumps and motors, fixed or adjustable. 

2. Performance of this stand has been highlighted by testing a radial 

piston pump, MOOG production, with capacity adjustable by means of a 

mechatronic system for positioning the stator ring between 0…32 cm 3 / rev.


3. The following performance of the pump resulted: a good 

proportionality between the control signal Uc, V and the flow displaced at 

constant rotational speed, l/min (from static characteristics); delay time = 116 

ms, stabilization time = 193 ms, over adjustment = zero (from response to step 

signal); attenuation = 30%, phase shift = 35 0 (from Bode diagram). 

REFERENCES 

Lepădatu I., Theoretical and Applied Research on Mechatronic Systems Adjusting the 

Flow of Hydraulic Rotary Generators by Eccentricity. Doctoral Thesis, 

“Politehnica” University of Bucureşti, 2010. 

Popescu T. C., Guşă Ion D. D., Călinoiu C., , Modern Instruments for Analysis of 

Hydrostatic Transmissions. P.U.B. Scientific Bulletin, Series D, 72, 4 (2010). 

TEHNICI MODERNE DE EXPERIMENTARE A 

POMPELOR HIDROSTATICE REGLABILE 

(Rezumat) 

Articolul îşi propune diseminarea unor lucrări de cercetare experimentală, 

realizate la IHP Bucureşti, pentru stabilirea performanŃelor reale ale pompelor hidrostatice 

reglabile. În acest scop s-au efectuat teste de laborator asupra unei pompe cu 

pistoane radiale de tip MOOG, care îşi poate varia automat debitul prin reglarea excentricitaŃii 

inelului stator. Lucrarea prezintă mijloacele moderne de experimentare utilizate 

în acest scop, tipul încercărilor experimentale, rezultatele obŃinute şi interpretarea lor. 

Pe un stand modern, s-a testat o pompă cu pistoane radiale, de tip MOOG, cu 

capacitate reglabilă între 0...32 cm 3 /rot. S-a etalonat o diafragmă pentru măsurarea 

debitelor, au fost ridicate caracteristicile în regim static şi dinamic ale pompei. Pentru 

acurateŃea măsurătorilor zgomotele au fost filtrate cu un filtru “trece-jos” cu frecvenŃa 

de tăiere de 100 Hz. Au rezultat: o bună proporŃionalitate între semnalul de comandă 

Uc, V şi debitul refulat la turaŃie constantă, l/min (din caracteristicile statice); timpul de 

întârziere = 116 ms, timpul de stabilizare = 193 ms, suprareglarea = zero (din 

răspunsul la semnal treaptă); atenuare = 30% , defazaj = 35 0 (din diagrama Bode).


Publicat de 



SecŃia 


SOME CONSIDERATIONS ABOUT COLD PLASTIC 

DEFORMATION OF BEARING RINGS 

BY 

OCTAVIAN PRUTEANU * , CONSTANTIN CĂRĂUŞU 

and LUCIAN TĂBĂCARU 

Received: June 12, 2011 



Department of Machine Manufacturing Technology 

Abstract. The equipment used to conduct experimental research enables the 

user to change the deformation force, the deformation feed, as well as the speed of 

the half-finish feeding roller. Paper describes the results obtained on the 

roughness, out-of-roundness and microhardness of the exterior and interior ring 

races of bearing 6210, further to changes in the three work parameters. The values 

detected were recorded in tables and presented by graphical representations, and 

they substantiate our conclusions related to the best work parameter ranges, where 

the quality parameter deviations are minimal. The paper also shows the 

mathematical processing of the microhardness results recorded for the ring race 

surfaces. 

Key words: plastic deformation, force, feed, speed, roughness, out-ofroundness, 

circularity, microhardness 


Plastic deformation of steels has enjoyed an increasingly widespread 

use in most of the industrial facilities around the world. As concerns the bearing 

manufacture industry, the method was first used by the 1 GPZ factory in 

Moscow and it was applied to cold plastic deformation of ring races of radial 

ball bearings. As for the Romanian bearing manufacture industry, this process 

was first employed by the company SC RulmenŃi SA of Bârlad, in 2005, further 

* Corresponding author: e-mail: pluteanu@yahoo.com

158 Octavian Pruteanu et al. 

to a research contract initiated by the Machine Manufacturing Technology 

Department of the Faculty of Machine Manufacturing and Industrial 

Management of Iaşi, due to the advantages it enjoys, namely: 

i) the existence of high precision and high productivity Japanese 

manufacturing equipment; 

ii) the advantages brought about by the improvement of specific part 

operation properties: resistance to wear and tear, improved endurance and 

increased profitability. 

The research conducted by SC RulmenŃi SA Bârlad consisted of the 

setting of the best values of the work parameters (deformation force, 

deformation feed, and speed of the half-finish feeding roller), required to 

achieve acceptable values for the roughness, circular shape and microhardness 

of the bearing ring races manufacturing to cold plastic deformation. 

2. Experimental Conditions 

The following conditions were created before conducting the 

experimental research. 

a) material: 100Cr6 – used especially in the manufacture of bearings, 

whose mechanical properties are shown in table 1, 

b) half-finish: hot rolled rings, whose sizes meet the requirements of the 

type 6210 bearing, 

c) equipment: special CRF 120 OR equipment for the exterior rings and 

CRF 70 IR equipment for the interior rings, 

d) tools: special mandrels and rollers for the type of bearing that is 

manufactured, 

e) measuring and control devices: for roughness: Taylor Hobson 

Formtalysurf, series 2, for out-of-roundness: Perthometyer Marsurf CD 120, for 

microhardness: Akashi MVK-D 

100Cr6 Force 0.2 

daN 

Table 1 

Mechanical properties of 100Cr6 steel 

Ultimate Yield Stretch Break 

strength 

daN 

point Rp 

0.2 

resistan. elonga. 

A% 

daN/mm 2 

Rm 

daN/mm 2 

Hardness 

HB 

Rolled 

7510… 

7550 

8900… 

8950 

82.45… 

95.60 

97.75… 

113.30 

3…5 329…345 

Annealed 5120 5697 65.20 72.50 7 219 

3. Results of the Experimental Research 

Table 2 shows the results obtained by the processing of the bearing 

6210 rings that is the roughness, out-of-roundness and microhardness of the 

races manufacturing to cold plastic deformation. The values of the work 

parameters shown in the table were used.

Work parameters Roughness 

µm 


Table 2 

Experimental results 

6210-10 – outer ring 

Quality parameters 

Circularity 

Microhardness, HV 

mm h1 h2 h3 h4 h5 

P = 17600 daN 0.30 0.16 302 309 297 283 281 

A = 34 mm/min 0.10 0.20 302 301 302 301 286 

n = 68 rot/min 0.11 0.15 311 303 308 289 265 

6210-20 – inner ring 

P = 7530 daN 0.34 0.05 283 295 301 303 294 

A = 30 mm/min 0.27 0.20 265 270 289 287 281 

n = 118 rot/min 0.21 0.07 281 283 276 275 282 

Remarks: a) P – deformation force; A – deformation feed; n – rpm of the half-finish 

feeding roller; b) The values of each variable parameter were recorded when the values 

of the other two parameters were constant. 

4. Graphical Representation of Experimental Results 

4.1. Influence of the Deformation Force 

The maximum forces allowed by the equipment CRF-120 OR – and 

CRF-70 IR – were used on the exterior and interior rings of bearing 6210. The 

other work parameters were kept constant and their values are shown in Table 2. 

Figs. 1 and 2 show the roughness of the ring races, whereas Figs. 3 and 

4 show the out-of-roundness of the same rings, and Figs. 5 and 6 are the 

representation of the race microhardness variation. 

Fig. 1 – Roughness of exterior ring race when the deformation force P=17600 daN.


4.2. Influence of the Deformation Feed 

The experimental results were recorded when the deformation feeds of 

the two pieces of equipment used were minimal, and they are shown in Figs. 7 

and 8 for race roughness, Figs. 9 and 10 for out-of-roundness and Figs. 11 and 

12 for race microhardness. 

Fig. 2 – Roughness of interior ring race when the deformation force P=7530 daN. 

Fig. 3 – Out-of roundness of exterior ring race when P=17600 daN.


Fig. 4 – Out-of-roundness of interior ring race when P=7530 daN. 

Fig. 5 – Microhardness of exterior ring race when the deformation force P=17600 daN.


Fig. 6 – Microhardness of interior ring race when the deformation force P=7530 daN. 

Fig. 7 – Roughness of exterior ring race when the deformation feed A=30 mm/min.


Fig. 8 – Roughness of interior ring race when the deformation feed A=30 mm/min. 

Fig. 9 – Out-of roundness of exterior ring race when the 

deformation feed A=30 mm/min.


Fig. 10 – Out-of-roundness of interior ring race when 

the deformation feed A=30 mm/min. 

Fig. 11 – Microhardness of exterior ring race when the deformation feed A=30 mm/min.


Fig. 12 – Microhardness of interior ring race when the deformation feed A=30 mm/min. 

4.3. Influence of the Rotation per Minute of the Part Feeding Roller 

The experimental results were recorded for the default rpm of the 

feeding rollers of the two pieces of equipment used, and they are shown in Figs. 

13 and 14 for surface roughness, Figs. 15 and 16 for out-of-roundness and Fig. 

17 and 18 for ring race microhardness. 

Fig. 13 – Roughness of outer ring race when the roller rpm n=68 rot/min.


Fig. 14 – Roughness of inner ring race when the roller rpm n=118 rot/min. 

Fig. 15 – Out-of roundness of outer ring race when the roller rpm n=68 rot/min. 

5. Conclusions on Quality Parameter Variation 

The influence of the technological parameters on the quality parameters 

consists of the following processes: 

i) the deformation force has a positive influence on the quality 

parameters, in the sense that the values of the Ra race roughness meet the 

requirements of the plastic deformation stage, i.e. 0.30 µm and 0.31 µm for the 

interior ring, respectively, the out-of-roundness has normal values, i.e. within 

the 0.16 mm and 0.053 mm range, whereas the microhardness of the upper coldworked 

layer of the races is 302 and 283 units, respectively. May therefore 

conclude that the values of the deformation force used in this experiment are 

adequate. 

ii) the values of the deformation feed are thought to meet all the quality 

parameter requirements, in the sense that the values of the Ra race roughness is 

0.10 µm and 0.27 µm, respectively, the out-of-roundness of both rings is 0.20 

mm, whereas the microhardness is 302 and 265 units, respectively.


Fig. 16 – Out-of-roundness of inner ring race when the roller rpm n=118 rot/min. 

Fig. 17 – Microhardness of outer ring race when the roller rpm n=68 rot/min. 

Fig. 18 – Microhardness of inner ring race when the roller rpm n=118 rot/min.


iii) the quality parameter values recorded are within the range of the 

current technical requirements for the rpm of the feeding roller, i.e. the values of 

the Ra roughness is 0.11 µm and 0.21 µm, respectively, the out-of-roundness is 

0.15 mm and 0.07 mm, respectively, and the race surface microhardness is 311 

and 281 units, respectively. 

REFERENCES 

Leonte P., Pruteanu O., Cărăuşu C., Şerban C., The Effect of Deformation Degree upon 

the Form, the Roughness and the Microhardness of Surface Processing by Cold 

Plastic Deformation. Proceedings of the 13th International Conference, Modern 

Technologies, Quality and Innovation, 21-23 May 2009, p. 371. 

Lupescu O., Netezirea suprafeŃelor prin deformare plastică (Plastic Deformation 

Surfacing). Technical Info Publishing House, Chişinău, Republic of Moldova, 

1999. 

Pruteanu O., Cărăuşu C., Tăbăcaru L., Grămescu T., Influence of Deformation Feed on 

the Roughness and Shape Precision of Cold Worked Surfaces. International 

Journal of Modern Manufacturing Technologies, I, 1, 61, Politehnium Publishing 

House (2009). 

CONSIDERAłII ASUPRA DEFORMĂRII PLASTICE 

LA RECE A INELELOR DE RULMENłI 

(Rezumat) 

Echipamentul utilizat pentru cercetările experimentale permite modificarea 

valorii forŃei de deformare, a avansului şi a vitezei rolei. Lucrarea descrie rezultatele 

obŃinute asupra valorilor rugozităŃii, abaterii de la circularitate şi microdurităii la 

exteriorul şi interiorul inelelor de rulment 6210, în funcŃie de variaŃia celor trei 

parametri de lucru. Valorile obŃinute au fost prezentate sub formă de tabele şi 

reprezentate grafic şi susŃin concluziile noastre privind valorile optime ale domeniilor 

de variaŃie ale parametrilor de lucru, în condiŃiile în care abaterile sunt minime. De 

asemenea, este prezentat modul de prelucrare matematică a valorilor microdurităŃii 

înregistrate pentru suprafeŃele inelelor de rulment.


Publicat de 



SecŃia 


THE COSTS ESTIMATION 

FOR THE MECHANICAL PRODUCTION 

BY 

BRUNO RĂDULESCU ∗ and MARA-CRISTINA RĂDULESCU 

Received: September 15, 2011 


” Gheorghe Asachi” Technical University of Iaşi, 


Abstract. This paper presents the decomposition and the definition of a 

technical cost that include supplier cost, subcontracting cost and production cost. 

Key words: production programme, cost, project budget. 


The companies live in a permanent competition environment and they 

are continuously searching the maximum profit from the markets. Because of 

that technological and economical concurrence, managing the project budgets or 

a product budget, is not only an advantage, but is an obligation if that company 

want to survive. 

2. The Costs Estimation 

Today, we consider that there are three different views of the estimated 

cost by type of decision. Indeed, the concern of management who is responsible 

for defining the company strategy is mostly the "boundary" or profit realizable 

on a given project. It is indeed vital for all businesses to generate profit to be 

able to continue to exist. As for the commercial department, which aims to 

position the best product over its competition, it seeks to know the best selling 

price for its product on that market; this means the product market price. 

∗ Corresponding author: e-mail: bruno_radulescu@yahoo.com

170 Bruno Rădulescu and Mara Rădulescu 

The selling price and the profit area, then form a set defining the 

economic constraints of the project requiring a maximum cost of production. 

The project cost has become a design parameter as well as technical 

specifications. 

We found that the designing phase of a product is 70% from the period 

spent from idea to emerge that product, but the influence in the final cost is less 

than 10% from the entire cost of the product Fig. 1 (Miller, 1988). 

Fig. 1 – Influence of design on manufacturing cost for the Ford vehicles (Miller, 1988). 

We notice also, that any advances in the product development means that 

the expenses to modify that product from that point are higher Fig. 2. It is 

important to manage the product cost earlier as possible in it’s the life cycle. 

Fig. 2 – Cost variation progress and their influence on the final cost. 

The price offer is more accurate if that price offer is based on the 

production criteria or on the complete production programme, allowing to know 

with precision all the involved costs and also the tool productivity. The price 

offer is becoming a criterion permitting to choose between different production 

solutions.


For some companies, such as subcontractors in mechanical engineering, 

the development of an quotation is very important. We consider that in this area, 

only 10% of quotes are converted into firm orders. The quality of the provided 

documents, the quotation accuracy and response time are all criteria that 

determine the quality of a quotation. 

3. Decomposition and Definition of Technical Cost 

We can consider the costs based on the ability to affect directly or not a 

product or an activity. They include: 

a) direct costs: it is, in most cases, disappearing with the production 

costs (labor, material, tools, use of tools machines ...); 

b) indirect costs: they represent the required expenses to produce, but 

cannot be assigned to a specific product (insurance, management ...). 

The distribution of indirect costs is a delicate operation. It depends 

mainly on the strategic choices made by company management. Therefore, we 

will not talk about those costs acquired called structure cost. 

In this paper, we take into consideration only the technical costs: supply 

cost, subcontracting cost and production cost. 

This cost includes expenses dependent only on the chosen solution and 

the means of production involved. 

3.1. Supply Cost 

The supply cost essentially represents the cost of raw materials. We 

consider two types of raw materials: the standard material: laminated, …; the 

special material: casting, forging, … 

Many studies show that the supply cost is a very important part of 

technical cost (average 50%) (Boothroyd, 1988). It is therefore important to 

determine its cost of obtaining in a realistic and fast manner (Ou-Yang & Lin, 

2011). 

In the case of a standard material, usually is associated to each type of 

raw material a cost per unit of weight and length, allowing to know easily the 

direct cost of a standard material. 

In the case of a special material, it is difficult to accept to be limited in 

the case of a quotation, to estimate an average cost of production using the 

average cost per kilo. The company that would sell complex parts will be 

underestimated. 

3.2. Subcontracting Cost 

There are three types of subcontracting cost: 

a) Subcontracting the means: some tasks require very specific operation 

or equipment that the company does not own, in which case it is necessary to 

use the subcontracting.


b) Subcontracting the capacity: this is a call to a subcontractor because 

the company is overloaded, the company can estimate the outsourcing, or 

estimate the subcontracting cost of work as if the operation was conducted in 

the company, or may have the effect of reducing the company income. 

c) Subcontracting to lower cost per hour: it often happens that the 

hourly cost of the company, for certain activities, is higher than that proposed 

by some subcontractors, it is still necessary to make a more detailed analysis of 

changes in hourly costs in this case, if one considers that the hourly cost Ch 

takes into account fixed and variable costs 

CV + CF 

Ch = , (1) 

Np 

where: CV – variable charges, CF – fixe charges, Np – number of products. 

3.3. Production Cost 

It is still important to note that the cost raw materials depends on the 

dimensional accuracy and required specifications. Therefore, choosing a method 

of obtaining can be done by comparing the cost of processing each of the 

manufacturing possibilities. 

The manufacturing cost is broken down into: 

a) manufacturing preparing cost; 

b) machining cost; 

c) cutting tools cost. 

3.3.1. Manufacturing Preparing Cost. The manufacturing preparing 

cost for all operations is concerning the definition of manufacturing strategy, 

development and installation and / or implementation of the devices needed to 

machining. They are: 

a) production programme (Weill, 1993); 

b) CNC program (Prudhomme, 1996); 

c) the griping devices (Gladel et al., 1992); 

d) editing tools; 

e) mounting the piece. 

All of these costs are very difficult to estimate. Indeed, these operations 

represent a human work (physical or mental). Yet the determination of human 

time (the costs) is subject to the rigors of the work. Many parameters can 

disrupt a normal execution of the operation: 

a) staff qualifications; 

b) the exact knowledge and respect for the procedure; 

c) the influence of the environment; 

d) learning;


e) the monotony … 

(Anselmett, 1995) proposed a relation to define the programming time 

depending on the difficulty of machining the piece 

T = 10min+ 8min/phase + 1.5min/tool + 2min/operation (2) 

CN 

where: TCN – programming time for CNC. 

3.3.2. Machining Cost. The machining cost is the cost generated when 

the machine is operating. The total time in this case corresponds to the total 

time of movement of the approaching tool and speed work, time to release the 

tool, automatic tool change when possible. Determining the cost of machining 

requires knowledge of hourly Cm machine that can be evaluated as follows. The 

simplified model of the machine cost per hour includes: 

a) Amortization of the machine and financial expenses 

P Pi 

A = + , (3) 

HN 2H 

where: A – amortization, P – toll machine price, H – hours of use per year, N – 

duration in years of amortization, i – interest rate. 

b) Operating costs 

Fr SI 

R = + + 0.6eW + 1.6CS 

, (4) 

H H 

where: Fr – repairs and annual maintenance, S – area occupied by the machine, I 

– local costs per m 2 per year, W – machine power in kW (used only 60%), e – 

cost of kilowatts per hour, CS – salary costs (1.6 CS includes payroll taxes). 

We have the hourly cost of the machine 

C m 

= A + R . (5) 

3.3.3. Cutting Tools Cost. The cutting tools cost is the replacement cost 

of these tools. This replacement is directly related to wear. We show that the 

cutting tools costs and machine use is dependent. Indeed, when the cutting 

speed Vc increases, the machining time decreases in the same proportions. 

However, the tool wear also increases and the cutting tools cost per unit 

increases. To define the optimum, it is necessary to model the law of tool wear. 

There are different models in the literature of the law of wear (Pallot, 1988). 

The model wearing the simplest and most commonly used is the simplified 

Taylor model that can be written as follows 

c 

n 

V T = Cte = α , (6)


where: Vc – cutting speed, T – life of the tool, n – Taylor coefficient. The 

economic tool life follows the relation 

⎛ 1 ⎞ Co 

T0 

= ⎜ −1 

n 

⎟ , (7) 

⎝ ⎠ C 

n 

C0 

0 

m 

V αT − 

= , (8) 

where: T0 – economic tool life, VC0 – economic cutting speed, Co – tools cost, 

Cm – hourly cost of the machine. 

From the economic tool life is estimated easily the number of 

replacements. We determine the replacement cost of a tool taking into 

consideration the type of tool (re-sharpened or removable inserts) as follows. 

Replacement cost of a re-sharpened tool 

C 0 = C + C + C , (9) 

OU 

where: COU – cost of the tool for lifetime T (between two sharpening), 

P 

af 

c 

O C OU = , (10) 

na 

where: PO – purchase price of the tool, na – number of sharpening means, Caf – 

sharpening cost, Cm – hourly cost of the machine. 

Caf = taCa , (11) 

where: ta – sharpening time, Ca – hourly cost of sharpening. 

where: tc – tool change time, Cc – cost of tool change. 

Cost of replacing a tool with removable inserts: 

CC = tcCm , (12) 

C = C + C + C 

0 OU f c , (13) 

where: COU – cost of the removable inserts, Cf – amortization of fastening 

elements , Cc – cost of removable inserts. 

C 

OU 

ZP 

= , (14) 

a 

where: Z – number of teeth, P– purchase price of a insert, a – number of edges 

to use.


C f = ZfTf 

, (15) 

where: f – cost of fasteners for a insert, Tf– life of fasteners (change in number 

of edges), a – number of edges to use. 

Ccc = Zt pCm , (16) 

where: tp – time for a inset changing. 

It is important to note that the estimated cost of cutting tools is very 

well managed with respect to annual purchases and the cost of sharpening. 

However, determining the costs of storage, recycling, regulation and 

management of cutting tools is generally unknown. It is in fact indirect costs 

such as: 

a) the purchase; 

b) receipt and control of supplies; 

c) the physical storage in the store; 

d) preparation of research tools and components; 

e) updating the file management of cutting tools. 

The estimate of these costs depends heavily on the organization of the 

company. To define, it is necessary to survey the various stakeholders: 

warehousemen, buyers, ... 


1. In this paper, we have demonstrated the importance and complexity of 

the estimated cost within the company. Then we identified the different 

components of cost. 

2. It was noted also that for all quotations, it is necessary to establish a 

database or knowledge of technical and economic reliability. Indeed, it is very 

important to note that most quotations are made by extrapolation of the past. 

REFERENCES 

*** Manuel de données technologiques en fraisage. CETIM, 1984. 

Anselmetti B., Génération automatique de devis pour l’usinage sur MOCN. Revue 

d’Automatique et de Productique Appliquée, 8, 1, 81-100 (1995). 

Boothroyd G., Estimate Cost at an Early Stage. Annals of CIRP (1988) . 

Gladel G., Gourdet D., Tous J.-L., Matériaux pour outils de coupe. Traité Génie 

Mécanique, 1992. 

Miller F.W., Design for Assembly: Ford’s Better Idea to Improve Products. 

Manufacturing System, 1988.


Ou-Yang C., Lin T.S., Developing an Integrated Framework for Feature-based Early 

Manufacturing Cost Estimation. The International Journal of Advanced 

Manufacturing Technology, 13, 9, 618-629 (2011). 

Pallot B. Prédétermination des temps d’usinage relatifs aux séries limitées – 

Application au tournage numérique. Thèse ENSAM Paris, 20 déc., 1988. 

Prudhomme G., Commande numérique des machines-outils. Traité Génie Mécanique, 

1996. 

Weill R.D., Conception des gammes d’usinage. Traité Génie Mécanique, 7-25, 1993. 

ESTIMAREA COSTULUI DE PRODUCłIE ÎN CAZUL PRODUSELOR 

(Rezumat) 

Această lucrare descrie descompunerea unui cost în componentele sale: costul cu 

aprovizionarea, costurile cu colaboratorii şi costul de producŃie.


Publicat de 



SecŃia 


THE ROLE OF PROFESSIONAL COMPETENCE 

IN BUSINESS ENTREPRENEURSHIP 

BY 

LUCIANA-CRISTIANA STAN *, and VLADIMIR MĂRĂSCU-KLEIN 



“Transilvania” University, Braşov, 

Department of Engineering and Management 

Abstract: In a market economy based on competition and risk, business 

sector development is a priority. For the Romanian economy, business sector 

development, especially skills are an important part of economic restructuring 

policy, which positively influence the pace of sustainable economic growth. By 

their characteristics: innovation, close ties with the community, high dynamic, 

optimal exploitation of local resources, creating jobs, developing small and 

medium impact, especially at lower levels, namely the local and regional levels. 

Entrepreneurship skills are reflected in fact the result of ideas and projects that 

illustrate the determination and attitude of the entrepreneur to identify and use 

opportunities that arise at a time. The motivation to create a new business is a 

prerequisite for the emergence of entrepreneurial initiative and receptiveness to a 

product or service is what brings success. 

Key words: responsible entrepreneurship, professional competence, 

questionnaires, sample. 


The research was NQFHE, that the National Framework of 

Qualifications in Higher Education skills fall into three categories: 

a) general professional skills − are covered by a range of skills which 

will enable further studies exercise professional roles in a wider field of activity, 

enabling use of the overall integrated, coherent, dynamic and open knowledge, 

* Corresponding author: e-mail: lucicri72@yahoo.com

178 Luciana-Cristiana Stan and Vladimir Mărăscu-Klein 

skills (eg cognitive, actionable, relational and ethical) and other acquisitions (eg 

values and attitudes) in a given field. 

b) specific skills − those skills are covered by running a specific program 

for graduate studies to cope with the demands of a specific profession, allow use 

of the overall integrated, coherent, dynamic and open knowledge, skills (eg 

cognitive, actionable , relational and ethical) and other acquisitions (eg values 

and attitudes) in the exercise of certain professions that 

2. Professional Competence in Entrepreneurial Business 

2.1. Implementation Methodology 

2.1.1. The Starting Point. Qualitative marketing research used a variety 

of investigative methods and techniques, mostly psychological. The most 

suitable types of qualitative research methods and research techniques are 

chosen nondirectiv depth interviews, plus paper and pencil interview focused on 

group discussions Focus group This type of survey is a technique based on 

interviewing relevant persons, highly qualified, available experience in areas 

related to business performance and maintain a high position on the economic 

plan. It is a survey based on direct personal interviews, informal, which 

involves both complex questions and a free discussion, which provides an 

opportunity for the interviewees to express their views. 

2.1.2. An Intermediate Result. The questionnaire was applied to a sample 

of 14 subjects randomly selected socio-demographic characteristics of which 

are outlined below: 

i) aged between 22 and 61 years with an average of 44 years; 

ii) 28.6% of surveyed have postgraduate (masters, doctorate), 64.3% are 

graduates and 7.1% have completed a technical school; 

iii) respondents are specializations in the fields of economics, 

engineering, socio-human, financial, management, technology, the Romanian 

language - respondents are the following professions: economist, engineer, 

administrator, school psychologist, librarian, manager, - 64.3% active and 

35.7% in the private sector in the state 

Presentation skills training in the sample In the sample, were presented 

skills, people in the sample will indicate which of these 9 would like to develop 

their skills. Thus: 

i) Communication skills in their mother tongue: 92.9% of subjects 

correctly asserts that speak the language of all respondents said that home and 

write correctly in their native language. 

ii) Communication skills in a foreign language: 85.7% of respondents 

stated that speak foreign languages, while only 14.3% say they do not speak a 

foreign language. 

iii) The level of competence in using information and communication 

technologies, digital technologies (computers, internet, mobile phone etc..): All


participants declare that they have 100% skills and abilities to use multi-media 

technology. 

iv) Math Skills : 78.6% of respondents say they are good at math, 14.3% 

declared that they have mathematical skills and 7.1% can not determine the 

degree of mastery of computer skills. 

v) Half of the respondents in the sample can easily solve mathematical 

problems or other sciences, 14.3% declared that they have these skills and 

35.7% can not appreciate the ease with which exact solves. 

vi) Ability to "learn to learn": 92.9% of respondents have the ability to 

learn new things and to integrate new knowledge into the old ones already 

existing, while 7.1% can not determine the degree of learning some new things. 

78.6% of respondents believe they have time and willingness to engage in a 

career development activity. 14.3% can not appreciate this and 7.1% did not 

have time /availability for this business. 

vii) Skills in interpersonal relationships: 92.9% of respondents declared 

that they have difficulty in communicating with colleagues working in while 

7.1% say they face obstacles in communicating effectively with their peers. 

viii) Also for a percentage of 92.9% of those polled say they have 

teamwork skills while 7.1% can not determine the degree of development of 

their teamwork skills. 

ix) Spirit of initiative and entrepreneurial skills: half of the respondents 

have the skills to start a business, 42.9% can not appreciate the skills of 

entrepreneurship, while 7.1% say they know how to start a business. 

x) In terms of "openness to innovation", 92.9% of respondents were happy 

to experience new things but 7.1% can not appreciate this. 64.3% of persons in 

the sample report that they feel the opposite fear of risk taking, 14.3% of 

subjects declared that they feel comfortable taking the risk, and finally, 21.4% 

of them do not and can appreciate the anxiety experienced in relation to risk 

taking. 92.9% of respondents are reluctant to again, while 7.1% say they are 

resistant to innovation. 

xi) Entrepreneurial skills main-respondents are in engineering and 

management fields, so that 60% of respondents engaged in economic 

engineering, 20% are employed in the economic, industrial 10% and 10% in 

other related fields (Baciu, 2011). 

Remark. Developing professional skills. Of the nine skills needed to 

develop personal and professional journey, say they would like to improve the 

following: the skills of "learning to learn, civic and social networking skills, 

communication skills in a foreign language; entrepreneurial skills, skills and use 

of digital multi-media technology. 

2.2. Necessary Skills and Personal Development Profesionale 

The questionnaire aims to identify the degree of development of the 

Route 9 skills necessary personal and professional development in key areas in


which respondents operate: language communication skills, communication 

skills in a foreign language, mathematical skills in science and technology, 

digital literacy, the use of multi-media technology, the skills of "learning to 

learn", civic and social networking skills, entrepreneurial skills, skills of artistic 

and cultural expression 

3. Application of Professional Competence 

This study is designed on the basis of questionnaires completed by 

project participants. Questionnaires were distributed to all participants in the 

marketing research aim is to identify the entrepreneurial skills acquired which 

brought the project participants. Based on the analysis of all questionnaires 

highlight the initial findings on participants' learning experiences, learned 

entrepreneurial skills, vision and business results immediately perceived in 

terms of entrepreneurship. form of tables and graphs: 

D e f i n i t i o n 1. The graphic representation in the level of studies 

(Table 1, Fig. 1) (Baciu, 2011). 

Table 1. 

Educational studies 

Studies Percent 

High School 7.1 

Education Technical 64.3 

Other Higher 28.6 

D e f i n i t i o n 2. Graphical representation depending on the 

specialization (Table 2, Fig. 2) (Baciu, 2011). 

Nr. 

crt. 

Table 2 

Specialization 

Specialization Percent 

1 Economics 28.6 

2 Financial 7.1 

3 Engineering 14.3 

4 Romanian language 7.1 

5 Management 7.1 

6 Sciences humanities 14.3 

7 Tehnology management 7.1 

8 Tehnology equipment 7.1

Studies 

28,6% 

Tehn. Equipm. 

7,1 

Tehn. 

7,1 

Scince humanit. 

14,3 

Man 

7,1 

Rom.langu 

7,1 


Fig. 1 – Representation based on studies. 

Specialization 

Fig. 2 – Representation on the specialization. 

7,1 

Econom 

28,6 

Financ 

7,1 

Engin. 

Remarks. 1. A percentage of 64.5% of respondents are university 

graduates. Only 7.1% are technical school graduates 

A percentage of 8.6% of respondents have completed other studies. 

T h e o r e m 1. Studies. 

P r o o f. A percentage of 28.6% of respondents are specialized in the 

economic profile which represents a majority 

T h e o r e m 2. Field activity. 

igh High 

school 

Education 

technical 

7,1 

% 

64,3 

% 

14,3


P r o o f. A percentage of 14.3% of respondents are specilize the profile 

of engineering to equal the number of respondents specilize the profile of social 

and human sciences. 

L e m m a 1. Areas that will initiate business. 

P r o o f. Specializations: finance, Romanian, management, technology 

and equipment technology have equal percentages of 7.1% 

Remark. The problems faced by potential women entrepreneurs- 

economic dependency and poverty are major problems especially in rural areas. 

C o r o l l a r y 1. The motivation to create a new business is a 

prerequisite for the emergence of entrepreneurial initiative for rewarding and 

successful 

4. Conclusions on the Leading Successful Business Entrepreneurs 

1. The results of this report indicates that women are strongly involved in 

entrepreneurial activities prior to the start of business (at a rate of 9.88%) while 

men have a somewhat higher percentage in developing business in recent years 

(16.44%). It was also found that the average age of those involved in 

entrepreneurial activities is between 33 and 35 years. 

2. However, it notes that people aged between 36 and 50 years are 

involved in a much greater extent in previous entrepreneurial start-up business 

activities (9.52%). In the case of the recently become entrepreneurs, the highest 

proportion is found in people aged between 26 and 40 (18.27%).The results 

indicate the importance of the entrepreneur's family presence in one or more 

persons to enhance entrepreneurial entrepreneurial activities. 

3. A share of 38.45% of the people involved in previous work reported 

the presence of an entrepreneurial start-up business in the family, the example 

most often given to the father (22.92%). For people newly entrepreneurs, 

20.63% of respondents have a family business, father is the family member 

most often given as an example (15.17%). 

4. More interestingly, the results indicate that respondents consider social 

aspects as major factors that motivated the decision to engage in entrepreneurial 

activities. 




POSDRU/88/1.5/S/59321. 

REFERENCES 

*** ANIMMC, A Youth Entrepreneurial Training Course. START Program, Bucureşti, 

2006. 

Baciu F., Role of Women Living in Entrepreneurship. available at: http://www. 

g2entrepreneurship.eu/Ro.../antreprenoriat accessed: 2011/03/10, 2011.


Bălăsescu M., Guidelines for Achieving the Business Plan, Marketing Plan, Marketing 

Program. Transilvania University, Braşov, 2004. 

Brătucu G., Public Service Marketing, Publishing Infomarket, Braşov, 2004. 

Drucker P F, System Innovation and Entrepreneurship - Practice and Principles. Ed. 

Enciclopedică, Bucureşti, 1993. 

Pascu I.,. Professional Skills and Entrepreneurship. available at: http://www.businessedu.ro; 

accessed: 2011/03/02, 2011. 

Popescu C.; Brătucu G., Marketing Research. Guideline – Practical. Transylvania 

University, Braşov, 1999. 

Rugman A., Collinson S., International Business. 4th Ed., Prentice Hall, 2006. 

ROLUL COMPETENłELOR PROFESIONALE 

ÎN ACTIVITATEA ANTREPRENORIALĂ 

(Rezumat) 

Pentru economia românească, dezvoltarea sectorului antreprenorial, dar, mai 

ales, competenŃele profesionale, constituie o componentă importantă a politicii de 

restructurare economică, ce influenŃează pozitiv ritmul creşterii economice durabile. 

Prin inovaŃie, legături strânse cu comunitatea, dinamică ridicată, valorificarea optimă a 

resurselor locale, crearea de locuri de muncă, intreprinderile mici si mijlocii 

influenŃează dezvoltarea, la nivele locale si regionale. IniŃiativele antreprenoriale 

reflectate în competenŃe profesionale reprezinta rezultatul unor idei si a unor proiecte ce 

ilustrează determinarea şi atitudinea întreprinzătorului faŃă de identificarea şi 

valorificarea unor oportunităŃi care apar la un moment dat.


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SOME ASPECTS REGARDING HYDRAULIC SYSTEMS 

WITH SECONDARY CONTROL 

BY 

IRINA TIłA * and IRINA MARDARE 





Abstract. The high costs for energy and the increased necessary of it lead to 

hydraulic systems working at a certain power with an optimum efficiency and 

consequently minimum energy consumption. The calculus of such systems is 

focused on the energy price and not on investment only. The hydraulic systems 

with secondary control are not only saving energy but also recovering it. In the 

paper two cases of such systems are presented. One can see the Simscape diagram 

for one of them and the simulation results for both. 

Key words: hydraulic system, secondary control, numeric simulation. 

1. General Considerations 

The high costs for energy and the increased necessary of energy lead to 

hydraulic systems working at a certain power, with an optimum efficiency and 

consequently minimum energy consumption. The calculus of such systems is 

focused on the energy price and not on investment only. 

The most common strategy for controlling hydraulic systems is still 

primary control. In addition to the primary controlled systems, constant pressure 

systems could be applied. This concept characterizes secondary controlled 

systems, where the hydraulic output units are connected to a constant pressure 

rail. Displacement control of the secondary units, support direct control of the 

output torque or force load (Rydberg 2005). 

* Corresponding author: e-mail: iddtita@yahoo.com

186 Irina TiŃa and Irina Mardare 

Using secondary control adaptive hydraulic systems in impressed 

pressure circuits one can obtain a substantial growth of efficiency. This control 

type uses hydrostatic displacement machine with variable displacement volume 

and implies that from a system with impressed pressure one can receive power 

using displacement control and not resistance control. The hydrostatic 

displacement machines may work both as pump and as motor. Hydraulic energy 

obtained when it works as a pump is transferred to other consumers or 

accumulated. 

Energy recovery is fulfilled when hydraulic machine works as a pump 

and the energy is stored in an accumulator (Backé, 2005), (Rydberg 2005). 

Using the concept of secondary control it is possible to constitute a feeding 

network like the one in electric networks. Using an accumulator as an energy 

storing device makes possible to design a system with improved endurance 

compared with a similar one without accumulator (Kirka, 2009). 

Tasks for the future can be seen in the recognition of possible new 

applications that could not be identified so far. (Murrenhoff, 2006). 

Simulation of the system is important in order to adjust parameters for 

dynamic characteristics improvement. The aim of this paper is to present a 

block diagram for Matlab simulation. 

2. Simscape Diagram and Simulation Results 

In Fig. 2 is presented the Simscape diagram for a hydraulic system with 

secondary control and fixed – displacement pump. The input signal represents 

the angular velocity of the hydraulic motor. It was considered the signal 

presented in Fig.1. 

The Simscape diagram has as subsystems the servocylider subsystem and 

the load subsystem. It has also some sensors and oscilloscopes for important 

parameters. 

Fig. 1 – Input signal. 

Fig. 3 shows the diagram for the servocylinder subsystem and Fig. 4 the 

diagram for the load at the axle of the variable displacement motor. For both 

cases the load is a constant.

Fig. 2 – The Simscape diagram for system with fixed - displacement pump. 

Signal 1 

Signal Builder2 

� �� 

�� 

4-Way Directional 

Valve 

A 

� � �� 

S 

Proportional and 

Servo-Valve Actuator 

P 

B 

T 

PS S 

PS2S2 Position_Valve 

Gas-Charged 

Accumulator1 

R 

C 

V 

P 


Motion Sensor 

HR2 

-C- 

Constant 

Solver 

Configuration 

f(x)=0 

MTR1 

B 

S PS 

S2PS 

A 

Flow Meter1 

Ideal Angular 

Velocity Source 

S 

PS S 

PS2S1 

Angular 

velocity 

R 

Position_SP 

Hydraulic Fluid 

C 

� �� 

MRR 

S PS 

PS2S 

PRV 

A B 

Flow Meter 

Flow 

Meter2 

HR 

A 

B 

A 

B 

A 

C 

B 

MRR1 

A 

S 

C 

W 

R 

A 

Ideal Rotational 

Motion Sensor 

B 

P 

Ideal Hydraulic 

Pressure Sensor 

HR1 

PS S 

PS-Simulink 

Converter1 

-C- 

Constant1 

�� 

Pump_pressure 

�� 

Bul. Inst. Polit. Iaşi, t. LVIII (LXII), f. 1, 2012 187


D 

MTR3 MTR2 

C 

A 

R 

B 

Servo-cylinder 

R C 

M 

C 

TS 

R � 

Fig. 

�� 

3 

�� 

– Subsystem servocylinder. Fig. 4 – Subsystem "load". 

�� 

In Fig. 5 one can see the position of the spool of the servovalve, the 

pressure at the pump outlet, the angular velocity at the motor axle and the flow 

rate at the motor inlet for the case of the system with fixed – displacement 

pump. 

A �� 

P 

Winch 

a b 

c d 

Fig. 5 – For the system with fixed displacement pump: 

a – position of the servovalve; b – pressure at the pump outlet; c – angular velocity of 

the hydrostatic machine; d – flow rate at the motor. 

For the hydraulic system with secondary control and variable – 

displacement pump, was also done the Simscape diagram. The input signal is 

presented in Fig. 1. In Fig. 6 one can see the position of the spool of the 

servovalve, the pressure at the pump outlet, the angular velocity at the motor 

M D 

R 

C 

S PS 

S2PS1 

MTR 

S 

R 

C 

Weight 

MTR1


axle and the flow rate at the motor inlet for the case of the system with variable 

– displacement pump. From the diagram in Fig. 5c, one can see the dynamic 

parameters of the system: time constant, error and damping ratio 

a b 

c d 

Fig. 6 – For the system with variable – displacement pressure – compensated pump: 

a – position of the servovalve; b – pressure at the pump outlet; c – angular 

velocity of the hydrostatic machine; d – flow rate at the motor. 


1. The hydraulic systems with secondary control are not only saving 

energy but also recovering it. 

2. The study of dynamic behaviour of such systems is possible using the 

Matlab/Simscape method. In this paper the Simscape diagram for a hydraulic 

system with secondary control and fixed displacement pump is presented. 

3. Comparing the two cases, one can see that the angular velocity at the 

hydrostatic machine follows the input law. Important differences are to be 

noticed regarding the pressure at the pump outlet and the flow rate at the 

hydromotor. 

Acknowledgements. The present work has been supported from the Grant PNII, 

2703/22-111/2008. 

REFERENCES 

Backé W., What Are the Prospects Facing new Ideas in Fluid Power? The Sixth 

International Conference on Fluid Power Transmission and Control (ICFP' 2005), 

Hangzou, China, 5-8, 2005, pp. 21-30.


Kirka A., Investigation of Parameters of Accumulator Transmission of Self Moving 

Machine. Engineering for Rural Development, 28-29.05.2009, Jelgava, 2009, pp. 

100-104. 

Murrenhoff H., Hydraulic Drives in Stationary Applications. 5th International Fluid 

Power Conference, Aachen, Germany, 2, 2006, pp. 11-36. 

Rydberg K.-E., Hydraulic Accumulators as Key Components in Energy Efficient Mobile 

Systems. The Sixth International Conference on Fluid Power Transmission and 

Control (ICFP' 2005), Hangzou, China, 2005, pp. 178-182. 

Mathworks Inc., on line at www.mathworks.com 

*** SimHydraulics: http://www.mathworks.com/products/simhydraulics/demos.html 

ASPECTE PRIVIND SISTEME HIDRAULICE CU REGLARE SECUNDARĂ 

(Rezumat) 

Sistemele hidraulice cu reglare secundară au avantajul că realizează nu doar 

economie de energie ci şi recuperarea acesteia în fazele de frânare. În acest mod se 

explică atenŃia de care se bucură atât pentru cercetători cât şi pentru utilizatori. Pentru 

analiza comportrii dinamice, utilizarea mediului Matlab, respectiv metodei Simscape, 

este potrivită în cazul sistemelor hidraulice în general şi implicit a celor cu reglare 

secundară. În lucrare este prezentată schema funcŃională Simscape pentru un sistem 

hidraulic cu reglare secundară şi pompa cu volum geometric constant. Sunt prezentate 

rezultatele simulării pentru acest caz ca şi pentru cazul unui sistem hidraulic cu reglare 

secundară şi pompă cu volum geometric variabil şi regulator de presiune. Se remarcă în 

ambele cazuri obŃinerea unei variaŃii a vitezei unghiulare la axului motorului hidraulic 

apropiata de cea impusă. DiferenŃe între cele două cazuri apar în ceea ce priveşte varia- 

Ńia debitului trimis către motorul hidraulic şi variaŃia presiunii la ieşirea din pompă.


Publicat de 



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MAN MACHINE INTERFACE ERGONOMIC DESIGN 

RELATED TO HUMAN FACTORS 

BY 

ALEXANDRA TOMA * and CORNEL CIUPAN 

Technical University, Cluj-Napoca, 

Department of Machine Tools and Industrial Robots 

Received: September 10, 2011 


Abstract. When a human-machine-interface gets less priority in a design analysis, 

unnecessary hazards may be imposed. Ergonomic and human factors should be 

incorporated as a part of this process, so they can show the importance of human machine 

interaction. The objective of this paper is to highlight these factors that are able to provide 

a platform where this emerging technology part interaction can increase productivity, 

quality, satisfaction, safety and health in a workplace. The article contribution is a 

theoretical one taking into consideration helpful ideas in a human machine interface 

ergonomic design. 

Key words: design, ergonomics, human factors, human-machine-interface. 


The rational efficient utility of means of production depends on the use of 

the most active and most dynamic element which is the human, who participates 

with his whole being and consists of physical, psychical and moral complexity. 

Therefore, beyond the human activity results, man-machine interaction strongly 

manifests. The mutual conditioning between the man-machine interaction is 

created by human, technological, physical and psycho-social elements which 

interconnect themselves in a common network, leading them to the same 

purpose. The use of machines keeps growing because the work system gets 

increasingly based on the socio-technical element and the human-machine 

becomes more prevalent in all domains of activity, says (Ispas et al., 1984). 

Sometimes humans and machines are equal partners and they should be treated 

* Corresponding author: e-mail: ines_design3@yahoo.com

192 Alexandra Toma and Cornel Ciupan 

that way and described as equal terms. The right contradiction idea is that 

humans must not be described as they are machines, neither should machines be 

described as if they were human, explains (Hollnagel, 2008). The human is a 

part of a process system and also can be a part of a larger production system. In 

addition of ergonomic activities, human factors benefits get incorporated. The 

operator using a machine is, at least one single level of a man-machine system 

that can be described. The paper is divided so as to present important ideas 

about human machine interaction system, concerning its ergonomic design. 

2. Man Machine System 

This quote given by (Salvendy, 2008) says: “Although it can refer to any 

type of interface device, the term human machine interaction usually refers to a 

display, a computer and software that serve as an operator interface for a 

controller or a control system. Ergonomics is trying to assemble information on 

human capacity and capability in order to use information in equipment 

designing and create a larger system. Both, human capacity and capability, 

describe man-machine interaction trying to reduce unnecessary stress 

concerning demands of the task being performed. A discrete category added to 

this fragment will explain that a design analysis will always depend on the bond 

between user and machine. Interfaces can be more or less similar to one 

another. The way they are designed or redesigned must enable the user to do 

reliable operations depending on elements that he sees, touches, hears, in order 

to perform control functions and receive feedback on the actions. Fig. 1 shows 

how this category finds its place as a benefit added to a human machine 

interaction. The operator can be put in charge as follows: 

Fig.1 – Conditions that put an operator in charge. 

The display is one of the most important elements that belongs to a 

human machine interface and is the first that realizes the actual communication 

between machine and user, and from here the operator knows if it is a good or a 

wrong one. While he makes an error, an onscreen message can be displayed, 

Fig. 2. This message can have no consequence over any of the tasks performed. 

Fig. 3 shows the way how an onscreen message display should look like. 

Any impact between a machine and an operator will always be for the 

first time visual and that is how communication between them becomes 

important. Explanations will be shown in Fig. 4. Whenever an interface change


is performed and there is a change in the command function, mapping the new 

interface, the designer should share as few features as possible with the former 

one. Operator shows a higher level of performance when the same commands 

are linked to the same functions, explains Smith S. E. in (Smith, 2011). 

Fig.2 – Error message display. 

Fig. 3 – A good display look should highlight. 

Fig. 4 – User interface visual communication.


When consistency cannot be achieved, for example the cost, a designer 

should consider is the next recommendation. The operator may disregard 

warning notices and signals because factors like stress and lack of vigilance can 

occur. For example, a safety device can be required from operator before he 

uses a given function. A good functionalism goal discovers how user behavior 

adapts to contains imposed by the domain for example, a machine program 

getting available. A good human machine interface design leads to individual 

cognition of every specific task centered on man machine interaction. 

Applications decision making to complex settings and leads to observing 

operators at work and incident studies and assumptions about human decision 

making can be formed. In the following chapter the environment shows its 

involvement when a human machine interface takes into account the human fit. 

3. Environment 

The environment from a system point of view is that which provides 

inputs to the system and which reacts to outputs from the system. The 

environment around the operator or user-crew ensemble however, is not only 

the atmosphere and weather, temperature around it or luminosity, but also the 

machine working program management system. 

Equipment design and environmental comfort and safety become 

facilities for human use and ergonomics is the supplement added to these 

operation facilities. 

a b 

Fig.5 – Schematic presentation: a − computer numeric control equipment; 

b − working panel. 

Human factors are shown in Fig. 6, as being a main incorporated system 

design part and the other Fig. 7, highlights what a human computer interaction 

provides.


Fig.6 – Human factors. 

Fig 7 – Human computer interaction. 

When it is about man machine interface function and design the main 

purpose is to make its user friendly, increasing market share and operator 

satisfaction (Flaspöler et al., 2004). 


1. The three dimensional sketch shown in the previous chapter represents 

a schematic example for a Computer Numeric Control machine where the 

working panel is the main important part that draws author attention concerning 

its design in order to improve the communication between machine and human. 

2. The paper shows how important is the human factor when it is about to 

design an interface. In the same time, from the lines above, synthetically results 

that the author has the strong desire to improve the ergonomics and the 

functionality of machine tools. 

3. This study is directed to an equipment with numerical command or a 

group of equipments with numerical command, namely on vertical milling 

tools. We suggest adapting a panel that is traditionally used for another type of 

machine tool, on vertical milling tool. 

4. It is pointed out the author concern to correlate the innovating 

intentions with the active STAS. It is emphasized the impact on improving the 

equipment ergonomics and functionality in the relation man – machine and the 

professional environment. It is also underlined the author interest in making a 

guide very necessary for the machine tool designers concerning the ergonomics 

and functionality. 

Acknowledgements. This paper was written at the Technical University of Cluj- 

Napoca and supported by the project “Doctoral studies in engineering sciences for 

developing the knowledge based society - SIDOC” contract no. 

POSDRU/88/1.5/S/60078, project co-funded from European Social Fund through 

Sectorial Operational Program Human Resources 2007-2013.


REFERENCES 

Flaspöler E., Hauke A., Reinert D, The Human-machine Interface as an Emerging Risk. 

EU-OSHA – European Agency for Safety and Health at Work (Sarodnik & Brau 

2006), (Schmersal 2005), (Montenegro 1999), (Koller, Beu & Burnmester, 

2004). 

Hollnagel E., Prolegomenon to Cognitive Task Design. Lawrence Erlbaum Associates, 

Inc, 2008. 

Ispas C., Ionescu I. A., Lazarovici V. A., Andriescu N. T., Machine-Tools Ergonomy. 

Ed. Tehnică, 1984. 

Salvendy G., Human factors and Ergonomics. Ed., by Erick Hollnagel, 2008. 

Smith S. E., Wise GEEK, Ed. by O. Wallace, January, (Tutherow & Lipták, 2002), 

(Baumann and Lanz, 1998), (Charwat, 1992), Internet address, 2011. 

FACTORII UMANI ASOCIAłI DESIGNULUI ERGONOMIC AL 

UNEI INTERFEłE OM-MAŞINĂ 

(Rezumat) 

Designul unei interfeŃe om-maşină va fi tot timpul important şi continuu 

îmbunătăŃit. Factorii umani trebuie luaŃi în considerare obligatoriu, pentru că omul este 

elementul principal, care le raspunde pozitiv sau negativ. 

AtenŃia autorilor a fost îndreptată asupra comunicării vizuale dintre om şi 

maşină, deoarece,acesta este primul si cel mai important lucru, care realizează contactul 

dintre om şi tehnologie. Pe viitor, dorinŃa autorilor este aceea de a îmbunătăŃi aspectul 

unui panou de comandă şi interfaŃa unui soft care aparŃine unei maşini cu comandă 

numeric; în speŃa, mărirea dimensiunilor display-ului şi poziŃionarea ergonomică a 

butoanelor, Ńinând cont de standardele de proiectare. Scopul acestui articol este de a 

ajuta proiectanŃii să respecte dorinŃa operatorului uman.


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STUDY ABOUT THE ROLE OF ACCUMULATORS IN 

HYDRAULIC SECONDARY CONTROL SYSTEMS 

BY 

LILIANA TOPLICEANU * , ADRIAN GHENADI and MARIUS PASCU 



”Vasile Alecsandri” University, Bacău, 

Engineering Faculty 

Abstract. The paper is focused on the functional role of accumulators in 

hydraulic secondary control systems. Mathematical models are presented and 

the same the simulation results for a gas-charged accumulator using 

SimHydraulics tool from programming language Matlab. 

Key words: accumulator, secondary control system. 


The concept of secondary control of the hydraulic systems working under 

impressed pressure entered the current use around 1980. The secondary control 

is mostly used where the conventional systems are no longer able to meet the 

technical requirements in terms of feedback dynamic, positioning and accurate 

control of the speed. The use of these systems is still rather reduced. 

2. The Role of the Accumulator in the Secondary Control Systems 

2.1. The Working Principle of the Secondary Control 

The concept of secondary control working under impressed pressure 

supposes the existence of a secondary unit of adjustable capacity which, when 

working at a preset pressure, tries to reach the adequate capacity in order to 

maintain a constant rotation speed. An adjustment of the capacity under 

constant pressure takes place, so that the rotation speed of the motor should 

* Corresponding author: e-mail: lili@ub.ro

198 Liliana Topliceanu et al. 

remain constant (or approximately constant). This implies at the same time an 

adequate adjustment of the hydraulic momentum, so that the mechanical 

momentum required by the functional program of the system could be adjusted 

The main advantage of these systems is the significant increase in 

efficiency. The adjustable and reversible hydrostatic units will take from the 

system only the power necessary for overcoming a certain momentum, or they 

will supply it with power when they start working as a pump (Kordak, 2003). 

The second important advantage is a completely new method of 

centralized power supply, similar to the electric systems, supplying several 

consumers with different loads from the same network (Fig. 1). 

Fig.1 – Hydraulic system with secondary control (Kordak, 2003). 

Through some hydraulic connectors, we can link to the main power 

supply network as many motors as are necessary for that specific system. There 

are no hydraulic resistors on the power supply lines. The pressure within the 

system is determined by the loading conditions of the accumulator. 

When the actuators work as a motor, the power is taken from the power 

supply network, and when they work as a generator, the power is reintroduced 

into the system and it can be used by other motors, it can be stored or can be 

used for producing other types of energy such as electric energy. We must 

notice that, when the pressure stays quasi-constant, the compression-dilatation 

effect of the liquid is eliminated, the system dynamic is much better as 

compared to the classical systems and the power efficiency is much higher. 

Designing such systems implies: setting the pressure value necessary in 

the centralized power source, calculating the range of resistance momentum 

which can be overcome by adjusting the displacement of the motors, 

determining the influence of transitory conditions on the functional stability of 

other consumers connected to the network, necessary number of accumulators. 

2.2. Hydraulic Accumulators 

The general function of these hydraulic elements is, as their name 

suggests, to accumulate the hydraulic energy during the first stage, and then to 

introduce it into the system in order to meet the energy demand or to control the


occurrence/development of certain specific functional phenomena. The 

accumulator charge sets the value of the pressure in the classical systems as 

well as in the more modern ones, which have secondary control. 

Hydraulic accumulators are used for: maintaining a relatively constant 

pressure within the system when the pistons of the hydraulic motors change 

their position; attenuating the flow pulsation of the pump; attenuating the peak 

pressure in the system; compensating for the volume fluctuation of the liquid 

when the pressure or the temperature in the system change; compensating the 

oil loss, within certain limits; recovering the breaking energy. 

From the large range of designed versions, the most frequently used are 

the gas-charged ones. They are made up of two distinct volumes, one of them 

containing gas under pressure, the second one being filled with the working 

liquid in the hydraulic system. Due to the compressibility of the gas in the 

accumulator, the container filled with gas has the role of elastic element. 

3. Modeling of Gas-Charged Accumulator 

The accumulators must have a minimum total volume and a maximum 

useful capacity. The useful capacity refers to the volume of the oil output for a 

pressure loss ∆p = pmax − pmin. These values, the maximum drive pressure and 

the minimum pressure, are given as preset data in choosing or designing of an 

accumulator. In the case of the secondary control hydrostatic systems, the 

transitory loading or unloading modes of the accumulator occur at high speed, 

in quick cycles. 

3.1. Dimension of the Accumulator 

As it was mentioned before, the most frequently used accumulators are 

the gas-charged ones. Taking into consideration this type of accumulator, the 

version with elastic chamber, its functioning is based on the following wellknown 

law 

n 

iVi 

f 

n 

f 

p = p V , (1) 

where pi and pf are the initial and the final pressure of the gas; Vi and Vf, the 

initial and the final gas volume; n = cp/cv is the polytrophic index; cp/cv - the 

specific heat of the gas under constant pressure, and the constant volume 

respectively. The value of the index n depends on the type of process, more 

exactly on its speed. In general, it is an adiabatic transformation if the process 

takes place in less than a minute (Oprean et al., 1982) and it is specific to the 

accumulators with membrane and elastic chamber. The accumulator volume 

necessary within a hydrostatic system is calculated using the following formula 

(Oprean et al., 1982).


1/ n 

⎛ p3 

⎞ 

⎜ ⎟ 

p 

V 

⎝ ⎠ 

Vx 

, (2) 

1 

1 = 

1/ n 

⎛ p3 

⎞ 

1− 

⎜ ⎟ 

p 

⎝ 2 ⎠ 

where: Vx is the volume to be discharged from the accumulator; p1 is the preload 

pressure in the accumulator; p2 is the maximum working pressure of the 

installation (for which there is a corresponding volume V2 of compressed gas); 

p3 is the minimum pressure in the installation. Setting the pre-load pressure is 

important from the point of view of accumulator efficiency in maintaining the 

desired value within the system. 

In the case of a polytrophic transformation, the optimal value of the preload 

pressure can be set using the formula indicated in the specialized literature 

( 1−1/ 

n) 1/ 

n 

2 p3 

p = p . (3) 

1 

If the discharge is an adiabatic one, the equation becomes: p1 = p2 0,291 p3 0,709 . In 

the case of a polytrophic process with n=1,25 then the Eq. (3) has the following 

content: p1=p2 0,200 p3 0,800 . 

The discharge time of the accumulator is an extremely important index, 

since it determines the type of process taking place and thus sets the value of the 

index n. The total volume of the accumulator Vt is calculated by adding the 

volume filled with gas V1 and the one filled by oil Vu. The volume filled with 

gas is usually a multiple value of the oil volume in full charge mode. The rate of 

these volumes in the common industrial applications is usually V1/Vu= 3/1. 

3.2. Mathematic Model of Accumulator 

Simplifying one can say that accumulator absorbs or releases a discharge 

Qac when there is a variation of the pressure p within the system. The 

mathematic model of accumulator without connecting pipe is given by equation 

(Călăraşu, 1999), (Oprean et al., 1982) 

Q 

ac 

V0 dp 

= . (4) 

E dt 

The model of the accumulator with bellows is given by equation of 

continuity and adiabatic gas transformation equations 

Q 

ac 

V 

= + + 

dt E dt E dt 

dV 0 dp V d p 

, 

χ χ 

= 1 1 . 

pV p V 

(5)


Working on the two equations we obtain 

Q 

ac 

⎛ 1/ χ 1/ χ 

V1 p1 V1 V1 p ⎞ 

1 dp 

= ⎜ + − ⎟ . (6) 

⎜ E 1/ χ E ( χ + 1)/ χ 

p np ⎟ 

⎝ ⎠ 

dt 

4. Simulation of Gas-Charged Accumulator 

The study of accumulator behavior in different operating conditions can 

be made using specialized software. One of the most frequently used, because 

of its flexibility, is SimHydraulics from Matlab. 

Fig. 2 – Accumulator testing circuit. 

In Fig. 2 one can see the scheme for functional tests of an accumulator, 

using predefine units of Matlab program. 

Simulation results are presented in Fig. 3. The accumulator is preloaded 

at 0.5MPa. The pressure source is switched at every 3s from 1MPa to 0,5 MPa 

but the accumulator performs its function and maintains the pressure and 

discharges a corresponding flow. The input signals are presented in Fig. 4.


Fig. 3 – Simulation diagrams. 


Fig. 4 – Input signals. 

1. Accumulator presence is required in the construction of drive systems 

with secondary control in order to stabilize the value of the pressure. The 

reaction time of accumulator affects the stability of motion of driven equipment. 

2. The type and characteristics of accumulator must be chosen in 

accordance with needed parameters of the system. 

Acknowledgements. The present work has been supported from the Grant 

(CNCSIS) PNII 2703/22-111/2008. 

REFERENCES 

Călăraşu D., Reglarea secundară a sistemelor de acŃionare hidrostatică în regim de 

presiune cavsiconstantă. Ed. Media-Tech, Iaşi, 1999. 

Marin V., Moscovici R., Teneslav V., Sisteme hidraulice de acŃionare şi reglare 

automată. Ed. Tehnică, Bucureşti, 1981. 

Oprean A., Ionescu Fl., Dorin Al., AcŃionări hidraulice Elemente şi sisteme. Ed. 

Tehnică, Bucureşti, 1982. 

Kordak R., Hidrostatic Drives with Control of the Secondary Unit. Rexroth Bosh 

Group, Vol. 6, 2003. 

STUDIU DESPRE ROLUL ACUMULATOARELOR IN FUNCłIONAREA 

SISTEMELOR HIDRAULICE CU REGLAJ SECUNDAR 

(Rezumat) 

Lucrarea realizeză o sinteză privind importanŃa acumulatoarelor în funcŃionarea 

sistemelor hidraulice cu reglaj secundar lucrând la presiune cvasi-constantă. Un 

exemplu de simulare funcŃională a acumulatorului, importantă mai ales în regimurile 

tranzitorii, este de asemenea prezentată.


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CONSIDERATIONS ABOUT CONTROL CHARTS USED IN 

STATISTICAL PROCESS CONTROL 

II. CONTROL CHARTS FOR INFREQUENT 

AND CUMULATIVE DATA 

BY 

CĂTĂLIN UNGUREANU ∗1 , IRINA COZMÎNCĂ 1 and RADU IBĂNESCU 2 



”Gheorghe Asachi”Technical University of Iaşi, 


2 Department of Theoretical Mechanics 

Abstract. Processes quality control is very important for organizational 

success and maintaining it on market. SPC is a reliable and economical method of 

monitoring the evolution of a productive process. In industrial practice are 

encountered cases when classical control charts, Shewhart type, are not 

recommended. The purpose of this paper is to analyze control charts using in 

these cases. Special attention is given to the EWMA and CUSUM charts. 

Key words: quality control, control chart, EWMA chart, CUSUM chart. 

1. Charts for Infrequent Data 

In many practical situations encountered in industry and services, 

available data are not on a large scale. There may be instances where a new set 

of data is rarely available, only once per cycle. Such cases include batches of 

chemicals, large and complex shapes castings, economic indicators monitored 

in time. SPC techniques can be successfully applied even in these cases, when 

literature does not recommend Shewhart control charts using (Oakland, 2003), 

(de Vargas et al., 2004), (Han et al., 2010). 

∗ Corresponding author: e-mail: cungurea@yahoo.com

204 Cătălin Ungureanu et al. 

1.1. Individual or Run Charts 

The simplest control chart for variable is that for individual 

measurements. In this case individual data values will be plotted, instead of 

sample’s averages. The central line represents the specification central value, 

past performance average or target value. Action lines are placed at three 

standard deviations from the center line. Warning lines can be traced to two 

standard deviations from the center. 

Individual or run charts have the advantage of simplicity and show 

changes in process alignment and accuracy. The limits of these charts are 

related to reduce sensitivity to detect small changes. The literature recommends 

the use of these charts, sometimes with different amplitudes diagrams, with 

clearly superior benefits in process control to a simple data table. 

1.2. The Zone Control Chart and Pre-Control 

This is an adaptation of the individual chart. Compared to the action and 

warning lines, two lines are placed at one standard error from the average. Each 

point marked on the sheet corresponds to a certain score, depending on the band 

in which it falls. A change occurred in the process if the cumulative score 

exceeds a certain value between two crossings of the line for average. Another 

pre-control chart variant divided into four allowable tolerance zones, two areas 

of green plants and two yellow zones, peripheral. Areas outside the 

specifications are considered red areas. The initial five consecutive 

measurements ensures that the process is under control and demonstrate a 

minimum capability CPK = 1.33. Then periodically two units are review, and 

depending on the areas where it fall the process variability evolution is 

concluded. 

Pre-control charts are simple and useful in many applications, but still 

suffering because of low sensitivity (Oakland, 2003). 

1.3. Control Charts with Moving Mean and Moving Range 

In practical industrial applications often meet situations where the output 

of a process are available at a relatively large time, from technological or 

specific economic reasons, related to the high cost of tests necessary to evaluate 

the quality characteristics. In these cases, the forming of samples sufficient to 

allow the use of conventional control charts is not possible and the process 

control can be delays with adverse consequences. 

In these situations charts for moving mean and range are recommended. 

The emergence of a new data leads to the forming of a new sample by removing 

the oldest set of information. This technique has several advantages: every data 

represent two points on two different charts, which means different things, 

isolated data has not react and less likely override, the effect on the process will 

be a calming.

Bul. Inst. Polit. Iaşi, t. LVII (LXII), f. 1, 2012 205 

Compared to individual files, moving average charts have a smoothing 

effect on the results, so that trends and changes can be seen more easily. 

Smoothing effect is greater, the greater the number of sample, but at the same 

time, highlighting trends is delayed. Another use of this technique is forecasting 

activities. 

1.4. Exponentially Weighted Moving Average Charts (EWMA) 

Some researchers have developed a type of charts which does not give 

equal importance to the newest data. These are EWMA charts, where the 

average is calculated with 

( 1− ) X −1 

λ . (1) 

X i = xi 

+ λ i 

x result, 1 

In Eq. (1) i X is the mean after the i X i− 

is the previous 

average, and λ is the smoothing constant, with values in the interval (0, 1). 

Many studies have been conducted to determine the influence of λ size on the 

sensitivity of EWMA charts. Most researchers (de Vargas et al., 2004), (Han et 

al., 2010), (Shu et al., 2007), (Borror et al., 1998), (Friker et al., 2008), 

recommend λ values in the range of 0.05…0.25. 

For example it is considered a manufacturing process of an adhesive, 

whose tracked quality characteristic is viscosity V[cSt]. Due to the peculiarities 

of the process, one can obtain a single value of viscosity on the day. Data for 

viscosity and the averages MA calculated with Eq. (1) ( λ =0.2) are presented in 

Table 1. 

Table 1 

Experimentals 

days V[cSt] MA days V[cSt] MA days V[cSt] MA 

1 59.1 59.82 5 61.8 60.2 9 60.5 59.69 

2 60.5 59.96 6 57.5 59.66 10 62.4 60.23 

3 54.3 58.63 7 56.2 58.97 11 58.2 59.83 

4 63.7 59.8 8 61.6 59.49 12 56.9 59.24 

In Fig.1 are presented individual and EWMA control charts. Control lines 

are determined with Eqs. (2) (de Vargas et al., 2004), 

λ 

AL = µ 0 ± 3σ 

2 − λ 

, 

WL = µ ± σ 

0 2 

CL = µ 0 . 

λ 

2 − λ 

, 

From previous experiments µ 0 = 60 and σ = 3.301 are known. 

(2)


Fig.1 – Control charts, 1-individual, 2-EWMA. 

2. Charts for Cumulative Data 

2.1. Cumulative Sum (CUSUM) Charts 

The technique of CUSUM charts was developed in the UK in 1950s and 

is one of the most powerful tools available to detect small changes in trends and 

data management (Shu et al., 2008). Compared to previous models, these charts 

take into account the entire volume of available data. 

Table 2 

Experimental data 

Sample Mean Cusum Sample Mean Cusum 

number X [mm] score Sk number X [mm] score Sk 

1 22.25 -0.25 11 22.35 -0.85 

2 22.55 -0.20 12 22.55 -0.90 

3 22.40 -0.30 13 22.40 -1.10 

4 22.65 -0.15 14 22.60 -1.10 

5 22.70 0.05 15 22.50 -0.90 

6 22.55 0.10 16 22.65 -1 

7 22.60 0.20 17 22.40 -1.35 

8 22.30 0 18 22.15 -1.60 

9 22.10 -0.40 19 22.35 -1.75 

10 22.20 -0.70 20 22.40 -1.85 

For each sample, the quality characteristic mean is determined and is 

compared with the target. Cusum score of sample k is then calculated by


S 

k 

∑ 

i= 

1 

= k 

( x − t) 

i 

. (3) 

In Eq. (3) x i is the i sample result and t is the target value. 

An example with some machined parts is shown in Table 2. The target 

value was 22.50 mm. 

The mean control chart is shown in Fig. 2. 

Fig. 2 – Mean control chart. 

Control limits were calculated using known relationships (Oakland, 

2003). Note that samples 5, 9 and 18 required an intervention to bring the 

process under control. 

Cusum chart is shown in Fig. 3. Sk score values are then represented 

graphically, the slope signifying the development process diagram. An upward 

slope means that the process is above target, a downward slope means that the 

process is below target and a horizontal portion of the development means 

exactly the desired value. 

The scale ratio is determined by the rapport of the distance between 

corresponding samples points on the abscissa and the double means standard 

error, determined by Eq. (4) (Oakland, 2003), on the ordinate. It is 

recommended that this ratio is between 0.8 to 1.5 (Oakland, 2003), 

σ 

SE = (4) 

n


Fig. 3 – Cusum chart. 

Fig. 4 – Manhattan chart. 

To see if the process is under statistical control compares the slope with 

this gradient. A slope exceeding the limit means the occurrence of special 

causes of process variability. The general appearance of the chart signify 

process tend to perform below the target. 

2.2. Manhattan Charts 

This chart is the particular case of cumulative sheet, showing successive 

stages in process evolution. Thus related products with slightly different 

features can be separated and can be delivered to customers with varying 

requirements.


An example is shown in Fig. 4. 

Based on mean control and Cusum charts analysis, have been identified 

several phases in process evolution (samples 1-3; 4-7; 8-11; 12-16; 17-20). For 

each average was determined and represented graphic. 


1. Using the most suitable type of control charts for the process be kept 

under control should be a concern for the quality control department and for the 

entire administration. 

2. EWMA charts are used when available data about the process arriving 

at a relatively high periods of time. It shows the general evolution of the process 

and has a calming role, to prevent process override. 

3. Cumulative charts highlight the general tendency of the process. They 

can be used to separate the different developmental stages of the process. 

REFERENCES 

Borror C. M., Champ C. W., Rigdon S. E., Poisson EWMA Control Chart. Journal of 

Quality Technology, 30, 352-361 (1998). 

de Vargas V. C., Lopes L. F. D., Souza A. M., Comparative Study of the Performance 

of the CUSUM and EWMA Control Charts. Computers & Industrial Engineering, 

46, 707–724 (2004). 

Friker R. D., Knitt M. C., Hu C. X., Comparing Directionally Sensitive MCUSUM and 

MEWMA Procedures with Application to Biosurveillance. Quality Engineering, 

20, 478-494 (2008). 

Han S. W., Tsui K. L., Ariyajunyab B., Kimb S. B., A Comparison of CUSUM, EWMA, 

and Temporal Scan Statistics for Detection of Increases in Poisson Rates. Quality 

and Reliability. Engineering International, 26, 279-289 (2010). 

Oakland J. S., Statistical Process Control. Butterworth Heinemann, Oxford, 2003. 

Shu L., JIang W., Wu S., A One-sided EWMA Control Chart for Monitoring Process 

Means. Communications in Statistics—Simulation and Computation, 36, 901-920 

(2007). 

Shu L. J., Jiang W., Tsui K. L., A Weighted CUSUM Chart for Detecting Patterned 

Mean Shifts. Journal of Quality Technology, 40, 194-213 (2008). 

CONSIDERAłII PRIVIND FIŞELE DE CONTROL UTILIZATE ÎN CONTROLUL 

STATISTIC AL PROCESELOR 

II. Fişe de control pentru date rare şi cumulative 

(Rezumat) 

Cea mai simplă cale de atingere a obiectivelor organizaŃiei din domeniul calităŃii 

este învăŃarea şi aplicarea de către toŃi salariaŃii a instrumentelor simple de control şi 

asigurarea a calităŃii, în special a diferitelor tipuri de fişe de control. Fişele EWMA se


folosesc atunci când sunt disponibile relativ puŃine date despre proces. Ele au un efect 

de calmare şi împiedică suprareglarea procesului. Fişele CUSUM prezintă tendinŃa 

generală de evoluŃie a procesului. O variantă a acestora poate separa diverse stadii de 

dezvoltare din proces.


Publicat de 



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FUNCTIONAL DIAGRAMS FOR MODELING 

THE RESERVOIR EMPTYING PROCESS 

THROUGH A SMALL EMPTYING ORIFICE 

BY 

DĂNUł ZAHARIEA * 

“Gheorghe Asachi” Technical University of Iaşi 




Abstract. In this paper the emptying process of a reservoir with constant 

cross-sectional area through an emptying small orifice will be investigated. Two 

types of reservoir will be considered: with constant head and with variable head. 

The pressure at the water surface and also the velocity of this surface, as well as 

the flow discharge coefficient of the emptying orifice will be considered. The 

fundamental equations governing the reservoir emptying process will be 

presented (the analytical method). For modeling and simulating the reservoir 

emptying process two functional diagrams based on MATLAB/Simscape 

computational approach suited to analyze the real physical systems consisting of 

real physical functional blocks will be presented (the functional diagrams 

method). The functional diagrams allow obtaining the graphical representation of 

the level of the liquid from the reservoir during the time of the emptying process 

but also the graphical representation of the flow rate of the liquid through the 

orifice. Using the measured values of flow rate, the velocity and the Reynolds 

number can be calculated. A comparative analysis of the numerical results for the 

emptying time for constant head and for variable head reservoirs, obtained with 

the analytical and the functional diagram methods will be presented. 

Key words: constant cross-sectional area reservoir, emptying time, emptying 

orifice, functional diagram, MATLAB. 

* e-mail: dzahariea@yahoo.com

212 DănuŃ Zahariea 


Safe operating procedures of the hydraulic systems equipped with 

reservoirs must consider, from the point of view of the industrial process 

management, two principal conditions: the volume of liquid inside the reservoir 

must be high enough for uninterrupted supplying the industrial process 

(emptying or consumption phase); the storage capacity of the reservoir must be 

high enough for a safety liquid volume storage thus the continuity of the 

industrial process being achieved (filling or supplying phase). The liquid 

volume inside the reservoir must be known on every moment of the industrial 

process, the more that for certain periods of time the two working phases can 

take place simultaneously. From this point of view three important elements can 

be mentioned: the supplying system, the emptying system and the reservoir. 

Reservoirs can have constant or variable cross-sectional area and constant or 

variable head. The emptying system can have an emptying orifice or an 

emptying pipe. In this paper the emptying process of a constant cross-sectional 

reservoir using an emptying orifice placed on the bottom side of the reservoir 

will be analysed using two methods: the analytical method, (Bartha, 2004), 

(Idelcik, 1984), (Panaitescu & Cacenko, 2001), based on the fundamental 

equations governing the emptying process and the functional diagrams method, 

(Zahariea, 2010), (Mathworks), based on MATLAB/Simscape computational 

approach to analyze the real physical systems. 

2. Analytical Method 

Let us consider a vertical reservoir with variable cross-sectional 

z z ∈ 0, 

H where H is the reservoir maximum level, Fig. 1a. The 

area Ar ( ) , [ ] 

emptying orifice has the section o A , the local head loss coefficient ζ o and the 

flow discharge coefficient µ o . The flow through the orifice generates a 

contraction phenomenon characterized by the contraction coefficient 

ε = Ac Ao 

. The liquid volume variation produced during the emptying time dt 

by level change from z to z − dz 

will be equal with the liquid volume flowing 

through the emptying orifice in the same time dt. 

Assuming A 

2 

= πD 4 = ct. (Fig. 1b) and considering the conditions t = 

r r 

= 0 ⇒ H = Hi 

and f H H T t = ⇒ = , where H i and H f are the initial and the 

final liquid levels, the emptying time can be expressed by 

A 

T = 

A 

Hi 

r 1 

dz 

∫ 

c H 

f 

v 

( z) 

, (1) 

where: v( z ) is the liquid velocity through the emptying orifice which will be 

obtained starting from the Bernoulli’s equation for incompressible flow between 

characteristic sections 1 and 2.


a b 

Fig.1 – Reservoir with emptying orifice: characteristics: 

a – variable cross-section reservoir, b – constant cross-section reservoir. 

Different expressions can be obtained depending of the reservoir type: 

i) for constant head reservoir, the emptying time is 

T 

c 

A 

Hi − H 

r 

f 

= 

⎡ 2 

A 

p1 − p 

⎛ o ⎞ ⎤ 

µ 2 1 

Hi 

o Ao g ⎢ − 

+ 

⎜ ⎥ 

A 

⎟ 

γ 

⎢⎣ ⎝ ⎠ ⎥⎦ 

ii) for variable head reservoir, the emptying time is 

2A 

⎛ 

r 

p1 − pa p1 − p ⎞ 

a 

Tv = ⎜ Hi + − H f + , 

2 ⎜ 

⎟ 

γ γ ⎟ 

⎡ ⎛ Ao 

⎞ ⎤ ⎝ ⎠ 

µ o Ao 2g ⎢1 − ⎜ ⎥ 

A 

⎟ 

⎢⎣ ⎝ ⎠ ⎥⎦ 

where µ o εφ = is the discharge coefficient, 1 1 φ = + ζo 

− the velocity 

coefficient, γ = ρg − the specific weight, ρ − the liquid density, p 1 − the 

pressure at the liquid surface in the reservoir, p a − the atmospheric pressure. 

The emptying time rate is defined by the emptying time of the variable head 

reservoir divided by the emptying time of the constant head reservoir 

k 

T 

c 

a 

; 

(2) 

(3) 

Tv 

= (4) 

T


In Fig. 2a are presented the comparative results for emptying times of 

constant and variable head reservoirs with respect to the percent of reservoir 

emptying ( 1 − kH 

[%], k H = H f Hi 

). 

In Fig. 2b the emptying time rate is presented with respect to the same 

parameter 1 − kH 

[%]. 

The numerical values of the characteristic parameters are: H i =1 m, the 

reservoir section 

r 

2 

r 4 

Ao 2 

πDo 

4 

A = πD with the reservoir diameter D r =1 m, the 

emptying orifice section = with the emptying orifice diameter D o = 

=.01 m, the discharge coefficient of the emptying orifice µ o = 0.7, the pressure 

inside the reservoir p 1 = p a . 

a b 

Fig. 2 – Reservoir with emptying orifice: numerical results: 

a – emptying time; b – emptying time rate. 

3. Functional Diagrams Method 

Based on the MATLAB/Simscape computational techniques, two 

functional diagrams for modelling and simulating the emptying process of a 

constant cross-sectional reservoir with constant and variable head has been 

developed. 

In Fig. 3a the functional diagram for constant head reservoir is presented. 

There are three subsystems: the level meter, Fig. 3b, the flow meter (Fig. 3c) 

and the calculation block which will calculate the velocity and the Reynolds 

number starting with the flow value (Fig. 3d). 

The condition for simulation stop is defined with the “Check Static 

Lower Bound” block which is set to stop the simulation when the liquid level 

will be equal with the imposed value f H (for this case H f =0). There are four 

“Display” blocks for the final values presentation of the four constant 

characteristic parameters: emptying time [s], velocity [m/s], Reynolds number 

and flow [m 3 /s].


a 

b c 

d 

Fig. 3 – Functional diagram for constant head reservoir: 

a – diagram, b – level meter subsystem, c – flow meter subsystem, 

d – calculation block subsystem. 

In Fig. 4 the functional diagram for variable head reservoir is presented. 

There are three subsystems, the same like in the previous case: the level meter 

(Fig. 3b) the flow meter (Fig. 3c) and the calculation block (Fig. 3d). There are 

four “Scope” blocks for the final values presentation of the four variable 


and flow [m 3 /s]. 

In Fig. 5 the numerical results are presented for the emptying process of a 

variable head reservoir (for this case H f =0). 

A comparative analysis of the numerical results for the emptying time for 

constant head and for variable head reservoirs, obtained with the analytical and 

the functional diagram methods is presented in Fig. 6a and Fig. 6b. The error


f 

c 

a 

c 

a 

c 

rates defined by ε = T −T 

T ⋅100 

[%] and ε = T − T T ⋅100 

[%] 

are presented in Fig. 6c and Fig. 6d. 

c 

Fig. 4 – Functional diagram for variable head reservoir. 

v 

f 

v 

a b 

c d 

Fig. 5 – Numerical results with respect to time [s]: 

a – level [m], b –flow rate [m 3 /s], c – velocity [m/s], d – Reynolds number. 

a 

v 

a 

v


a b 

c d 

Fig. 6 – Comparative analysis: 

a – emptying time, b – error rate - for constant head reservoir; 

c – emptying time, d – error rate - for variable head reservoir. 


1. The analytical method is based on the fundamental relationship for the 

emptying time for both reservoirs, with constant head and with variable head. 

The emptying time for variable head reservoir is greater than for constant head 

reservoir, and this difference is increasing with the percent of emptying. The 

emptying time rate is also increasing with the percent of emptying having the 

maximum value of 2 for full reservoir emptying. 

2. The functional diagrams method is based on the MATLAB/Simscape 

development methodology using Simscape functional blocks like: reservoir 

with constant head, reservoir with variable head, orifice with constant section, 

flow meter, etc. The Simulink blocks are used for data input and for data output 

respect to the functional diagram; input and output converters are required. 

3. For the functional diagram with constant head reservoir the output data 

are constant, so simple “Display” blocks have been used for single value data 

display. For the functional diagram with variable head reservoir the output data 

are variable, so “Scope” blocks have been used for graphical representation. For 

the variable head reservoir, the numerical values of flow, velocity and Reynolds


number are decreasing in time, but starting with the same values obtained for 

the constant head reservoir. 

4. For both functional diagrams with constant head and variable head 

reservoir a simulation stop condition has been used. This condition consists in a 

comparative test between the actual liquid level and the imposed final liquid 

level in the reservoir. When this condition will become true, the simulation will 

be stopped and the final simulation time will be displayed. 

5. A very good agreement between the numerical results obtained by both 

analytical and functional methods has been observed for emptying times, as 

well as for error rates: 

a 

T c and 

a 

v 

T are the emptying times determined using the 

f f 

analytical method for constant and variable head reservoirs; T c and T v are the 

emptying times determined using the functional diagram method. The 

maximum error rate is no greater than 0.02% for constant head reservoir and 

0.055% for variable head reservoir. 

REFERENCES 

Bartha I., Javgureanu V., Marcoie N., Hidraulică. Ed. Performatica, Iaşi, 2004. 

Idelcick I. E., Îndrumător pentru calculul rezistenŃelor hidraulice. Ed. Tehnică, 


Panaitescu V., Tcacenco V., Bazele mecanicii fluidelor. Ed. Tehnică, Bucureşti, 2001. 

Zahariea D., Simularea sistemelor fzice în MATLAB. Ed. PIM, Iaşi, 2010. 

*** Mathworks, Simscape Model and Simulate Multidomain Physical Systems. 

www.mathworks.com 

DIAGRAME FUNCłIONALE PENTRU MODELAREA PROCESULUI DE 

GOLIRE A UNUI REZERVOR PRINTR-UN ORIFICIU DE GOLIRE 

(Rezumat) 

În lucrare se analizează procesul de golire al unui rezervor cu secŃiune constantă 

printr-un orificiu de golire plasat la baza rezervorului. Sunt considerate două tipuri de 

rezervoare: cu sarcină constantă şi cu sarcină variabilă. Sunt prezentate două metode de 

analiză: metoda analitică şi metoda diagramelor funcŃionale. În cadrul metodei analitice 

se prezintă relaŃiile fundamentale pentru calculul timpului de golire, la determinarea 

cărora s-au considerat: presiunea la nivelul liber al lichidului din rezervor, viteza de 

coborâre a nivelului liber, coeficientul de debit al orificiului de golire. În cadrul metodei 

funcŃionale se prezintă două diagrame funcŃionale elaborate în MATLAB/Simscape la 

elaborarea cărora s-au utilizat elementele funcŃionale specifice pentru simularea 

sistemelor fizice reale. Diagramele funcŃionale permit obŃinerea reprezentărilor grafice 

pentru variaŃia cantităŃii de lichid din rezervor (şi prin calcul, a poziŃiei nivelului liber al 

lichidului), precum şi pentru debitul de lichid care trece prin orificiul de golire (şi prin 

calcul, a vitezei de curgere şi a parametrului Reynolds). Se prezintă o analiză 

comparativă a rezultatelor obŃinute prin cele două metode de analiză (metoda analitică şi 

metoda diagramelor funcŃionale).


Publicat de 



SecŃia 


FUNCTIONAL DIAGRAMS FOR MODELING THE RESERVOIR 

EMPTYING PROCESS THROUGH AN EMPTYING PIPE 

BY 

DĂNUł ZAHARIEA * 

“Gheorghe Asachi”Technical University of Iaşi 




Abstract. In this paper the emptying process of constant head and variable 

head reservoirs with constant cross-sectional area through an emptying pipe will 

be investigated. The pressure at the water surface and also the velocity of this 

surface, as well as the head loss due to friction along the emptying pipe will be 

considered. Two methods will be presented: the analytical method based on the 

fundamental equations governing the reservoir emptying process and the 

functional diagrams method based on MATLAB/Simscape computational 

approach suited to analyze the real physical systems. A comparative analysis of 

the numerical results for the emptying time for constant head and for variable 

head reservoirs, obtained with the analytical and the functional diagram methods 

will be presented. 

Key words: constant cross-sectional area reservoir, emptying time, emptying 

pipe, functional diagram, MATLAB. 


The supplying systems, as well as the emptying systems connected to a 

reservoir, are generally composed by network pipes. Let us consider the 

simplest case in which the emptying system of a supply reservoir is composed 

by a single pipe connected at the right bottom side of the reservoir. The 

reservoir emptying time problem can be formulated as follows: if all 

geometrical and functional characteristics of the reservoir and the emptying pipe 

* e-mail: dzahariea@yahoo.com


are known, what is the time for the liquid level to decrease from the initial value 

to a final imposed value? Three complementary parameters must be determined: 

the flow rate, the velocity and the Reynolds number for the liquid flow through 

the emptying pipe. 

Two resolution methods can be developed: the analytical method and the 

functional diagram method. The analytical method consists in determining the 

equation of the emptying time (Bartha, 2004), (Panaitescu & Tcacenko, 2001). 

Despite the apparent simplicity of this method, there are two important aspects 

to be mentioned: for the variable head reservoir case, the linear loss head 

coefficient λ is liquid level dependent, thus the velocity coefficient ϕ will be 

liquid level dependent too, (Idelcik, 1984). Considering this observation, the 

integrals involved in the emptying time equations for variable head reservoir 

can be solved only using numerical methods. The second important aspect 

refers to obtaining of the linear loss head coefficient. Three different 

relationships will be used for laminar, turbulent and transition fluid flow 

regimes. The functional method allows developing functional diagrams used for 

complete analyse of the reservoir emptying process (Zahariea, 2010), 

(Mathworks). All four parameters will be determined: the liquid level, the flow 

rate, the liquid velocity and the Reynolds number. 

In this paper, both the analytical and the functional diagram methods for 

analysing the emptying process of a reservoir through an emptying pipe will be 

presented. 

2. Analytical Method 

Let us consider a vertical reservoir with constant cross-sectional 

2 

area Ar = πDr 

4 = ct. , ∀ z ∈[ 

0, 

H ] where H is the reservoir maximum level, 

Fig. 1. The characteristics of the emptying pipe are: diameter D c , 

2 

c = c 4 , length c 

section A πD L , local head loss coefficients at the input ζ i and 

ζ , linear head loss coefficient λ . 

output e 

Fig. 1 – Reservoir with emptying pipe: characteristics.


The liquid volume variation produced during the emptying time dt by 

level change from z to z − dz 

will be equal with the liquid volume flowing 

through the emptying pipe in the same time dt. 

Using the conditions t = 0 ⇒ H = Hi 

and f H H T t = ⇒ = , where H i 

and H f are the initial and the final liquid levels, the emptying time T will be 

H 

i 

Ar 

1 

T = dz 

A ∫ , (1) 

v z 

c H f 

where v(z) is the liquid velocity through the emptying orifice which will be 

obtained starting from the Bernoulli’s equation for incompressible flow between 

characteristic sections 1 and 2. 

Different expressions can be obtained depending of the reservoir type: 

i) for constant head reservoir, the emptying time is 

T 

c 

( ) 

= 

φ( H 

Ar 

) A 2g 

Hi − H f 

p1 − p 

γ 

i c a Hi 

+ 

ii) for variable head reservoir, the emptying time is 

T 

v 

H 

i Ar dz 

= ∫ 

A 2g 

p − p 

f φ( z) z + 

γ 

c H 

1 a 

where γ = ρg is the specific weight, ρ − the liquid density, p 1 − the pressure 

at the liquid surface in the reservoir, p a − the atmospheric pressure, φ − the 

velocity coefficient. 

The velocity coefficient is defined by 

1 

φ( z) 

= 

, 

2 

⎛ Lc ⎞ ⎛ Ac 

⎞ 

1 + ⎜ λ( z) + ζi + ζ e ⎟ − ⎜ ⎟ 

⎝ Dc ⎠ ⎝ Ar 

⎠ 

where linear head loss coefficient is obtained using different expressions, 

depending on the Reynolds number: 

λ L = 64 ℜe, ℜe ≤ ℜ eL = 2000 , 

{ ( ) } 2 

1 

T 

1.8 lg 6.9 e e 

, 

1.11 

3.7 c 

e eT 4000 

D 

λ = 

− ⋅ ⎡ 

⎢ 

ℜ + ∆ 

⎣ 

⎤ 

⎥⎦ 

ℜ ≥ ℜ = , 

λT − λL 

λ = λL 

+ 

ℜe − ℜe 

( ℜe− ℜe L ) , ℜ eL < ℜ e < ℜe 

T . . 

T L 

, 

; 

(2) 

(3) 

(4) 

(5)


The emptying time rate is defined by the emptying time of the variable 

head reservoir divided by the emptying time of the constant head reservoir 

k 

T 

Tv 

= . 

(6) 

T 

In Fig. 2a are presented the comparative results for emptying times of 

constant and variable head reservoirs with respect to the percent of reservoir 

emptying ( − kH 

k = H H ). In Fig. 2b the emptying time rate is 

1 [%], H f i 

presented with respect to the same parameter 1 − kH 

[%]. 

The numerical values of the characteristic parameters are: H i =1 m, the 

2 

reservoir section Ar = πDr 

4 with the reservoir diameter D r =1 m, the 

2 

emptying pipe section Ac = πDc 

4 with the emptying pipe diameter D c =0.01 

m, the emptying pipe length L c = 2 m, the local head loss coefficients at the 

input ζ i = 0.5 and output ζ e = 1, the pressure inside the reservoir p 1 = p a . 

a b 

Fig. 2 – Reservoir with emptying pipe: numerical results. 

a – emptying time, b – emptying time rate. 

3. Functional Diagrams Method 

Based on the MATLAB/Simscape computational techniques two 

functional diagrams for modelling and simulating the emptying process of a 

constant cross-sectional reservoir with constant and variable head has been 

developed. 

In Fig. 3a the functional diagram for constant head reservoir is presented. 

There are three subsystems: the level meter (Fig. 3b) the flow meter (Fig. 3c) 

and the calculation block which will calculate the velocity and the Reynolds 

number starting with the flow rate value (Fig. 3d). 

c


The condition for simulation stop is defined with the “Check Static 

Lower Bound” block which is set to stop the simulation when the liquid level 

will be equal with the imposed value f H (for this case H f =0.0125 m). There 

are four “Display” blocks for the final values presentation of the four constant 


and flow [m 3 /s]. 

a 

b c 

d 

Fig. 3 – Functional diagram for constant head reservoir: 

a – diagram, b – level meter subsystem, c – flow meter subsystem, 

d – calculation block subsystem. 

In Fig. 4 the functional diagram for variable head reservoir is presented. 

There are three subsystems, the same like in the previous case: the level meter, 

(Fig. 3b) the flow meter (Fig. 3c) and the calculation block (Fig. 3d). There are 

four “Scope” blocks for presentation of variable characteristic parameters: 

emptying time [s], velocity [m/s], Reynolds number and flow [m 3 /s]. 

In Fig. 5 the numerical results are presented for the emptying process of a 

variable head reservoir (for this case Hf =0.0125 m).


A comparative analysis of the numerical results for the emptying time for 

constant head and for variable head reservoirs, obtained with the analytical and 

the functional diagram methods is presented in Fig. 6a and Fig. 6b. The error 

f a a 

f a a 

rates defined by ε = T − T T 100 [%] and ε = T − T T 100 [%] are 

c c c c 

presented in Fig. 6c and Fig. 6d. 

v v v v 

Fig. 4 – Functional diagram for variable head reservoir. 

a b 

c d 

Fig. 5 – Numerical results with respect to time [s]: 

a – level [m], b –flow rate [m 3 /s], 

c – velocity [m/s], d – Reynolds number.


a b 

c d 

Fig. 6 – Comparative analysis: 

a – emptying time, b – error rate - for constant head reservoir; 

c – emptying time, d – error rate - for variable head reservoir. 


1. The analytical method is based on the fundamental relationship for the 

emptying time. The linear loss head coefficient will be calculated using three 

different relationships depending of the fluid flow regimes. An iterative 

computation block have been developed having the validation condition based 

on the Reynolds number criterion. The emptying time for variable head 

reservoir is greater than for constant head reservoir, and this difference is 

increasing with the percent of emptying. The emptying time rate is also 

increasing with the percent of emptying. 

2. The functional diagrams method is based on the MATLAB/Simscape 

development methodology using Simscape functional blocks for simulate real 

physical elements, like: reservoir with constant head, reservoir with variable 

head, pipe, flow meter, etc. The Simulink blocks are used for data input and for 

data output respect to the functional diagram, when input and output converters 

are required. 

3. For both functional diagrams with constant head and variable head 

reservoir a simulation stop condition has been used. This condition consists in a


comparative test between the actual liquid level and the imposed final liquid 

level in the reservoir. When this condition will become true, the simulation will 

be stopped and the final simulation time will be displayed (this time correspond 

with the emptying time). 

4. A good agreement between the numerical results obtained by both 

analytical and functional methods has been observed for emptying times, as 

well as for error rates: 

a 

T c and 

a 

v 

T are the emptying times from analytical 

f f 

method for constant and variable head reservoirs; T c and T v are the emptying 

times from the functional diagram method for constant and variable head 

reservoirs. The maximum error rate is no greater than: 3.35% for constant head 

reservoir and 5.5% for variable head reservoir. 

REFERENCES 

*** Mathworks, Simscape Model and Simulate Multidomain Physical Systems. 

www.mathworks.com 

Bartha I., Javgureanu V., Marcoie N., Hidraulică. Ed. Performatica, Iaşi, 2004 

Idelcick I. E., Îndrumător pentru calculul rezistenŃelor hidraulice. Ed. Tehnică, 


Panaitescu V., Tcacenco V., Bazele mecanicii fluidelor. Ed. Tehnică, Bucureşti, 2001. 

Zahariea D., Simularea sistemelor fzice în MATLAB. Ed. PIM, Iaşi, 2010. 

DIAGRAME FUNCłIONALE PENTRU MODELAREA PROCESULUI DE GOLIRE 

A UNUI REZERVOR PRINTR-O CONDUCTĂ DE GOLIRE 

(Rezumat) 

În lucrare se analizează procesul de golire al unui rezervor cu secŃiune constantă 

printr-o conductă de golire. Sunt considerate două tipuri de rezervoare: cu sarcină 

constantă şi cu sarcină variabilă. Sunt prezentate două metode de analiză: metoda 

analitică şi metoda diagramelor funcŃionale. În cadrul metodei analitice se prezintă 

relaŃiile fundamentale pentru calculul timpului de golire, la determinarea cărora s-au 

considerat: presiunea la nivelul liber al lichidului din rezervor, viteza de coborâre a 

nivelului liber, pierderile locale şi liniare de sarcină de pe condcucta de golire. În cadrul 

metodei funcŃionale se prezintă două diagrame funcŃionale elaborate în 

MATLAB/Simscape la elaborarea cărora s-au utilizat elementele funcŃionale specifice 

pentru simularea sistemelor fizice reale: rezervorul cu sarcină constantă sau variabilă, 

conducta de golire, traductorul de debit, etc. Diagramele funcŃionale permit obŃinerea 

reprezentărilor grafice pentru variaŃia cantităŃii de lichid din rezervor (şi prin calcul, a 

poziŃiei nivelului liber al lichidului), precum şi pentru debitul de lichid care trece prin 

orificiul de golire (şi prin calcul, a vitezei de curgere şi a parametrului Reynolds). 

Pentru cele două tipuri de rezervoare (cu sarcină constantă şi cu sarcină variabilă) se 

prezintă o analiză comparativă a rezultatelor obŃinute prin cele două metode de analiză 

(metoda analitică şi metoda diagramelor funcŃionale).


Publicat de 



SecŃia 


A PLEADING FOR THE MODERNIZATION OF 

MACHINE TOOLS IN NATIONAL INDUSTRY 


Accepted for publication: Sptember 2, 2011 

BY 

DAN POPESCU * 

University of Craiova, 

Department of Electrotechnics 

Abstract. Numerical controls have a great role in increasing the efficiency 

of the electric drive of the machine tool, thus in the energy saving. Measurements 

were done for two of those (one equiped with numerical control and a converter 

for the asynchronous motor of the main drive, and another with no numerical 

control and with a converter for the d.c. motors of the main drive). Measurements 

were done using a kit which was specialized in aquiring and handling energetic 

data. This project presents the diagrams resulted from the measurements in a 

comparative study. The monitored parameters were the following: actual values 

of voltage and currents during the three phases, the content of harmonics, or the 

frequency spectrum for all three phases, the frequency of the feeding voltage, 

active and reactive power, energy, and the power factor. 

Key words: numerical controls, active and reactive power, energy, converter. 


Lately, the technical needs in the area of mechanic processing of materials 

have evolved. The increased complexity of the machine and of the workpiece 

calls for more qualified and more thorough manpower. These two attributes are 

usually dominated by human subjectivism: tiredness, neglection, inattention, 

lack of promptitude in critical situations - which make humans imperfect 

operators. In time, mankind was prone to discover methods of avoiding 

unpleasant situations during the productive process. One of these methods was 

automatization, with its top element, numerical control (Minciu & Predincea, 

1985). Numerical controls are mainly necessary for increasing work 

* e-mail: popescumincu@yahoo.com

228 Dan Popescu 

productivity while maintaining high precision. It is because of this need that 

they have continuousy evolved, until human error was eliminated from the 

execution of workpieces. As shown in their name,computer assisted numerical 

controls(CNC) have as main element a computer, similar to a pc, not very 

powerful, but adapted to industrial needs (increased protection for adapting to 

the environment and an operating safety suited to demands) . 

2. Functioning of CNC Equipped Machine Tools 

A CNC system also consists of a peripherics connection unit, an external 

memory, an interface between the computer and the machine and a control 

panel. All these equipments (system hardware) function according to the basic 

software. The software is based on Windows NT, which is compatible with the 

numerical control software, for starting and operating, of the original 

manufacturing equipment. For instance, let’s consider a Sinumerik 840 

numerical control made by Siemens, with which a lathe is equipped. The lathe is 

placed in a light mechanical processing department and it provides the serial 

production of the pieces which are to be produced. The lathe is also equipped 

with a tool storage space, thus ensuring low time intervals for the execution of 

technological operations. The processing programs are stocked on a memory 

card. The command system of the machine consists of: 

i) the control panel, which includes the operator panel, the display, the 

control panel of the machine, keyboard and mouse; 

ii) mains infeed module (MS), which provides energy for the drive 

modules, the regenerativ feedback, braking or braking resistance modules; 

iii) the numerical control unit (NCU); 

iv) the supplies power for the axis modules- it generates continuous 

tension for the convertors of the motors; 

v) blocks of the feed drive digital system, one for each axis, depending of 

the configuration of the machine tool. They hold tension and frequency PWM 

convertors, ensuring high energetic performances. 

The input/output modules and the feed drive modules are connected with 

a Profibus 12 MBaud base bus (Mărgineanu, 2005). 

Numerical controls also have a great role in increasing the efficiency of 

the electric drive of the machine tool, thus in the energy saving. First of all, they 

assess a machine which is extremely well built. This machine must be precise, 

in order to ensure the meeting of processing demands, it must not have friction 

or displacement during steering. 

The direct consequence is a smaller power needed for the drive. Second 

of all, numerical controls allow choosing and respecting an optimal working 

condition for the advance and for the splinting, also forcing the static frequency 

and tension changers, which also increase their efficiency, to have an economic 

energetic state. These were also the conclusions of the measurements that were 

done, because the equipment allowed comparison between the same type of 

machines, some of them with numerical controls and some without.

Bul. Inst. Polit. Iaşi, t. LVIII (LXII), Fasc. 1, 2012 229 

3. Measuring Process- Observations 

Measurements were done for the boring machine AFP 160 CNC 

(equiped with numerical control and Fanuc converter for the asynchronous 

motor of the main drive) and for the the boring machine AFP160 (with no 

numerical control and with a converter for the continuos power motor of the 

main drive). Measurements were done using a kit (AR-5 network analyser) 

which was specialized in aquiring and handling energetic data. We must 

mention that the two machines executed the same operation (finishing). 

The boring dust resulted from the processing was compared and 

dimensions or the chips were found to be similar. This project presents the 

diagrams resulted from the measurements in a comparative study. The 

monitored parameters were the following: actual values of tensions. In the case 

of active powers, in the most charged phase, the energy consumption is three 

times higher for non-upgraded machine, and this ratio is also mantained in the 

case of reactive powers, disfavouring the non-upgraded machine, currents 

during the three phases, the content of the harmonics or the frequency spectrum 

for all three phases, the frequency of the feeding tension, active and reactive 

power and energy, and the power factor. 

Frequency spectrums of the harmonics content of the voltage waves and 

the variation of harmonics while monitoring look like in Figs. 1 and 2. We can 

see that especially harmonics three and five are almost two times smaller for 

machine no 1. This means in the case of non-upgraded machine, the superior 

harmonics of the flux and rotoric current (of different orders and sequences) 

determin disturbing couples to appear (Ciobanu, 2008). The amplitude of those 

couples is independent of the charge, overlapping asincronous couple produced 

by the fundamental. If the feeding frecquency is high, their effect is unnoticed. 

At small rotation speeds of the asyncronous motors, the swinging couples can 

produce an abrupt(or in several steps) movement of the engine rotor. 

Fig.1 – Armonics for machine AFP160 CNC. Fig.2 – Armonics for machine AFP160. 

Their presence actually limits the minimum speed at which the motor can 

be used. There are different ways to reduce the effects of these couples, either 

operating on the motor, or on the constructive plan of the convertors (Bizon,


2008), (Câmpeanu, 2008). At low speeds, although the harmonics content 

increases, the effective value of the current decreases with the decreasing of the 

speed. All these couples, dued to the superior harmonics, tend to decrease the 

maximum torque of the engine, in comparison to network feeding. 

Fig. 3 – Active power for AFP 160CNC. Fig. 4 – Active power for AFP160. 

Fig. 5 – Reactive power for AFP160CNC. Fig. 6 – Reactive power for AFP160. 

This means that a reserve is necessary (variation of 10-15% of the 

engine power), in order to handle the effects of the harmonics. In the case of 

active powers, in the most charged phase, the power consumption is two times 

higher for for non-upgraded machine (Figs. 3 and 4) and this ratio is also 

mantained in the case of reactive powers, disfavouring the non-upgraded 

machine (Figs. 5 and 6).This thing leads to obvious drastic decrease of the 

power factor which can cause additional costs in order to compensate it.


Variation graphs of the actual values of currents look like in Figs. 7 and 

8. During the most charged phase we can notice a significant growth of the 

absorbed current, especially when the machine drive breaks and accelerates. 

Fig.7 – Current consumption for AFP160CNC. Fig.8 – Current consumption or FP160. 

4. Economic Considerations 

These benefits are noticeable if we do a basic economic calculation- 

determination of the recovery period and the project profitability (Leca, 1997). 

Table 1 

Calculus of static indicators 

Nr. Indicators M.U. Results 

1 Economy by introducing CSF kW/hour 1,4 

2 Functioning hours hours/year 3840 

3 Annual energy economy MWh 5,36 

4 The price/ energy lei /MWh 311.98 

5 The cost of the energy save/ year lei/year 1677 

6 The cost of the installation lei 560 000 

From the absorbed active power diagrams we notice a difference of 

about 1,4kW in favour of upgraded machine. Considering that the machinery 

works in only two shifts, that the owner has proposed a moderate yearly profit, 

about 100 000 lei after he spent 560 000 lei on upgrading the machinery, by 

carrying the appropriate calculations we obtain the static indicators grouped in 

Table 1. Calculus of static indicators (Leca, 1997) shows an acceptable recovery 

period, less than five years, which confirms once more the profitability resulted 

from the machinery upgrades. 

It should be noted that there are no additional maintenance costs 

because on one side the machinery is under warranty for a long period of time, 

and on the other side modern equipments are higly reliable. It may also be noted 

that the owner may not bear all modernization costs.


The costs can decrease spectacularly by accesing European funds 

intended for energy saving. 


1. We can conclude that as numerical controls evolve, they become more 

and more accesible to operators and to projecting engineers. Difficulties remain 

for software developers, who must find more and more accesible programmes, 

thus lowering the costs by reducing the work time, reducing the energetic 

consumption (by optimization of feed controls) and reducing expenses with 

personnel. 

2. The boards of factories in the processing branch can easily decide the 

modernization of the existing machine tools, considering both the high 

productivity and the small amount of time in which the investment will be 

recovered. According to calculations, in all the cases we studied the investment 

is easily recuperated, thus attractive. The attractivity increases as there are 

numerous European programmes in the energy field, according to which the 

investor only has to cover part of the investment. All these also contribute to 

general economic development, by creating and developing import, assembling 

and service firms, and also creating firms with connected activities. 

3. The considerable progress in the field of electrical systems with 

variable speed made the investment necessary for such systems to be just a 

fraction of the total cost of the energy used during their functioning. 

4. All these also contribute to general economic development, by 

creating and developing import, assembling and service firms, and also creating 

firms with connected activities. From all the above, we can conclude the 

following advantages of using static converters: precise control of the energetic 

process and especially of the system’s speed; very low maintenance; energy 

save; smooth start/stop, without blows of controllable acceleration and 

deceleration; almost inexistent noise. 

Command pulse time modulation inverters, which eliminate the superior 

harmonics. (Delesega&Andea, 2002), (Mircea, 1999) have represented an 

important step in the establishment of the driving with an asynchronous motor 

as a solution for the future. This is why, while the static tension and frequency 

converters are being modernized, the balance is towards the driving with an 

asynchronous motor against the D.C. one. 

REFERENCES 

*** Convertor valutar. available at: http://www.dobanzi.ro accessed: 2011-06-10. 

*** Siemens SINUMERIK 840Di Documentation, 2001. 

Bizon N., Convertoare. Ed. Matrix Rom, Bucureşti, 2008. 

Câmpeanu A., Maşini electrice. Ed. Universitaria, Craiova, 2008. 

Ciobanu L., Sisteme de acŃionări electrice. Ed. Matrix Rom, Bucureşti, 2008.


Delesega I., Andea P., Procese de comutaŃie. Calitatea energiei electrice. Ed. 

Orizonturi Universitare, Timişoara, 2002. 

Leca A., Principii de management energetic. Ed.Tehnică. Bucureşti, 1997. 

Mărgineanu I. , Automate programabile. Ed. Albastră, Cluj Napoca, 2005. 

Minciu C., Predincea N., Maşini unelte cu comandă numerică. Ed. Tehnică, Bucureşti, 

1985. 

Mircea I., Sisteme eficiente energetic pentru instalaŃii cu debite reglabile. Ed. 

Universitaria, 1999. 

O PLEDOARIE PENTRU MODERNIZAREA MAŞINILOR UNELTE 

ACTUALE DIN INDUSTRIA NAłIONALĂ 

(Rezumat) 

O contribuŃie deosebită o au comenzile numerice în eficientizarea acŃionărilor 

electrice a utilajelor, şi deci în economia de energie pe care o aduc. În primul rând, ele 

impun un utilaj bine construit din punct de vedere mecanic. Acesta trebuie să fie precis, 

pentru asigurarea exigenŃelor de prelucrare, să nu aibă frecări şi jocuri mari în ghidaje. 

ConsecinŃa directă este o putere necesară pentru acŃionare mai mică. În al doilea rând, 

comenzile numerice permit alegerea şi respectarea unui regim optim de avans şi 

aşchiere, impunând şi convertizoarelor statice de tensiune şi frecvenŃă, care la rândul lor 

au un randament din ce în ce mai ridicat, un regim economic din punct de vedere 

energetic. Acest lucru a reieşit şi din măsurătorile făcute, unde graŃie dotării 

întreprinderii s-au putut face comparaŃii pe acelaşi tip de maşini, unele fiind 

convenŃionale, iar altele fiind proaspăt echipate cu comenzi numerice. 

Pentru confirmarea celor de mai sus în ceea ce priveşte economia de energie 

pentru utilajele modernizate s-au făcut comparativ mai multe măsurători pe utilaje de 

acelaşi tip, încărcate cu aceleaşi repere în vederea prelucrării. S-a urmărit ca regimurile 

de aşchiere să fie aceleaşi pentru acurateŃea comparaŃiilor ulterioare. 

S-au făcut măsurări pe maşina de alezat şi frezat cu pinolă AFP180 CNC, 

echipată cu comandă numerică şi convertizor Fanuc pentru motorul asincron de 

acŃionare principală, şi pentru maşina de alezat şi frezat cu pinolă AFP180, fără 

comandă numerică, echipat cu convertizor pentru motorul de curent continuu de 

acŃionare principală. 

Parametrii monitorizaŃi au fost: valorile efective ale tensiunilor şi curenŃilor pe 

cele trei faze, conŃinutul în armonici sau spectrul de frecvenŃă al tensiunilor pentru toate 

fazele, frecvenŃa tensiunii de alimentare, puterea şi energia activă şi reactivă, precum şi 

factorul de putere. Se observă în cazul puterilor active că pe faza cea mai încărcată 

consumul este de circa de două ori mai mare la utilajul nemodernizat, iar în cazul 

puterilor reactive proporŃia se respectă în defavoarea aceluiaşi utilaj. Acest lucru 

conduce evident şi la scăderea drastică a factorului de putere ceea ce determină costuri 

suplimentare pentru compensarea lui. 

Pe faza cea mai încărcată se observă o creştere semnificativă a curentului 

absorbit mai ales în momentele de frânare şi accelerare ale platoului utilajului. 

Din calculele economice făcute reiese că în toate cazurile studiate investiŃiile se 

recuperează rapid, ceea ce face să fie atractivă investiŃia. Atractivitatea constă şi în 

existenŃa mai multor programe europene în domeniul energetic în care investitorul vine 

numai cu un anumit procent din suma investită. Toate acestea contribuie şi la


dezvoltarea economică generală, prin apariŃia şi activitatea firmelor de import, montaj şi 

service în domeniu, dar şi la apariŃia altor firme din domenii pe orizontală ale activităŃii. 

Se poate trage concluzia din cele prezentate că pe măsura evoluŃiei lor, 

comenzile numerice devin din ce în ce mai accesibile atât operatorilor, cât şi inginerilor 

din proiectarea uzinală. DificultăŃile se deplasează în zona dezvoltatorilor software care 

trebuie să găsească programe din ce în ce mai accesibile, reducând astfel costurile prin 

reducerea timpului de lucru, reducerea consumurilor energetice prin optimizarea 

comenzilor acŃionărilor şi reducerea cheltuielilor de calificare cu personalul.

buletinul institutului politehnic din iaşi - Universitatea Tehnică ...

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