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BULETINUL 

INSTITUTULUI 

POLITEHNIC 

DIN IAŞI 

Tomul LVIII (LXII) 

Fasc. 4 

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

BULETINUL INSTITUTULUI POLITEHNIC DIN IAŞ I 

BULLETIN OF THE POLYTECHNIC INSTITUTE OF IAŞ I 

Tomul LVIII (LXII), Fasc. 4 2012 

CONSTRUCŢII DE MAŞINI 

S U M A R 

ANTONINO LA ROCCA (Aglia), VINCENZO LA ROCCA (Italia) şi 

MASSIMO MORALE (Italia), Recuperarea energiei generate în procesul 

de vaporizare a gazelor naturale lichefiate (engl., rez. rom.) ..................... 

DANIEL DRAGOMIR-STANCIU şi MARIAN STANCU, Proiectarea şi 

testarea unui gazeificator cu circulaţie descendentă, cu puterea 50 kW 

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

Pag. 

GRAŢIELA MARIA ŢÂRLEA, ION ZABET, MIOARA VINCERIUC şi 

ANA ŢÂRLEA, Studiu teoretic comparativ privind îmbunătăţirea din 

punct de vedere al eco-eficienţei unui sistem frigorific ca urmare a 

înlocuirii hidrocarburilor şi amestecurilor de HFC cu amoniac (engl., 

rez. rom.)..................................................................................................... 35 

NICOLAE BARA şi MARIN BICA, Influenţa factorilor de mediu şi 

constructivi asupra performanţelor vaporizatoarelor (engl., rez. rom.)....... 43 

IONEL OPREA, Metoda multicriterială de alegere a agenţilor frigorifici 

pentru pompele termice (engl., rez. rom.)................................................... 51 

RĂZVAN FLORIN BARZIC, IULIANA STOICA, ANDREEA IRINA 

BARZIC şi GHEORGHE DUMITRAŞCU, Proprietăţi termice ale 

nanocompozitelor polistiren/nanotuburi de carbon obţinute prin metoda 

forfecării (engl., rez. rom.).......................................................................... 59 

IONEL IVANCU, DANIEL DRAGOMIR STANCIU, IONUŢ CRÎŞMARU, 

DAN TEODOR BĂLĂNESCU şi GEORGE OVIDIU RĂU, Modelarea 

unui schimbător de căldură cu plăci în curent încrucişat (engl., rez. rom.) 65 

ION ZABET şi GRAŢIELA-MARIA ŢÂRLEA, Rezultate experimentale 

privind un compressor scroll cu injecţie de vapori (engl., rez. rom.) ....... 71 

VICTOR PANTILE, CONSTANTIN PANĂ şi NICULAE NEGURESCU, 

Studiu asupra performanţelor motorului cu aprindere prin scânteie 

alimentat cu hidrogen (engl., rez. rom.)............................................. 81 

ALEXANDRU RADU, CONSTANTIN PANĂ şi NICULAE NEGURESCU, 

Aspectre privind utilizarea bioetanolului în motoare cu aprindere prin 

scânteie (engl., rez. rom.)............................................................................ 91 

1 

29

IOAN HITICAS, LIVIU MIHON, DANILA IORGA şi WALTER SVOBODA, 

Nivelul emisiilor poluante ale motoarelor cu ardere internă care 

utilizează combustibili clasici şi alternativi (engl., rez. rom.) ................... 101 

COSTIN DRAGOMIR, CONSTANTIN PANĂ, NICULAE NEGURESCU şi 

ALEXANDRU CERNAT, Investigaţii teoretice experimentale ale 

motorului cu aprindere prin scânteie supraalimentat (engl., rez. 

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

EUGEN RUSU, CONSTANTIN PANĂ şi NICULAE NEGURESCU, Investigarea 

efectului adăugării hidrogenului în adaos la un motor alimentat cu 

benzină (engl., rez. rom.)............................................................................. 117 

ADRIAN SABĂU, CONSTANTIN DUMITRACHE şi MIHAELA BARHA- 

LESCU, Analiza funcţionării motorului diesel cu combustibilii metan, 

metanol şi diesel (engl., rez. rom.) ........................................................... 

ANCA ELENA ELIZA STERPU şi ANCA IULIANA DUMITRU, Reguli de 

amestecare pentru predicţia proprietăţilor de ardere ale combustibililor 

pentru motoarele diesel ale autovehiculelor (engl., rez. rom.) .................. 

MARIUS RECEANU, DAN DĂSCĂLESCU şi ADRIAN SACHELARIE, 

Optimizarea controlului debitului de aer la un sistem de răcire al unui 

motor termic (engl., rez. rom.) ................................................................... 153 

FlORIN POPA, EDWARD RAKOSI şi GHEORGHE MANOLACHE, 

Modelarea unui sistem de propulsie auto cu funcţionare stabilă (engl., 

rez. rom.)..................................................................................................... 161 

MARIANA LUPCHIAN, Determinarea componentelor sistemului de propulsie 

la bordul unui petrolier (engl., rez. rom.).................................................. 171 

VASILE GABRIEL NENERICA, VASILE HUIAN şi DORU CĂLĂRAŞU, 

Inroducere în sistemul de injecţie Common Rail (engl., rez. 

rom.)............................................................................................................ 177 

COSTICĂ ATANASIU, BOGDAN LEIŢOIU şi ŞTEFAN SOROHAN, 

Ruperea prin forfecare a unui material cromoplastic (engl., rez. 

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

OVIDIU NIŢĂ, VASILE BRAHA şi ANDREI MIHALACHE, Determinarea 

variaţiei rugozităţii tablelor subţiri în funcţie de deformaţia obţinută la 

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

OVIDIU NIŢĂ şi VASILE BRAHA, Determinarea curbelor limită de 

deformare utilizând metoda umflării hidraulice (engl., rez. rom.) ............. 205 

MARIAN TEODOR POPESCU, NICOLAE POPA, CONSTANTIN 

ONESCU şi RADU NICOLAE DOBRESCU, Consideraţii privind 

comportarea oţelului XC45 la frecare uscată la temperatura de 80 o C 

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

PETRONELA PARASCHIV şi CIPRIAN PARASCHIV, Rezultate numerice 

obţinute la modelarea dinamică a articulaţiei cotului (engl., rez. rom.) .... 221 

109 

127 

141 

187 

197 

213

VICTOR COTOROS şi EMIL BUDESCU, Aspecte privind biomecanica 

articulară: articulaţia gleznei (engl., rez. rom.) .......................................... 

VICTOR COTOROS şi EUGEN MERTICARU, Fractura maleolei tibiale: 

model virtual şi analiza cu metoda elementelor finite (engl., rez. rom.) .... 

IOAN ŢENU, RADU ROŞCA, PETRU CÂRLESCU şi VIRGIL VLAHIDIS, 

Stand de laborator pentru studiul impactului traficului agricol şi a 

lucrărilor tehnologice asupra proprietăţilor fizice ale solului (engl., rez. 

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

EMILIAN MOŞNEGUŢU, VALENTIN NEDEFF, OVIDIU BONTAŞ, 

NARCIS BÂRSAN şi DANA CHIŢIMUŞ, Studiul obţinerii traiectoriei 

unei particule solide pe o suprafaţă plană oscilantă (engl., rez. rom.)........ 

GELU NUŢU şi IOAN BĂISAN, Stand de laborator pentru studiul tăierii 

tulpinilor vegetale cu aparate de tăiere de tip cuţit-deget 

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

GELU NUŢU şi IOAN BĂISAN, Cercetări privind tăierea tulpinilor vegetale 

cu aparate de tăiere de la combinele de recoltat cereale (engl., rez. 

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

MIRELA PANAINTE, VALENTIN NEDEFF, CIPRIAN OLARU, 

CLAUDIA TOMOZEI şi OANA IRIMIA, Studiul comportării produselor 

alimentare supuse mărunţirii prin tăiere prin metoda analizei de 

textură (engl., rez. rom.).............................................................................. 283 

229 

243 

255 

263 

271 

277

BULETINUL INSTITUTULUI POLITEHNIC DIN IAŞ I 

BULLETIN OF THE POLYTECHNIC INSTITUTE OF IAŞ I 

Tomul LVIII (LXII), Fasc. 4 2012 

MACHINE CONSTRUCTION 

ANTONINO LA ROCCA (UK), VINCENZO LA ROCCA (Italy) and 

MASSIMO MORALE (Italy), Energy Recovery from Regasification of 

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

DANIEL DRAGOMIR-STANCIU and MARIAN STANCU, Design and 

Testing of a 50 kW Downdraft Gasifier (English, Romanian 

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

GRAŢIELA MARIA ŢÂRLEA, ION ZABET, MIOARA VINCERIUC and 

ANA ŢÂRLEA, Theoretical Eco-Efficiency Comparative Study Case, 

Hydrocarbons, Ammonia and HFC Mixture Alternative Retrofit 

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

NICOLAE BARA and MARIN BICA, Influence of Environmental and 

Constructive Factors on the Vaporizers Performance (English, Romanian 

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

IONEL OPREA, An Improved Method for Heat Pumps Refrigerant Choice 

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

RĂZVAN FLORIN BARZIC, IULIANA STOICA, ANDREEA IRINA 

BARZIC and GHEORGHE DUMITRAŞCU, Thermal Properties of 

Polystyrene/Carbon Nanotubes Composites Prepared by Shear Casting 

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

IONEL IVANCU, DANIEL DRAGOMIR STANCIU, IONUŢ CRÎŞMARU, 

DAN TEODOR BĂLĂNESCU and GEORGE OVIDIU RĂU, 

Modelling of a Plate Cross Flow Heat Exchanger (English, Romanian 

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

ION ZABET and GRAŢIELA-MARIA ŢÂRLEA, Experimental Results on a 

Vapour Injection Scroll Compressor (Engliah, Romanian summary)........ 

CORNELIU CRISTESCU PETRIN DRUMEA, CĂTĂLIN DUMITRESCU și DRAGOȘ ION GUȚĂ, Experimental Research Regarding the Dynamic Behavior of Linear Hydraulic Servo-Systems (English, Romanian Summary) ................................................................................ 

C O N T E N T S 

VICTOR PANTILE, CONSTANTIN PANĂ and NICULAE NEGURESCU, 

A Study on Performance of a Hydrogen Fuelled Spark Ignition Engine 

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

Pp. 

1 

29 

35 

43 

51 

59 

65 

71 

81

ALEXANDRU RADU, CONSTANTIN PANĂ and NICULAE NEGURESCU, 

Aspects Regarding the Use of Bioethanol in Spark Ignition Engines 

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

IOAN HITICAS, LIVIU MIHON, DANILA IORGA and WALTER 

SVOBODA, Emission Level from Internal Combustion Engine Using 

Fosil and Alternative Fuels (English, Romanian summary) ...................... 101 

COSTIN DRAGOMIR, CONSTANTIN PANĂ, NICULAE NEGURESCU 

and ALEXANDRU CERNAT, Theoretical and Experimental Investigations 

of the SI Engine Turbocharging (English, Romanian summary) ... 109 

EUGEN RUSU, CONSTANTIN PANĂ and NICULAE NEGURESCU, 

Investigation the Effect of Hydrogen Addition in a Spark Ignition Engine 

(English, Romanian summary).................................................................... 117 

ADRIAN SABĂU, CONSTANTIN DUMITRACHE and MIHAELA BAR- 

HALESCU, Analysis of Diesel Engine Operation on Methane, Methanol 

and Diesel Fuels (English, Romanian summary) ....................................... 127 

ANCA ELENA ELIZA STERPU and ANCA IULIANA DUMITRU, Mixing 

Rules for Predicting the Ignition Properties of Automotive Diesel Engine 

Fuels (English, Romanian summary).......................................................... 141 

MARIUS RECEANU, DAN DĂSCĂLESCU and ADRIAN SACHELARIE, 

Optimization of the Air Flow Control for a Car Engine Cooling System 

(English, Romanian summary) ................................................................... 153 

FlORIN POPA, EDWARD RAKOSI and GHEORGHE MANOLACHE, 

Modeling a Car Powertrain System with Stable Operation (English, 

Romanian summary)................................................................................... 161 

MARIANA LUPCHIAN, Determination of Propulsion System Components on 

Board a Tanker Ship (English, Romanian summary).................................. 171 

VASILE GABRIEL NENERICA, VASILE HUIAN and DORU CĂLĂRAŞU, 

Introduction to Common Rail Injection System (English, Romanian 

summary)..................................................................................................... 177 

COSTICĂ ATANASIU, BOGDAN LEIŢOIU and ŞTEFAN SOROHAN, 

Study of a Chromoplastic Material Shear Breaking (English, Romanian 

summary)..................................................................................................... 187 

OVIDIU NIŢĂ, VASILE BRAHA and ANDREI MIHALACHE, Determination 

of Sheets Metal Roughness Variation Depending on Tensile Strain 

(English, Romanian summary).................................................................... 197 

OVIDIU NIŢĂ and VASILE BRAHA, Determination of Forming Limit 

Diagrams Using Hydraulic Bulging Test (English, Romanian 

summary)..................................................................................................... 205 

91

MARIAN TEODOR POPESCU, NICOLAE POPA, CONSTANTIN 

ONESCU and RADU NICOLAE DOBRESCU, Considerations 

Regarding the Behaviour of Steel XC45 at Dry Friction at a Temperature 

of 80 o C (English, Romanian summary)..................................................... 213 

PETRONELA PARASCHIV and CIPRIAN PARASCHIV, Numerical Results 

Achieved through the Dynamic Shaping of the Elbow Joint (English, 

Romanian summary.) ................................................................................. 221 

VICTOR COTOROS and EMIL BUDESCU, Dynamic Analysis of Ankle 

Joint – 3D Biomechanics Model (English, Romanian summary).............. 229 

VICTOR COTOROS and EUGEN MERTICARU, Tibial Malleolus Fracture: 

Virtual Model and Analysis with the Finite Element Method (English, 


IOAN ŢENU, RADU ROŞCA, PETRU CÂRLESCU and VIRGIL VLAHIDIS, 

Laboratory Stand for Traffic Impact Study of Agriculture and 

Technology Works on Physical Properties of Soil (English, Romanian 

summary)..................................................................................................... 255 

EMILIAN MOŞNEGUŢU, VALENTIN NEDEFF, OVIDIU BONTAŞ, 

NARCIS BÂRSAN and DANA CHIŢIMUŞ, The Study to Obtaining the 

Trajectories of Solid Particles on an Oscillatory Flat Surface (English, 

Romanian summary) .................................................................................. 263 

GELU NUŢU and IOAN BĂISAN, Laboratory Stand for the Study of Cut 

Strains Vegetable with Knife-Fingers Type Cutting Apparatus (English, 


GELU NUŢU and IOAN BĂISAN, Researches Concerning the Cutting of 

Strains Plant with Cutting Devices from Combine Harvesters (English, 


MIRELA PANAINTE, VALENTIN NEDEFF, CIPRIAN OLARU, 

CLAUDIA TOMOZEI and OANA IRIMIA, Study the Behaviur of Food 

Materials with Soft Textured Subject to Shred by Cutting through the 

Texture Analysis Method (English, Romanian summary) ......................... 283


Publicat de 

Universitatea Tehnică „Gheorghe Asachi“ din Iaşi, 

Tomul LVIII (LXII), Fasc. 4, 2012 

Secţia 


ENERGY RECOVERY FROM REGASIFICATION OF LNG 

BY 

ANTONINO LA ROCCA 1 , VINCENZO LA ROCCA ∗2 

and MASSIMO MORALE 2 

1 University of Nottingham, UK, 

Department of Mechanical, Materials and Manufacturing Engineering 

2 Università di Palermo, Italy, 

Dipartimento dell’Energia 

Received: April 2012 

Accepted for publication: June 2012 

Abstract. The international forecasts on energy consumption in the next 

future show an increasing trend due also to growing markets, and among this the 

leading emerging economies connected to BRICS countries (Brazil, Russia, 

India, China and South Africa). We need more and more energy. On the other 

side the environmental problem is arising and the climate change is a more and 

more pressing emergency. The latest environmental disasters, as the oil spill in 

the Gulf of Mexico in 2010 and the earthquake with the following tsunami in 

Japan in 2011, struck the public opinion and so, as in many European countries, 

the nuclear program is under a deep revision or even stopped. 

So it is necessary to utilize as well as possible all energy resources. Among 

these resources, natural gas is undoubtedly one of the most used. A lot of gas 

pipelines were build and many others are under construction or in project. At the 

same time many countries utilizing gas are far from production countries, for 

example Japan that actually is reconverting the national energy plan pushing up 

on natural gas. Therefore the only way to utilize gas in these countries is to 

achieve it by transportation with ships, and the better way is in liquid state, so it 

is possible to maximize mass per unit volume. 

The transport of natural gas is made by gas pipeline from gas field to the 

harbour where are located the gasification terminal. In these industrial sites the 

gasification process takes place with consumption of the natural gas itself, i.e. an 

huge amount of energy is used in order to liquefy the gas. 

∗ Corresponding author: e-mail: vincenzo.larocca@unipa.it

2 Antonino La Rocca et al. 

At the terminal of arrival, the regasification process of LNG returns it back 

to gas state before its transport in the pipeline network. Again energy 

consumption is required and, moreover, a huge amount of cold is produced. The 

recovery of cold may be of capital interest because it has a relevant 

environmental impact, usually on the sea near the regasification site, and, by an 

energetic point of view, it is not a suitable operation to waste a lot of energy. 

So it is possible to consider both the production of electrical energy 

recovering the energy available as cold, using the cryogenic stream of LNG 

during regasification as cold source in an improved CHP Plant (Combined Heat 

and Power), and to utilize directly cold in some suitable process where cold is 

need. 

This paper, after a survey of the natural gas market, deals with some 

feasible process utilizing waste cold from regasification of LNG, showing an 

application in order to produce electric energy and another case study which 

reutilize cold directly. 

Key words: LNG regasification, energy efficiency, energy from waste, 

CHP plants. 

1. Introduction 

Nowadays the international markets are afflicted by a deep economic 

crisis. In Europe many countries was obliged to reduce their expenses, to 

increase taxation, to cut wages and jobs; many firms was obliged to restructure 

their factories and these policies led particularly to an increasing 

unemployment. In the Euro Area the unemployment rate in March 2012 reach 

the value of 10.9%, an increasing value from the value of 9.9% only one year 

ago. Among the European countries Spain and Greece have the highest values 

of unemployed, reaching respectively the values of 24.1% and 21.7% (data: 

Eurostat website). 

On the other side in the world the energy production and consumption are 

still increasing, but this is due only to the growing economies. European Union 

registers a negative trend in the primary energy production and, moreover, it has 

pledges to cut its energy consumption by 20% by 2020. 

The emerging economies connected to BRIC countries (Brazil, Russia, 

India and China) collectively consume around 27% of the world’s primary 

energy demand, 17% of oil demand, 19% of gas demand and nearly half of 

global coal demand. 

All these scenarios are reported in the latest IEA World Energy Outlooks. 

In these reports we can see an increasing demand of energy and the linked 

request to stop the climate change, related the use of fossil fuels. 

So the main keywords are: convert to renewables and use in a better way 

all the energy resources, performing also the CCS, i.e. the Carbon Capture and 

Sequestration, for fossil fuels if it is possible.

Bul. Inst. Polit. Iaşi, t. LVIII (LXII), f. 4, 2012 3 

The energetic scenario prospected by IEA and other statistical reports, as 

for example the BP Energy Outlook 2011, will be dominate again by fossil fuels 

and nuclear, with a more and more increasing role of all renewables in the next 

years until 2035. Considering the increasing discontent in many countries about 

nuclear energy we have to consider a shift on traditional reliable energy 

resources.A recent news point that Japan stop all nuclear reactors and many 

analysts doubt that these plant will operate again, due to the shock that struck 

the Japanese people after the tsunami. The reopening will be subject to an 

approval by Prefectures, a political choice that can’t take into account the public 

opinion; nuclear energy in the last years covered about 30% of energy 

production in Japan. Germany also is rethinking about nuclear energy and stated 

to stop in the next 10 years all the plants. In Italy there is no nuclear plant 

operating and the nuclear program, when some approach was made to 

reintroduce it, was stopped for many years, and maybe forever, by a referendum 

that took place in 2011. 

So despite all, the role of fossil fuels is still essential and among these, 

the natural gas (NG) isone of the most attractive. Many gas pipelines are 

operating now and many others will be realized in the next future. The 

Liquefied Natural Gas, LNG, will be increasing its role and it is easy to think 

that Japan, for example, actually the leader country for number of LNG 

regasification plants, will convert on LNG the lack of nuclear energy 

production. BP in his outlook points that the “global natural gas trade 

increased by a robust 10.1% in 2010. A 22.6% increase in LNG shipments was 

driven by a 53.2% increase in Qatari shipments. Among LNG importers, the 

largest volumetric growth was in South Korea, the UK and Japan. LNG now 

accounts for 30.5% of global gas trade. Pipeline shipments grew by 5.4%, led 

by growth in Russian exports”. 

2. The Use of NG and LNG 

The NG is a leader among the fossil fuels. In 2010 the world production 

was 3193.3 billion of cubic metres. The NG international trade was 975.22 

billion of cubic metres of which 297.63 as LNG. 

LNG made possible to use gas where isn’t possible to transport it by 

pipeline directly from the extraction wells. The natural gas reaches by pipeline 

the terminal plant near the sea. Here the liquefaction process take place and it 

uses energy given directly by the gas itself. So the natural gas, in liquid state, is 

stored in suitable tanks located into a ship that sails to the arrival terminal, 

where the gas is regasified and inserted into the distribution pipeline. The 

regasification process uses energy, given once again by the gas itself, and, at the 

same time, a huge amount of cold is produced and it need to be utilized or 

discharged in some way; usually cold is lost into the sea water utilized in the 

process.


2.1. Energy Recover from Regasification Process of LNG 

As suggested by several scientists (Cravalho et al., 1977; Buffiere & 

Vincent, 1972; Snamprogetti, 1978, Dispenza et al., 2002), there are several 

possibilities to recover energy from the regasification process of LNG. 

Among these we may consider to storage the cold energy by means of a 

suitable media carrier (a refrigerant), as may be Nitrogen or oxygen, and then to 

transport it back to the gasification site; this solution, very attractive from a 

thermodynamic point of view, is not practicable: in fact we need additional 

storage and process plants at both the gasification and the regasification site, 

moreover the handling of LNG and refrigerant is more complicated. Another 

possibility is to use the regasifying LNG to produce liquid oxygen and liquid 

Nitrogen. It is also possible to utilize cold directly in some foodstuff factory that 

need cold. Till now these options had no fortune because they involved a plant 

more complicated and economically there was no convenience. But now, with 

energetic costs so high, it is necessary to change point of view and operate with 

the maximum recover of energy. 

The production of electric energy may be realised in several ways 

(Dispenza et al., 2006). A first possibility is to enhance the overall electric 

efficiency of Electric Utilities in Power stations working with a cycle with 

Steam Turbines, lying near the regasification site, by lowering the temperature 

of the condenser utilizing the water rejected by the Open Rack units. The idea 

has been applied in Japan. 

Another new idea is applied also in Japan at Himeji LNG Terminal: it 

pertains to an application where cold energy of LNG during regasification, in an 

Electric Utility lying inside the Terminal area which has an Electric Power 

capacity of 50 MW and works with a combined cycle with Gas Turbine and 

Steam Turbine, is utilized to cool gas turbine inlet air. 

An improved Process was proposed in the past (Snamprogetti, 1978) by a 

patent which was proved in-field in the Regasification site of SNAM (ENI 

Group) located in Panigaglia, La Spezia, Italy. The process pertains the use of a 

CHP Binary Cycle composed of two Brayton Cycles: the Top Cycle is an open 

cycle working with a conventional Gas Turbine and the Bottom closed Cycle 

works with Nitrogen which is heated recovering the reject heat of the top cycle 

and then it goes into a cryogenic Heat Exchanger into which LNG is regasified. 

Now the development of innovatory technology in the field of Gas 

Turbines, Compressors and Heat Transfer Equipment and devices offers a new 

perspective. Then, it seems very interesting the analysis of some improved 

Cycles of such kind addressed to recovery exergy of cold during the 

regasification of LNG producing Electric Energy.


2.2 The Regasification Process of LNG 

The Italian natural gas transmission system works with major pipelines at 

a pressure rating up to more than 70 bar, then the LNG regasified has to be sent 

in the transport pipeline at a pressure of the same order of magnitude. 

Fig. 1 – Thermodynamic process during regasification of LNG (La Rocca, 2010). 

In principle it is possible to pump natural gas at this pressure after 

regasification at a low pressure (Fig. 1, path AB’D’D), but this option isn’t 

applied in Europe. When the regasification is carried out at low pressures and, 

then, the gas is compressed at the pressure required by pipelines, the pumping 

power, WNG, is very high. Instead, if the LNG is pumped in liquid phase, at a 

pressure lying in the pressure rating required by pipelines (usually a 

hypercritical pressure in Europe) there is convenience because the pumping 

power, WLNG, is lower (see Fig.1, path ABCD). The ratio WNG/WLNG is higher 

than 20. Moreover, with regasification at hypercritical pressures, the heat 

exchange by the LNG side is very good and heat transfer equipment is more 

reliable. 

In this paper a summary and some results of researches carried out at the 

Dipartimento dell’Energia of Palermo are reported. 

3. Production of Cold from Regasification of LNG 

The research developed on this subject includes both cryogenic 

applications and industrial use of cold at very low temperature inside the 

regasification site and cold utilization by end users far from the regasification 

site in deep freezing facilities and for space conditioning in the commercial 

sector (e.g.: Supermarkets and Hypermarkets). 

An application that want recovery cold from a regasification plant of 

LNG, need to be as close as possible to the regasification site, but this is very 

difficult due to safety restriction. LNG is flammable and a fire may take place in 

case of leakage, so it is necessary to fix safety distances with other production 

site or other activities.


There are still many problems to transfer cold recovered outside the 

regasification site, because the long distances need a suitable brine or safe 

fluids. 

In a previous work of the Research Team at Dipartimento dell’Energia of 

Palermo University, which Authors belong, is proposed the process shown in 

Fig.2, concerning a modular multipurpose regasification facility in order to 

recover LNG physical exergy for cold utilization. 

Fig. 2 –A proposed plant for cold recovery during LNG regasification 

(La Rocca, 2010). 

The design of the process allows the use of cold at two different 

temperatures: one lower for industrial use and one higher, suitable for other 

applications. Two applications utilizing cold recovered are here considered: a 

cluster of Agro Food Factories and a large Hypermarket. 

The plant proposed is modular and has a regasification capacity of 2 × 10 9 

Stm 3 /y when working 270 d/y 24 h a day all the year around. 

The facility requires an intermediate Power Cycle, working with a 

suitable fluid such as Ethane, able to produce 3MWof electric power, but the 

own purpose of the modular unit is to deliver cold suitable for industrial and 

commercial use in the proper temperature range utilizing Carbon Dioxide as 

secondary fluid to transfer cold from regasification site to far end users.


Fig. 3 reports a simplified thermodynamic cycle of the Power Cycle 

working with Ethane which is included in the Modular Unit. 

The evaporator section is constituted by KE1, KE2, KE3 and there is at 

exit a dryer/separator. The preheater section is represented by the Heat 

exchanger CU. The condenser section includes the Ethane condenser and the 

Ethane receiver from which liquid Ethane is pumped in CU. Ethane pump is 

driven by a low pressure Ethane expander, while GT is the Ethane power 

expander. 

Fig. 3 – Power Cycle working with Ethane (La Rocca, 2010). 

In the process (Fig. 2) cold can be produced at lower temperatures 

suitable for industrial applications (-90 °C, -80 °C, -35 °C) in the Heat 

Exchangers CU (4 MW), CU1 (9.5 MW) and at higher temperatures (-35 °C up 

to 0 °C) in the Kettle type evaporators KE1 (3.5 MW), KE2 (14.2 MW), KE3 (5 

MW) and in the Heat Exchanger CU2 (16.1 MW). The Open Rack units will be 

used to manage the matching of duty. 

Two other Service Cycles working with CO2 are included in the design of 

the modular plant related facilities (Figs. 4-5) in order to transfer cold to the 

end-user. 

Part of the cycles operates inside the regasification site: condensing the 

gaseous CO2 phase returning from the end users facilities and pumping the 

liquid phase of CO2 obtained in the heat exchangers CU1 and KE2 (respectively 

for the Agro Food Factories and for the Hypermarket: note also that Kettle 

reboilers 1, 2, 3 are used as heating heat exchangers for the power cycle 

working with Ethane recovering cold which is utilized for a lot of processes, 

Open Rack units operate when the modular unit is standing). 

Part of the cycles operates inside the end users facilities: the CO2 liquid 

phase evaporates releasing cold. The transfer of liquid phase of CO2 from the 

regasification site to end users is made via a liquid phase CO2 pipeline (CO2 is 

pumped in liquid phase inside the regasification facility). The transfer of 

gaseous phase of CO2, obtained during the process of refrigeration inside the


end users facilities and returning to the regasification site, is made via a gaseous 

phase CO2 pipeline (the pressure losses along the flow path are accounted in the 

design of the service cycles, see Figs. 6 and 7). 

Fig. 4 – The Service Cycle for the Agro Food Factories (La Rocca, 2010). 

Fig. 5 – The Service Cycle for the Hypermarket (La Rocca, 2010). 

The pumping process of the CO2 in liquid phase is a good practice for 

saving energy (Dispenza et al., 1979). 

The process proposed works at hypercritical pressures (LNG in the 

regasification process has pressure lying in the range 80 ÷ 75 bar). 

3.1. The Feasibility Study 

The recovery of cold from LNG regasification in the proposed process 

considers two kind of facilities as end user: first a cluster of Agro Food 

Factories, where there is need of cold for the freezing process of foods and to 

maintain these at low temperature; second Supermarkets and Hypermarkets 

where the need of cold is mainly for space air conditioning and some cold is 

required for the refrigeration of display cases and cabinets in the sales area and 

for the preparation of foodstuffs and their cold storage in cold storage rooms.


The two clusters have a mean distance of 2 km from the regasification site. 

Fig. 6 – The cooling cycle for transfer an utilization of cold in Agro Food Factories 

diagram: courtesy of IIR (La Rocca, 2010). 

Fig. 7 – The cooling cycle for transfer an utilization of cold in Hypermarket 

diagram: courtesy of IIR (La Rocca, 2010).


The mean single regasification capacity in the regasification sites planned 

in the World is 8 × 10 9 Stm 3 /y. The modular regasification facility reported in 

Fig. 2 has a regasification capacity of 2 × 10 9 Stm 3 /y and it has a potential 

capacity to delivery cold power, recovered for cold utilization, of more than 

50MWand can produce 3 MW of electric power. 

The power required by the cluster including Agro Food Factories 

amounts to 9 MW of cold. The duty can be satisfied by the use of cold delivered 

in the exchanger CU1. 

The end use in the Hypermarket pertains mainly to the process of space 

air conditioning and a quantity of cold is required in the refrigeration utilities. 

The whole amount of power required is of 6 MW of which 100 kW delivered as 

cold at -35 °C and the remaining part delivered as cold at -10 °C. The whole 

duty can be satisfied by the use of cold delivered in the Kettle evaporators KE2. 

The remaining cold recovered can be used inside the regasification site 

for a lot of other processes which requires cold. 

3.2. The Transfer of Cold Between the Regasification Facility and the End Users 

This phase of the research activity at Dipartimento dell’Energia of 

Palermo University required a thorough analysis of a lot of options. The transfer 

of cold between the regasification facility and the clusters of Agro Food 

Factories and the Hypermarket takes place by means of two pipelines into 

which travels Carbon Dioxide: in the feeding pipeline it is in liquid phase and in 

the return pipeline it is in gaseous phase. 

Carbon Dioxide is liquefied in the regasification facility recovering cold 

available in the regasification process (Fig.2, Units: CU1 and KE2) and it is 

pumped in liquid phase. The selected option allows a considerable saving of 

pumping power (e.g.: it results for the Case Study of the Agro Food Factories 

that for the gaseous phase it would be more than 30 time higher). Carbon 

Dioxide in liquid phase goes to the clusters of users in each factory and feeds 

evaporators of cold utilities while the gaseous phase returns back to the 

regasification facility where it is liquefied and then pumped in the feeding 

pipeline. 

3.3. Agro Food Factories 

A process sheet of the condensation facility and of the liquid Carbon 

Dioxide pumping station lying in the regasification site is reported in Fig. 4 

(exchanger CU1, receiver A and Carbon Dioxide Liquid Pump). In the same 

Figure it is reported a schematic sheet indicating some clusters of end users 

(CLi), the storage system (Bi) of the liquid Carbon Dioxide installed ahead of 

the end users in a Factory. The gaseous phase returning from the evaporators 

comes in the receiver (Ci). From this device installed in each Factory it goes 

(almost dry) in a common receiver (C, installed in the cluster area) and then it


goes in the return pipeline collecting all the gaseous Carbon Dioxide returning 

from the whole cluster of Factories. 

Fig. 6 represents the process which takes place in the secondary line 

(which includes the Carbon Dioxide pipeline) on the p-h diagram. The line 1-2 

indicates the condensation process of the gaseous phase which returns from the 

Agro Food Factories. The evaporation process takes place in the evaporators (at 

the end user, line 5-6). The pumping of the liquid phase is indicated by the line 

2-3 and line 3-4 indicates the pressure losses along the path in which the liquid 

phase travels. Line 4-5 indicates, for sake of simplicity, the whole adiabatic 

throttling process which takes place in all evaporators. For sake of simplicity 

pressure losses are not represented on the diagrams, but these were duly taken 

into account in the study. 

The line 6-1 indicates the process which pertains to the pressure losses 

during the return of the gaseous phase. The pipeline ahead of the end uses has a 

mean diameter of 8 inch (thickness 8.18 mm, pressure rating 40 bar, seamless 

steel API 5L). The pipeline downstream of the end uses (return pipeline) has a 

mean diameter of 18 inch (thickness 11.13 mm, pressure rating 40 bar, seamless 

steel API 5L). The insulation material is polyurethane foam, which has high 

thermal efficiency and is mechanically strong. The system is capable of working 

in a temperature range of -200 °C to 150 °C. The condensation and pumping 

facility (in the regasification site) and the facility lying in the cluster will be 

monitored on-line by means of a master PC (cluster controller) which receives 

signals by the data acquisition and control peripherals (remote units, at the end 

uses). The main task of the supervision system will be to assure the right 

feeding of the evaporators at the end use. 

3.4. Hypermarket 

Fig. 5 shows a process sheet of the condensation facility and of the liquid 

Carbon Dioxide pumping station lying in the regasification site (exchanger 

KE2, receiver A’ and Carbon Dioxide Liquid Pump). In the same Figure it is 

reported a schematic sheet of some end users, the storage system (B’) of the 

liquid Carbon Dioxide installed ahead of the end users. In the receiver (C’) 

comes the gaseous phase returning from the evaporators; from this device it 

goes (almost dry) in the return pipeline. In Fig. 7 it represents the process which 

takes place in the secondary line (which includes the Carbon Dioxide pipeline) 

on the p-h diagram. The processes represented both for the end use at -35 °C 

and for the end use at -15 °C, are similar to those explained for Fig. 4. 

Note that here there is need to equip the facility with a compressor (see 

Fig. 5) which pumps the gaseous phase coming out from the evaporators 

working at a lower temperature up to a pressure level suitable for mixing the 

gaseous phase with that coming out from the Air Treatment Units in the 

Hypermarket. The pipeline lying ahead of the end uses has a mean diameter of 6


inch (thickness 10.97 mm, pressure rating 80 bar, seamless steel API 5L). The 

pipeline lying downstream of the end uses (return pipeline) has a mean diameter 

of 12 inch (thickness 17.48 mm, pressure rating 80 bar, seamless steel API 5L). 

The insulation material is polyurethane foam, which has high thermal efficiency 

and is mechanically strong. The system is capable of working in a temperature 

range of -200 °C to 150 °C. 

3.5. The Regasification Modular Unit, Thermodynamic Analysis 

The regasification facility is a cogeneration system which has the aim to 

produce electric power and release cold at various temperature levels and to 

regasify LNG. 

The cogeneration system proposed uses as heat source the cooling duty of 

various kinds of refrigeration processes, as cold source the LNG to be regasified 

(first step of regasification process) is used. Other cold is recovered in the heat 

exchangers CU1 and CU2. The main design criteria adopted in the study for the 

regasification modular unit are the following: 

i) there is need to manage cold for some process different from cryogenic 

applications at a suitable temperature lying in a range allowing the use of 

available secondary fluids; 

ii) following the above criterion the facility includes also a suitable power 

plant heated by heat released by secondary fluids carrying cold far from the 

regasification site or inside the same regasification site; 

iii) a part of cold available during the regasification process of LNG can 

be used inside the same regasification site area (e.g.: air liquefaction producing 

Nitrogen, Oxygen, Argon, deep freeze industrial warehouse and cold storage 

warehouse, etc.). 

The main operating parameters of the power plant thermodynamic cycle 

working with Ethane (pressures and temperatures) shall be selected on the basis 

of some criteria adopted in design: 

i) lower pressure shall be higher than external atmospheric pressure (to 

avoid air infiltration in the plant: in the feasibility analysis it was selected as 1.3 

bar); 

ii) higher pressure shall be selected at a value allowing to have a 

temperature near -40 °C, in dependence of working fluid selected; 

iii) Ethane: lower pressure 1.3 bar (tsat=-83.85 °C), higher pressure 7.8 bar 

(tsat=-39.91 °C). 

The pinch temperature difference in the condenser between the working 

fluid and LNG (at the exit of the device) was selected as 10 K. 

Numerical simulations have been performed in the study with an 

algorithms developed at Dipartimento dell’Energia, using as a main tool the 

software EES, Engineering Equation Solver, of S.A.Klein and A.Alvarado, 

release 7.934.


Making thermodynamic analysis it is useful to analyse some details: 

The regasification process using Open Rack technology is fully 

dissipative, moreover cold is released in the sea near the regasification site and 

it gives rise to environmental problems. 

Some other regasification technologies (e.g.: Fired heaters with 

water/intermediate fluid, Submerged combustion vaporizers, CHP Plants) use a 

lot of LNG vaporized, although in improved CHP Plants (e.g.: Dispenza et al., 

2009 a-b) a considerable cold exergy recovery there is, this improves 

noteworthy the plant performance. 

The modular unit proposed don’t use heat produced burning LNG 

regasified, moreover it allows to recovery cold physical exergy to be used in a 

lot of various processes and the regasification process reaches a suitable 

thermodynamic performance reducing process irreversibility. 

The use of CO2 as a secondary cold transfer fluid requires the study of 

innovative systems into which dissipative service cycles are used, these are 

based on the concept of: pumping CO2 in liquid phase, transfer of CO2 in liquid 

phase to end users, use of CO2 in liquid phase evaporators to end users to 

delivery cold available for refrigeration processes, transfer of CO2 in gaseous 

phase returning it to regasification site. 

The exergy of LNG and of NG considered in the Thermodynamic 

analysis is the Physical Exergy. Actually, all processes which evolve in the 

Modular Unity proposed (Dispenza et al., 2009 c) don’t include burning of 

LNG regasified. The input exergy of LNG to be regasified include the physical 

exergy of cold and the physical exergy due to pumping LNG in liquid phase. 

The output exergy of NG is its physical exergy. The Dead State considered is a 

Restricted Dead State: the fixed quantity of matter under consideration is 

imagined to be sealed in an envelope impervious to mass flow, at zero velocity 

and elevation relative to coordinates in the environment, and at the temperature 

T0 and pressure P0. Thermodynamic analysis includes both Energy and Exergy 

analysis. A methodology has been built, duly defining each parameter included 

in the detailed analysis reported below (Nomenclature). Main results obtained 

derived by numerical simulation using the model derived at DREAM (Dispenza 

et al., 2009 c) are reported later. 

For sake of clearness some results of Exergy analysis are shown later in 

Fig. 8 in graphical format (Sankey Diagrams). 

3.6. Energy and Exergy Efficiency of the Modular Unit 

In the plant working with Ethane cold is delivered to a secondary fluid or 

a cooling process in the heat exchangers CU, KE1, KE2, KE3 (Fig. 2). 

At the same time Ethane is heated and it vaporizes. The cold duty of each 

device composes the whole cold duty for this part of the regasification facility.


The condenser of the Ethane Cycle is cooled by the LNG which is to be 

regasified. Moreover the completion of the LNG regasification takes place in 

heat exchangers CU1 and CU2. The cold released in the equipment is also 

utilized. An appropriate Energetic efficiency parameter (Dispenza et al., 2009 c) 

for such complex kind of modular plant is a COP which is defined as follows. It 

can be estimated as the percentage recovered of cold energy supplied to the 

system (in the condenser of the Ethane Cogeneration Plant and in the LNG heat 

exchangers CU1 and CU2): 

COP Ethane Cogeneration Plant. The COP for this part of the modular 

plant is defined by the formula 

Pel + QCU + QKE1 + QKE2+ QKE3 QR 

COPEthCP 

= × 100 = × 100, (1) 

Q Q 

R el 

cond cond 

Q = P + Q + Q + Q + Q 

CU KE1 KE2 KE3 

. (2) 

COP of the Whole Regasification Facility. This COP is defined by the 

formula 

COP 

reg 

Q + E + Q + Q 

= 

Q + P 

R NG CU1 CU2 

LNG LNG 

× 100. 

The Exergy efficiency ε for the modular plant can be estimated as the 

percentage recovered of cold exergy supplied to the system (the exergy of the 

LNG stream to be regasified). 

Exergy Efficiency of Ethane Cogeneration Plant. It is defined by the 

formula 

(3) 

Eroare! . (4) 

Exergy Efficiency of Whole Regasification Facility. It is defined by 

ε 

Re g 

ERE + ENG + ERCU1+ ERCU 

2 

= × 100. 

(5) 

E 

LNG 

3.7. Main Results of Thermodynamic Analysis 

In this section main results of thermodynamic analysis performed by 

numerical simulation are reported. Considering the power cycle working with 

Ethane represented in Fig. 3, main parameters pertaining the numerical 

simulation for a typical steady-state assessment has been derived. 

Mean pressure level in the kettle reboilers is 7.8 bar (-40 °C) and the 

vapour is superheated at a temperature of 10 °C reaching point 5. Mean pressure 

level in the Ethane condenser is 1.3 bar (-84 °C). Isentropic efficiency of the


Ethane power expander was assumed 0.89, while isentropic efficiency of the 

Ethane low pressure expander (which drives Ethane pump) 0.83. Isentropic 

efficiency of the Ethane pump is assumed to be equal to 0.86; while the 

efficiency of the Electric generator and mechanical efficiency of the system are 

respectively equal to 0.94 and 0.91. Ethane flow rate is 35.4 kg/s. LNG flow 

rate to be regasified is 62.7 kg/s, when the modular unit operates 9 months/year 

it regasifies 2 × 10 9 Stm 3 /year. 

For the case study analysed for the modular unit proposed the significant 

Energetic efficiency parameters above defined have the following 

values:COPEthCP= 161% and COPReg= 201%, while the Exergy efficiency as 

above defined have the values: εEthCP= 35% and εReg= 64%. 

It is worth of mention that ε would be a 42% if the same Regasification 

Plant of Fig. 2 regasifies LNG at a pressure below that pseudo-critical, as 

proposed in eastern countries. Moreover, ε would be a 48% for a simple 

regasification plant based on Open Rack technology working at hypercritical 

pressure and a 32% when the same works at a pressure below that of the 

pseudo-critical state of LNG. Note that, the exergy “recovered” in both the last 

two cases is only that of the natural gas regasified. 

The simulation analysis for the whole regasification facility, represented 

in Fig. 2, gives the main results reported in Fig. 8, which shows main results 

obtained performing the Exergy analysis of each part of the process. 

The input exergy is represented in Fig. 8 by the exergy of the stream of 

LNG to be regasified. This quantity includes the exergy of cold contained in the 

LNG stream to be regasified and the LNG liquid pumping power. 

The Dead State parameters used in the Exergy analysis are: P0= 1 atm 

and T0 = 15 °C (288.15 K). 

The flow rate of LNG to be regasified in the plant is 62.7 kg/s. 

kW. 

Fig. 8 – Main results of exergetic analysis (La Rocca, 2010). 

The exergy of LNG at inlet in the regasification facility (ELNG) is 69,666


In Fig. 8: ENG (47.7% of ELNG) is the exergy of the exit stream of NG 

regasified; EL.CO (25.7% of ELNG) is the exergy of cold available during the 

regasification process of LNG delivered in the Ethane condenser (see Fig.2); 

EL.CU1 (6.4% of ELNG) is the exergy of cold available during the regasification 

process of LNG delivered in the LNG cryogenic heater CU1; EL.CU2 (4.3% of 

ELNG) is the exergy of cold available during the regasification process of LNG 

delivered in the cryogenic heater CU2; ED.PS (15.9/% of ELNG) represents the 

exergy destroyed by the whole amount of pressure losses in the regasification 

facility of Fig. 2 by the LNG side. 

Fig. 8 shows also a detail of the above parameters and splits each part of 

exergy exchanged in the regasification facility indicating the item pertaining to 

exergy recovered in the Ethane condenser, in the cryogenic LNG heaters and 

the item pertaining the pressure losses in the whole regasification facility. 

The exergy delivered in the Ethane condenser requires a thorough 

analysis, the results obtained are shown in Fig. 8. The item ER.EL is the part of 

LNG exergy which is recovered as electric power produced by the Power Cycle 

working with Ethane. This cycle has a mean overall efficiency of a 14% and it 

produces a rated power of 3.2 MW (16.7% of input exergy delivered by LNG in 

the Ethane condenser), exergy losses in the Ethane pump unit (ED.PU) amounts 

to a 0.10% of input exergy. The item ER.PU is the part which is recovered and 

then used to drive the Ethane pump and it represents a 0.30% of input exergy. 

The ER.CU indicates the exergy recovered in the Ethane preheater CU (see Fig. 2) 

and ED.CU indicates the exergy destroyed, the nomenclature used in the Fig.8 

with the prefix ER (e.g.: ER.CU1) indicates the exergy recovered in the device 

named with the same symbol in Fig. 2, those with the prefix ED (e.g.: ED.CU1) 

indicates the exergy destroyed in each device. The item ED.P (7.1% of input 

exergy) indicates the exergy destroyed by the pressure losses along the Ethane 

Loop. 

4. Production of Electric Energy from Regasification of LNG 

The analysis performed at Dipartimento dell’Energia, Palermo 

University, on this subject includes two ventures: one of them pertains to a CHP 

Binary Cycle composed of two Brayton Cycles whose Bottom cycle works with 

Helium and another has the Bottom cycle working with Nitrogen but the 

pressures of the cycle have been selected, after an optimization analysis, higher 

than those pertaining the Process described in (Snamprogetti, 1978). Heat 

transfer Equipment proposed are also analysed designing a suitable heat transfer 

matrix to improve the processes analysed. 

4.1.The Process Proposed 

The process proposed for this application is shown in Fig. 9. The Top 

cycle includes a modern gas turbine system working at high temperatures which


is shown in the higher part of figure at left. The bottom Cycle, which is shown 

in the higher part of figure at right, works with Helium, the higher pressure is 

22.6 bar and the lower is 3 bar; the temperature of Helium at the inlet in GTbottom 

turbine is 480°C, then Helium coming out of gas turbine gives heat to LNG in 

the Cryogenic Regasifier and it is cooled down to -129°C, it is then pumped by 

the Cryogenic Compressor C2 and at the exit it has a temperature of 74°C. 

Fig. 9 – The modular process proposed for regasification of LNG 

(Dispenza et al., 2006). 

The heat exchanger managing the LNG to be regasified has a heat 

transfer matrix made with tubes having a special extended surfaces (Fig. 10). 

LNG crosses the units from the tube side while Helium flows from the 

shell side. Tubes have turbulence promoters inside and the hollow tie rods on 

the shell side have windows of rectangular shape with rounded corners, this 

feature assures lateral injection in the stream crossing the matrix enhancing 

turbulence. 

In the plant proposed there are four Heat Exchangers: a train of two is in 

series and then the trains are connected in parallel. The shell of a unit has an 

external diameter of 1.20m and is 8.50m long; there are 300 tubes of Stainless 

Steel suitable for cryogenic exploitation which have an external diameter of 

(3/4)”. The cold box has the following dimensions: 3.60m ×3.60m ×9.50m. The 

exchangers for heating Helium are two units with a feature of the heat transfer


matrix seeming that of Fig.10, but the inner of tubes has also an extended 

surface and there are inside turbulence promoters. The shell of a unit has an 

external diameter of 2.90m and is 12.00m long; there are 450 tubes of Stainless 

Steel which have an external diameter of 1.5 inch. 

Fig. 10 –Tubes proposed for the cryogenic regasifer of LNG (Dispenza et al., 2009 a). 

The Open Rack facility is required for operation for which isn’t produced 

Electric Energy or during maintenance operation. The Plant layout for a module 

with a regasification capacity of 2 ×10 9 Stm 3 /y requires a whole surface of a 

1,500 m 2 . The Plant operating with Nitrogen has seeming feature, but it requires 

higher pressures inside the Bottom Cycle and its performance is lower. 

The use of Helium calls for R. and D. on the related technology for a 

large worldwide diffusion of this kind of CHP Plant, but it is very interesting to 

look forward. Helium is easy available in USA, Russia and Italy, moreover, as 

far as it pertains to the technology, much R. and D. was carried out in past years 

for Nuclear Reactors Power Cycles. 

The use of Nitrogen has the advantage that technology is easy available 

on commission in the Petrochemical World. 

4.2. The Proposed CHP Plants 

Main characteristics of CHP Binary Cycles proposed are reported in the 

following sections. Some other features of the CHP cycle proposed have been 

described previously. 

4.2.1. The Proposed CHP Plants Working with Helium. The values 

reported pertain to a modular plant with a regasification capacity of 2 ×10 9 

Stm 3 /y when working 24 h a day all a year around (270 days). The mean net 

power inlet in bottom cycle is 74.32 MW, the mean Electric production 27.53 

MW and 44.03 MW is the thermal power exchanged in the cryogenic regasifier. 

The mean electric efficiency is of a 37%. The mean electric efficiency of top 

cycle is of a 27%; the thermal power furnished is 168.7 MW and gross thermal 

power recovered by flue gas is 80.78 MW while the Electric power produced is


44.96 MW. The gas consumption is a 5,7% of the quantity regasified but only a 

3% is attributable to the regasification process. The overall electric efficiency of 

the CHP cycle is 43% and the whole efficiency of the CHP cycle is 69%. 

Considering a mean Lower Heating Value for NG of 36,209 kJ/Stm 3 (8,560 

kcal/Stm 3 ) the gas consumption is 108.69 ×10 6 Stm 3 /y which correspond to 

94.00 ktoe/y of primary energy. The amount of electric energy produced is 

469.73 GWh/y. Bearing in mind these values, the primary energy need in Italy 

to produce the same amount of electric energy is of 103.33 ktoe/y, so there is a 

primary energy saving of 9.33 ktoe/y. 

4.2.2. The Proposed CHP Plants Working with Nitrogen. The process 

proposed is seeming to the other working with Helium (Fig. 9). The Top cycle 

includes a modern gas turbine system working at high temperatures which is 

shown in the higher part of figure at left. The bottom Cycle, which is shown in 

the higher part of figure at right, works with Helium, the higher pressure is 37.4 

bar and the lower is 5 bar; the temperature of Nitrogen at the inlet in GTbottom 

turbine is 479°C, then Nitrogen coming out of gas turbine gives heat to LNG in 

the Cryogenic Regasifier and it is cooled down to -129°C, is then pumped by 

the Cryogenic Compressor C2 and at the exit it has a temperature of 0°C. The 

values reported pertain to a modular plant with a regasification capacity of 2 

×10 9 Stm 3 /y when working 24 h a day all a year around (270 days). 

The mean net power inlet in bottom cycle is 67.96 MW, the mean 

Electric production 22.28 MW and 44.03 MW is the thermal power exchanged 

in the cryogenic regasifier. The mean electric efficiency is of a 33%. The mean 

electric efficiency of top cycle is of a 27%; the thermal power furnished is 146.3 

MW and gross thermal power recovered by flue gas is 70.04 MW while the 

Electric power produced is 39.00 MW. The gas consumption is a 4,9% of the 

quantity regasified but only a 3% is attributable to the regasification process. 

The overall electric efficiency of the CHP cycle is 42% and the whole 

efficiency of the CHP cycle is 72%. Considering a mean Lower Heating Value 

for NG of 36,209 kJ/Stm 3 (8,560 kcal/Stm 3 ) the gas consumption is 94.24 ×10 6 

Stm 3 /y which correspond to 80.00 ktoe/y of primary energy. The amount of 

electric energy produced is 397.05 GWh/y. Bearing in mind these values, the 

primary energy need in Italy to produce the same amount of electric energy is of 

87.00 ktoe/y, so there is a primary energy saving of 7.00 ktoe/y. 

4.3. Comparison Between CHP Modular Plants Working with Helium or Nitrogen 

The advantage offered by use of innovative modular CHP plants 

proposed in comparison with OR evaporators technology is mainly attributable 

to reduced environmental impact of cold released in the sea near the 

regasification site (mean avoided release of cold: 740…750 kJ/kg of LNG 

regasified); other benefits follow from the use of such a kind of CHP innovative


plant (e.g., saving of fossil energy resources, reduced greenhouse gas emission, 

etc.). 

As far as it pertains to regasification, considering prudentially the whole 

cost of production only attributable to electricity production, regasification costs 

are included in a low range because, actually, only the incidence of 

regasification of LNG using the OR back-up system during the period of its 

operation is to be considered apart. 

Considering the two different kind of innovative modular CHP plants, 

operating respectively with Helium or with Nitrogen, for the same quantity of 

LNG regasified (requiring a thermal energy of 740…750 kJ/kg of LNG) a 

modular plant working with Helium produces an amount of electric energy of 

0.38 kWhe/kg of LNG regasified while a modular plant working with Nitrogen 

produces 0.29 kWhe/kg of LNG regasified (24% less). 

Cryogenic heat exchangers for regasification of LNG have the same 

nominal duty (45 MWt) and their cost is almost the same (5% less for Nitrogen 

plants, note that Nitrogen works at higher pressures and its molecular weight is 

higher than that of Helium, but this last has better thermo-physical properties 

and the mass flow rate required is less: 0.5 kgHe/kgLNG; while for Nitrogen 1.8 

kgN2/kgLNG). 

The duty of a compressor using Nitrogen gas is less than 55%, compared 

with that using Helium, and a gas turbine using Nitrogen gas has a duty less 

than 44% compared with that using Helium. As a result of a thorough feasibility 

study, done by the Authors and pertaining to conceptual design of machines to 

be used in CHP plants working with Helium or with Nitrogen), on the contrary, 

it results that CHP modular plants working with Helium, for the same quantity 

of LNG regasified, can produce 24% more electric energy. 

OR technology for a nominal regasification capacity of 2040 MStm 3 /y 

(287 days/y, 24 h/day; 2.6 months/y not in operation) requires 11.728 GWh/y of 

electric energy for pumping sea water (it is considered a mean decrease of 8 °C 

of sea water temperature from inlet in OR to exit) and for auxiliary services of 

facility. The CHP plants proposed when working both with Helium or Nitrogen 

have a mean electric production efficiency ηe = 0.43 (a little more when Helium 

is used). 

Note that in Italy, now, combined cycles of electric utilities (top gas 

turbines and bottom steam turbines) have been built using modern equipment 

and have a mean electric production efficiency of 52%. But, modular CHP 

plants proposed for LNG regasification based on Brayton cycles working with 

inert gases allow a saving of 11.728 GWh/y (equivalent electric power need of 

7.7MWe for pumping sea water when OR technology is used). Thus, to have a 

figure for comparison of order of magnitude, a mean overall electric efficiency 

can be accounted for 49% working with Helium and 47% working with 

Nitrogen. Considering the whole amount of electric utilities operating in Italy, 

penalty paid results in 3% working with Helium and 5% working with Nitrogen.


4.4. Thermodynamic Analysis 

Thermodynamic analysis is based on a careful numerical simulation 

which has been performed with a software developed at DREAM, using as a 

main tool the EES, Engineering Equation Solver, of S.A. Klein and A. 

Alvarado, Professional version 6.567. The results obtained by exergetic analysis 

are presented separately for the plant working with Helium and the plant 

working with Nitrogen. 

The regasification facility is a CHP (cogeneration system), which has the 

aim to produce electric power and heat to regasify LNG. It can be defined a 

process efficiency based on the first thermodynamic principle (energetic 

efficiency) and a process efficiency based on the second thermodynamic 

principle (exergetic efficiency) by two different point of view. 

4.4.1. Energetic Efficiency: (a) First definition: The whole facility 

(see Fig. 9) can be considered a CHP system producing electric energy and heat. 

Then, the input energy is represented by heat furnished in the top cycle burning 

natural gas in the top turbine gas generator - CC. The outputs are: electric power 

generated in the top and bottom Brayton cycles and heat supplied to LNG which 

regasifies in cryogenic regasifiers. This approach allows to define the first 

thermodynamic principle energetic efficiency with 

η 

NHP 

Peltop + Pelbottom + Qreg 

= . 

(6) 

Q 

NGburned 

(b) Second definition: Heat furnished in top cycle, the energy of 

cold furnished by LNG to be regasified and the pumping power of LNG stream 

in liquid phase, which regasifies at hypercritical pressures, are considered as 

input items. As output items they can be considered: the electric power 

generated in the top and bottom Brayton cycles and the exergy available in the 

NG regasified at a pressure suitable for direct transfer in pipeline network. So, a 

COP can be defined as a first thermodynamic principle energetic efficiency 

parameter with 

COP 

CHP 

Pel + Pel + Ex 

= 

Q + Q + W 

top bottom NG 

NGburned OLNG pLNG 

The parameters defined by the relationships of Eqs.(6) and (7) for the 

CHP plants analysed are: 

ηCHP COPCHP 

Working fluid in bottom cycle: Helium 0.69 0.51 

Nitrogen 0.72 0.64 

4.4.2. Exergetic Efficiency. (a) First definition: This kind of 

analysis (second thermodynamic principle) considers the plant as a CHP plant 

. 

(7)


for cogeneration of electric energy and heat to be supplied at LNG to be 

regasified. The exergetic efficiency ζCHP for the cogeneration system can be 

defined by 

η 

CHP 

Peltop + Pelbottom + ExQreg 

= . 

(8) 

Ex 

NGburned 

(b) S e c o n d d e f i n i t i o n : Exergetic efficiency of the CHP cycle is 

defined by 

ζ 

CHP 

Pel + Pel + Ex 

= 

Ex + Ex + Ex 

top bottom Ng 

NGburned air LNG 

where ExLNG=ExcoldLNG+WpLNG; Exair is the exergy of air going to the suction side 

of top cycle compressor; air is cooled by sea water coming out from OR units. 

The definition of Eq (9) properly accounts for all kind of exergy items 

involved in such a complex process. The parameters defined by the 

relationships of Eqs (8) and (9) for the CHP plants analysed are: 

ζCHP ζ * CHP 

Working fluid in bottom cycle: Helium 0.51 0.49 

Nitrogen 0.53 0.47 

The exergy efficiency defined by the Eq (9) is the most appropriate, 

because it accounts for all kind of exergy item involved in such a complex CHP 

plant. The results show that a plant working with Helium has a higher exergetic 

efficiency. 

4.5. Results of Exergetic Analysis for the Plant Working with Helium. 

A detailed analysis has been performed, considering all items which 

compose the whole top cycle process. There is a need for this approach to 

derive a comprehensive synthesis of exergy losses and exergy recovered. Fig. 

11 shows results obtained for the top cycle of a CHP plant working with 

Helium. Input exergy is given by 

the exergetic efficiency is defined as 

. 

(9) 

ExNGburned+ Exair= 169.975.79 kW; (10) 

ζ 

topCHP 

Pel + Ex 

= 

Ex + Ex 

top Qrec 

NGbumed air 

, 

(11) 

the input item includes the exergy of heat power furnished in top cycle (gas 

generator CC, Fig. 9) and exergy of air going to the suction side of top cycle 

compressor; air is cooled by sea water coming out from OR units. In Fig. 11 

items are reported in detail of exergy losses in main plant components and


exergy output. Fig. 12 shows the results obtained for the bottom cycle working 

with Helium. Input exergy is given by 

Ex + Ex = 110.738.00 kW . (12) 

ING Qrec 

The input item includes the exergy of LNG to be regasified and the 

exergy of heat power furnished in bottom cycle, which is recovered by flue gas 

of the top cycle. The exergy efficiency for the bottom cycle is given, then, by 

Pelbottom + ExNG 

ζ bottomCHP = 

. 

(13) 

Ex + Ex 

LNG Qrec 

Exergy output is represented by electric power generated in bottom cycle 

and the exergy of stream of natural gas regasified. In Fig. 12 items are reported 

in detail of exergy losses in main plant components and exergy output. 

Fig. 11 – Exergetic analysis:CHP with Helium, top cycle (Dispenza et al., 2009 a). 

Fig. 12 –Exergetic analysis:CHP with Helium, bottom cycle (Dispenza et al., 2009 a) 

4.6. Results of Exergetic Analysis for the Plant Working with Nitrogen 

An exergetic analysis has been performed considering all items which 

compose the whole top cycle process. Fig.13 shows the results obtained for the 

top cycle of a CHP plant working with Nitrogen. Input exergy is given by


ExNGburned + Exair 

= 136,455.22 kW. 

(14) 

The exergetic efficiency is defined by means of Eq (11). Fig. 14 shows 

the results obtained for the bottom cycle working with Nitrogen. Input exergy is 

given by 

ExLNG+ ExQrec = 100,021.00 kW. (15) 

The exergetic efficiency is defined by means of Eq (13). 

Fig. 13 –Exergetic analysis: CHP with Nitrogen, top cycle (Dispenza et al., 2009 a). 

Fig. 14 –Exergetic analysis: CHP with Nitrogen, bottom cycle (Dispenza et al., 2009 a). 

5. Conclusions 

1. In this paper is presented a wide review of possibilities of energy 

recover from LNG regasification analysed by the Research Team of 

Dipartimento dell’Energia, University of Palermo, which Authors belongs. 

2. In this works is shown how is possible to utilize a process, the LNG 

regasification, recovering the cold energy that otherwise have to be discharged 

in the sea water. The cold may be utilised in the regasification site in order to 

help electric energy production with innovative CHP plants or far from the site 

in activities that need cold energy, avoiding the consumption of electric energy


produced in some typical power plant and limiting the emissions of greenhouse 

gases and the consumption of other fossil fuels. 

3. Maybe these proposal can effectively contribute to the objectives fixed 

by various Conferences on Climate Change and Sustainable Development. 

Nomenclature 

CHP Combined Heat and Power plant 

COP Coefficient of performance 

E, Ex Exergy (kW) 

ED Exergy destroyed (kW) 

EL Exergy of cold available during the regasification process of LNG (kW) 

ELNG Exergy of LNG to be regasified (kW) 

ENG Exergy of the exit stream of NG regasified (kW) 

ER Exergy recovered (kW) 

Exair Exergy of air going to top cycle compressor (kW) 

ExcoldLNG Cold exergy of LNG to be regasified (kW) 

ExLNG Whole exergy of LNG to be regasified (kW) 

ExNG Exergy of exit stream of NG regasified (kW) 

ExNGburned Exergy of NG burned in CC (kW) 

Exergy input in bottom cycle (kW) 

ExQrec 

ExQreg 

Exergy of thermal power required by regasification process (kW) 

LNG Liquefied Natural Gas 

NG Natural Gas 

OR Open Rack units 

P, p Pressure (bar) 

Pel Electric power generated (kW) 

PLNG Pumping power of LNG to be regasified (kW) 

Q Cycle power input (kW) 

Q’LNG Cold power delivered by LNG to be regasified (kW) 

QCU Cold power delivered in CU (kW 

QCU1 Cold power delivered in CU1 (kW) 

QCU2 Cold power delivered in CU2 (kW) 

Power input from NG burned in gas generator CC (kW) 

QNGburned 

Qreg 

Thermal power required by regasification process (kW) 

T, t Temperature (K) 

WNG Pumping power required by NG 

WLNG – Pumping power required by LNG 

Greek symbols 

ε, ζ – exergetic efficiency η – energetic efficiency 

Subscripts 

e – electric bottom – bottom cycle 

t – thermal sat – saturation condition 

top – top cycle


REFERENCES 

*** BP Energy Outlook 2030. January 2011, London, UK, http://www.bp.com. 

*** Eurostat web site: http://europa.eu/documentation/statistics-polls/index_en.htm 

Buffiere J.P., Vincent R, La récupération des frigories du GNL et l’ajustement du gaz 

au terminal de Fos sur Mer. GNL 3, Session III, Paper 8 (1972). 

Cravalho E.G., McGrath J.J., Toscano W.M., Thermodynamic Analysis of the 

Regasification of LNG for the Desalination of Sea Water. Cryogenics (March 

1977). 

Dispenza C., Dispenza G., La Rocca V., Panno G., Ricerca sul risparmio energetico 

nella rigassificazione del GNL, Innovazione tecnologica di impianti energetici: 

Studio teorico e sperimentale di metodologie per la progettazione e la verifica. 

Unità di Ricerca dell’Università di Palermo, 3, DREAM, Università di Palermo, 

Palermo. 2005. 

Dispenza C., Dispenza G., La Rocca V., Panno G., CHP Plants for Production of 

Electrical Energy During Regasification of LNG Recovering Exergy of Cold. 

Proceedings of ASME/ATI 2006 Conference Energy; Production, Distribution 

and Conservation, Vol. II. Milan, May 14-17, 2006, p. 593-603. 

Dispenza C., Dispenza G., La Rocca V., Panno G., Exergy Recovery During LNG 

Regasification: Electric Energy Production (I). Applied Thermal Engineering, 29, 

380-387 (2009 a). 

Dispenza C., Dispenza G., La Rocca V., Panno G., Exergy Recovery in Regasification 

Facilities - Cold Utilization: A Modular Unit. Applied Thermal Engineering, 29, 

3595-3608 (2009 c). 

Dispenza C., Dispenza G., La Rocca V., Panno G., Exergy Recovery During 

LNGregasification: Electric Energy Production (II). Applied Thermal 

Engineering, 29, 388-399 (2009 b). 

Dispenza C., La Rocca V., Pomilla A., Criteri per il dimensionamento di una pipeline 

per il trasporto di etilene allo stato aeriforme. Proceedings of “34° Congresso 

Nazionale ATI”, 8-13 Ottobre, Palermo, 1979. 

Dispenza G., La Rocca V., Panno G., Impianti di produzione combinata di energia 

elettrica e calore di processo integrati in terminali di rigassificazione di gas 

naturale liquefatto. Proceedings “Conference in memory of prof. Salvatore Amyr 

Culotta”, DREAM, Università di Palermo, Palermo, 2002. 

La Rocca V., Cold Recovery During Regasification of LNG. (I) Cold Utilization far 

from the Regasification Facility. Energy, 35, 2049-2058 (2010). 

Snamprogetti, More Energy from LNG, Electric Eergy from LNG Rgasification SP/BBC 

Process 2. Published by Snamprogetti ENI Group, AMSEL, Linate, Milano, Italy, 

1978. 

RECUPERAREA ENERGIEI GENERATE ÎN PROCESUL 

DE VAPORIZARE A GAZELOR NATURALE LICHEFIATE 

(Rezumat) 

Conform prognozelor, consumurile energetice pe glob vor avea tendinţă de 

creştere în viitorul apropiat, datorită creşterii pieţelor de consum. În acest sens se


remarcă grupul BRICS (Brazilia, Rusia, India, China şi Africa de Sud), al ţărilor cu cea 

mai rapidă dezvoltare a pieţelor din lume. Aşadar avem nevoie de tot mai multă energie. 

Pe de altă parte, schimbările climatice generate de poluarea tot mai accentuată a 

mediului înconjurător reprezintă o problemă tot mai apăsătoare, care necesită luarea 

urgentă de măsuri. Marile dezastre ecologice din ultima perioadă, cum ar fi deversarea 

de petrol din Golful Mexic, petrecută în anul 2010, sau cutremurul urmat de ţunami, 

care a avut loc în Japonia în 2011, au marcat profund opinia publică şi au făcut ca multe 

state, cum ar fi cele europene, să-şi reevalueze profund programele nucleare sau chiar să 

le stopeze. 

Aşadar se impune utilizarea cât mai raţională a tuturor resurselor energetice. 

Indubitabil, gazul natural reprezintă una dintre cele mai utilizate astfel de resurse, motiv 

pentru care s-au realizat numeroase reţele de transport a acestuia. Pe lângă reţelele deja 

realizate, există multe altele aflate în fază de construcţie sau de proiectare. Totodată, 

există ţări consumatoare de gaze naturale care se află la distanţă foarte mare de ţările 

furnizoare. Un exemplu în acest sens este Japonia, ţară care, in prezent, îşi reevaluează 

politica energetică, principala resursa energetică promovată fiind gazul natural. În astfel 

de cazuri, singura posibilitate de furnizare a gazelor naturale este cu ajutorul 

transportului naval. Varianta cea mai bună este lichefierea lor prealabilă deoarece starea 

lichidă permite transportarea masei maxime de gaze naturale într-un spaţiu cu volum 

dat, aşa cum este cazul rezervoarelor de combustibil. 

De la locul de extracţie, gazele naturale sunt transportate prin conducte până la 

terminalul de îmbarcare, amplasat într-un port. Acest terminal este de fapt o unitate 

industrială care are ca scop lichefierea gazelor naturale. Procesul de lichefiere 

presupune consumuri energetice foarte mari care sunt acoperite utilizând chiar o parte 

din gazele ce urmează a fi transportate. 

La terminalul de sosire are loc re-gazeificarea (vaporizarea) LNG şi, ulterior, 

introducerea gazelor naturale obţinute în reţelele existente. Procesul de re-gazeificare 

presupune, de asemenea, un consum foarte mare de energie şi generează, la rândul lui, o 

cantitate foarte mare de frig. Recuperarea frigului reprezintă o problemă de interes 

major, atât din punct de vedere vedere energetic cât mai ales ecologic: în cazul în care 

nu ar avea loc recuperarea, energia respectivă ar fi risipită, fiind disipată în mare, fapt ce 

ar avea un important impact ecologic asupra mării, în zona terminalului. 

Pentru recuperarea frigului pot fi luate în considerare atât varianta producerii de 

energie electrică într-o instalaţie cogenerativă, care să utilizeze sectorul criogenic al 

terminalului de re-gazeificare ca sursă rece, cât şi utilizarea directă a frigului în procese 

în care acesta este necesar. 

După o privire de ansamblu asupra pieţei gazelor naturale, în lucrare sunt 

discutate câteva posibile variante de utilizare a frigului rezidual rezultat din procesul de 

re-gazeificare a LNG. Este analizată o instalaţie de producere a energiei electrice şi este 

realizat un studiu de caz privind utilizarea directă a frigului.


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DESIGN AND TESTING OF A 50 kW DOWNDRAFT GASIFIER 

BY 

DANIEL DRAGOMIR-STANCIU ∗ and MARIAN STANCU 

1 “Gheorge Asachi” Technical University of Iaşi, 

Department of Mechanical and Automotive Engineering 

Received: November 22, 2012 

Accepted for publication: November 29, 2012 

Abstract. The paper presents design, manufacture and testing of a 

downdraft gasifier with a thermal output of 50 kW. The constructive solutions 

adopted, the main dimensions and the test results are presented. 

Key words: downdraft gasifier, design, testing. 

1. General Considerations 

This gasifier is intended to be a part of on microcogeneration unit with 

electrical power of 10 kW. The wood gas generator will be designed for 

operation in conjunction with a piston engine. Producer gas, the gas generated 

when wood is gasified with air, is the fuel for a piston engine driving an 

electrical generator. 

Was chosen a downdraught gasifier, in which primary gasification air is 

introduced at or above the oxidation zone. The producer gas is removed at the 

bottom of the apparatus, so that fuel and gas move in the same direction. The 

advantages of downdraft gasiefier are the flexible adaptation of gas production 

to load and low sensitivity to charcoal dust and tar content of fuel. The open 

top permits fuel to be fed more easly and allows easy access. 

Four distinct processes take place in a downdraft gasifier as the fuel 

makes its way to gasification. They are: drying of fuel, pyrolysis – a process in 

which tar and other volatiles are driven off, combustion, reduction. 

∗ Corresponding author: e-mail: ddragomir03@yahoo.com

30 Daniel Dragomir-Stanciu and Marian Stancu 

Fig. 1 shows schematically these regions for downdraft gasifier. 

Air 

g 

g s 

Pyrolysis 

Combustion 

Reduction 

Biomass 

Fig. 1 – Processes in the downdraft gasifier. 

System should contain a cyclone for separating solid particles and a filter 

for retaining tar. 

2. Design of the Downdraft Gasiefier 

2.1. Design Guidelines for Downdraft Gasiefiers 

One of the most important parameters for the downdraft gasifiers is so 

call “hearth load”. 

The hearth load Bg is defined as the amount of producer gas reduced to 

normal (p, T) conditions, divided by the surface area of the “throat” at the 

smallest circumference and is usually expressed in m³/cm²/h. Alternatively the 

hearth load can be expressed as the amount of dry fuel consumed, divided by 

the surface area of the narrowest constriction (Bs), in which case hearth load is 

expressed in kg/cm²/h. 

Because one kilogramme of dry fuel under normal circumstances 

produces about 2.5 m³ of producer gas, the relation between Bg and Bs is given 

by: 

Air 

B = 2.5 B . 

(1) 

In the case of a “single” throat design, hearth diameter at air inlet height 

should be 100 mm larger than “throat” diameter. Height of the reduction zone 

should be more than 200 mm. Height of the air inlet nozzle plane should be 100 

mm above the constriction. Throat inclination should be around 45°…60°. 

Gas


2.2. Main Dimensions and Construction of the Gasifier 

From preliminary calculations is known the wood consumption of the 

gasifier 15.1 kg/h, for a heating value of generated gas Hi = 1503.5 kcal/ Nm 3 . 

This gasifier uses 1.5 kg biomass to produce 1 kWh electricity. 

Diameter of throat at smallest cross-sectional area dt =70 mm ; 

Once the throat diameter has been fixed, further important gasifier 

dimensions can be determined: surface area “throat” St = 0.003846 m 2 , hearth 

load Bg = 0.742, height h of the nozzle plane above the smallest cross-section 

of the throat h = 100 mm. 

Assumption gasifier to be equipped with 5 nozzles. The necessary air 

massflow is 15.68 Nm 3 /h. 

i) nozzle diameter dn = 8 mm 

ii) ratio between nozzle flow area, Sn, and throat area, St , Sn/St = 0.072 

iii) nozzle air outlet velocity w = 15.4 m/s. 

Arrangement of the nozzles is shown in Fig. 2. 

72 

Φ 260 

Φ 70 Φ 170 

Fig. 2 – Arrangement of the nozzles. 

The gas generator is manufacturated by welding, using carbon steel S235 

JR, except for the heart of reactor using a steel alloy W1 4828 for high 

temperature. The gas generator is represented in Fig. 3. 

For optimal functioning of gas generator is neccesary a preheating of the 

air. For air preheating was adopted an innovative solution: air flows to the 

nozzles through five semicircular pipes, welded on the outside surface of the 

reactor. These pipes can be seen well in the photo in Fig. 4. 

4. Testing and Results 

It was performed the first series of tests, using as fuels fir, beech and 

wood pellets. 

Before entering in the burner, gas was mixed with air in an ejector. 

The following parameters were measured: 

i) temperature in combustion zone ( smallest section): 900...1200°C; 

ii) temperature in bottom reduction zone: 700...800 °C; 

iii) pressure after reduction zone: 1.015...1.02 bar; 

iv) temperature of the gas after filter: 100…120 °C .

32 Daniel Dragomir-Stanciu and Marian Stancu 

Fig. 3 – The gas generator and the cyclone. 

Fig. 4 – The reactor.



1. The pressure and temperature values indicate a good functioning of gas 

generator. 

2. Flame shape and color indicate a high calorific value of the gas. 

3. The presence of the tar require better filtration and gas cleaning. 

4. Experiments will continue, and will accurately determine the gas flow, 

gas composition and calorific value. 

5. The next phase of testing will be using of the gas as combustible for a 

piston engine and engine power measurement. 

REFERENCES 

Mathieu P., Dubuisson R., Performance Analysis of a Biomass Gasifier. Energy, 

Conversion and Management, 43, 8, 1291-1299 (2002). 

Mukunda H.S., Shrinivasa D.U., Open Top Wood Gasifiers. Renewable Energy, 

Sources for Fuels and Electricity. Island Press, 1993, pp. 699-728. 

Rajvanshi A.K., Biomass Gasification. In Alternative Energy in Agriculture. Vol. II, 

Ch. 4, Ed. D. Yogi Goswami, CRC Press, 1986, pp. 83-102. 

Reed T.B., Das A., Handbook of Biomass Downdraft Gasifier Engine Systems. SERI/SP 

271-3022, Solar Research Institute, Golden CO, 1988. 

Schmidt D., Martin K., Biomass Gasifier Power System Feasibility; Final Report for 

Building Materials Holding Corporation; EERC Publication 2005-EERC-10-02; 

Energy & Environmental Research Center, Grand Forks, ND, 2005. 

Wandera P.R., Altafinib C.R., Barretob R., Assessment of a Small Sawdust Gasification 

Unit. Biomass and Bioenergy, 27, 467-476 (2004). 

PROIECTAREA ŞI TESTAREA UNUI GAZEIFICATOR 

CU CIRCULAŢIE DESCENDENTĂ, CU PUTEREA 50 kW 

(Rezumat) 

Lucrarea prezintă proiectarea, realizarea şi testarea unui gazeificator cu puterea 

de 50 kW pentru o instalatie de microcogenerare. Gazeificatorul este cu circulaţie 

descendentă, biomasa utilizată fiind lemnul. Au fost efectuate calculele pentru 

dimensionarea gazeificatorului şi s-a stabilit soluţia constructivă. Pentru preîncălzirea 

aerului s-a adoptat o soluţie nouă, aerul fiind trimis spre zona de combustie prin canale 

semicirculare sudate pe corpul reactorului. Primele încercări şi teste au indicat o bună 

funcţionare a gazeificatorului. Experimentele vor continua, urmând să fie găsite soluţii 

pentru o reducere mai pronunţată a gudronului, iar în continuare se vor efectua teste 

pentru combustia gazului într-un motor cu piston.


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THEORETICAL ECO-EFFICIENCY COMPARATIVE STUDY 

CASE, HYDROCARBONS, AMMONIA AND HFC MIXTURE 

ALTERNATIVES RETROFIT 

BY 

GRAŢIELA MARIA ŢÂRLEA ∗2 , ION ZABET 1 , MIOARA VINCERIUC 2 

and ANA ŢÂRLEA 2 

1 Romanian General Association of Refrigeration of Bucharest 

2 Technical University of Civil Engineering of Bucharest 

Received: September 20, 2012 


Abstract. Several natural refrigerant alternatives are compared on the basis 

of their cycle coefficient of performance and TEWI factor (Total Equivalent 

Warming Impact – in respect with SR EN 378-1). 

These natural alternatives: ammonia and mixtures have zero or low global 

warming potential (GWP). 

The theoretical study uses as reference a single stage refrigeration system 

which works with R 404A. To implement the international Legislation, in the 

future it is necessary to retrofit HFC refrigerants with ecological refrigerants 

(with zero ODP and zero GWP). 

Energy efficiency is directly related to global warming and greenhouse 

gases emissions 

Key words: energy efficiency, TEWI, natural refrigerants, refrigeration 

system. 


The retrofitted single stage refrigeration system is used for chilled 

products in a cold storage (0…4°C). Initial refrigeration system is composed by: 

condenser, scroll compressors, evaporators, refrigerant vessel and thermostatic 

valves. 

∗ Corresponding author: e-mail: gratiela.tarlea@gmail.com

36 Graţiela Maria Ţârlea et al. 

The single stage refrigeration system was upgraded by replacing the 

scroll compressor (working with R404A) with open drive screw compressor 

(working with R717). The new refrigeration system is composed by: condenser, 

open drive screw compressor, ammonia vessel, ammonia pump, cooling oil 

vessel, and equalization column and ammonia evaporators. 

In the following lines the eco-efficiency comparative study between three 

refrigerants (R717, R404A and R507A) are described. The eco-efficiency 

comparative study consists in performance coefficient, annual energy 

consumption and TEWI factor (Ţârlea et al., 2010). 

2. Study Case 

In this section, the method and step calculation for eco-efficiency of 

single stage refrigeration system will be described. Fig. 1 shows the schematic 

representation of the single stage refrigeration system which must be retrofitted 

from R404A to R717. The R404A refrigerant from the discharge compressor 

port goes to the condenser (where appear the condensation process), then from 

the condenser outlet it goes into liquid receiver. From the liquid receiver the 

refrigerant enter in evaporator 1 and 2 (where appear evaporation process) and 

after the evaporator the refrigerant goes to suction port of the compressor. 

First the working conditions are described; secondly the theoretical ecoefficiency 

parameters (presented in the introduction) and thirdly the results of 

the theoretical study are shown. 

The working conditions are given in Table 1. The refrigeration system 

works 8 hours/day (nhours) and 261 days/year (ndays). The life cycle of the 

refrigeration system is 15 years (nyears). The refrigerant charge (m) is: R404A – 

14kg, R717 – 6.45 kg and R507A – 14.11 kg. The requirement electrical power 

(Pe) is shown in Table 1. The carbon dioxide emission (β) is 0,6kg/kWh. The 

recovery factor (αrec) is 75%. The refrigerant leak (msc) is 8%. 

The electrical power values were calculated and established by each 

refrigeration system, depending on properties and evaporation temperature 

(Zabet, 2011). 

Refrigerant Φ0 

kW 

M 

kg 

Table 1 

Working conditions 

Tc 

K 

TSH 

K 

TSUB K T0 

K 

263.2 273.2 278.2 

Pe , kW 

R404A 12.3 14 3.70 5.25 5.47 

R717 10 6.45 308.2 10 3 3.55 4.00 3.85 

R507A 10 14.11 

3.80 3.94 4.07 

where: Φ0 is refrigerant capacity, T0 evaporation temperature; TC condensing 

temperature, TSH superheating temperature; TSUB subcooling.


Fig. 1 – Schematic representation of refrigeration system: where: T1 – refrigerant 

temperature at suction compressor, ºC; T2 – refrigerant temperature at discharge 

compressor, ºC; T3 refrigerant temperature at condenser inlet, ºC; T4 – refrigerant 

temperature at condenser outlet, ºC; T5 – air temperature at inlet evaporator 1, ºC; T6 – 

air temperature at outlet evaporator 1, ºC; T7 – refrigerant temperature at discharge of 

evaporator 1, ºC; T8 – air temperature at inlet evaporator 2, ºC; T9 – air temperature at 

outlet evaporator 2, ºC; T10 – refrigerant temperature at discharge of evaporator 2, ºC; 

T11 – air temperature at condenser inlet, ºC; T12 – air temperature at condenser 

outlet, ºC. 

2.1. Calculation of the Theoretical Eco-efficiency Parameters 

The theoretical eco-efficiency parameters discussed in this paper are: the 

performance coefficient, annual energy consumption and TEWI factor. For 

finding the eco-efficiency parameters, classic thermodynamics relations 

together with parameters defined in Table 1 were used (Ţârlea et al., 2010). 

i) Finding the performance coefficient 

First of all was calculated the refrigerant mass flow by 

0 M& 

Φ 

= , 

Δh 

(1) 

ev


where Δhev 

is enthalpy difference at evaporation working condition, kJ/kg; M is 

the refrigerant mass flow rate, kg/s. 

Secondly, the mechanical power for the compressor was found, by 

(2) 

, 

W& = M& Δh 

cp cp 

where: Wcp is the mechanical power, kW; 

& 

compression working conditions, kJ/kg. 

Δ hcp 

is the enthalpy difference at 

Finally, the results of the coefficient of performance (COP) by 

Φ0 

COP = . 

W& 

ii) Finding the annual energy consumption 

The annual energy consumption (Eannual) is calculated according SR EN 

378-1 by 

E = n n P 

(4) 

iii) Finding the TEWI factor 

cp 

annual hours days e. 

The TEWI factor is calculated by SR EN 378-1 by 

( ) 

TEWI = GWP ⋅ Ln + GWP ⋅m1− α + n E β (5) 

years rec years annual 

where: GWP – global warming potential [-]; L – refrigerant leak during a year 

working, kg 

(3) 

L= m m. 

(6) 

sc 

Table 2 shows the global warming and ozone depletion potential values 

for analysed refrigerants (Ţârlea et al., 2011). 

Table 2 

Global warming and ozone depletion potential values 

Refrigerant GWP ODP 

R404A 3260 0 

R717 0 0 

R507A 3300 0 

3. Theoretical Eco-Efficiency Study Results 

The theoretical results are presented in the next lines by tables and 

figures. Table 3 shows the study results for the coefficient of performance, 

annual energy consumption and TEWI factor (Zabet & Ţârlea, 2011). 

As it is shown in Fig. 2, the electrical power for R717 refrigerant is 15% 

smaller than R404A and R507A and this electrical power savings has a great 

impact over the environment pollution and maintenance cost.


Table 3 

The study results 

Refrigerant COP [-] E annual , kWh TEWI, tonsCO 2 

R404A 4.484 6.328 7.724 10962 8352 8227 164.8 141.3 140.2 

R717 4.812 6.689 8.101 7726 7412 7934 69.5 66.7 71.4 

R507A 4.461 6.273 7.639 11421 8039 8498 170.3 139.9 144 

Fig. 2 – Electrical power versus evaporation temperature. 

Fig. 3 shows the variation of the coefficient of performance’s versus 

evaporation temperature. One could observe that the coefficient of performance 

for R717 refrigerant is higher by 6% than R404A and by 7% than R507A 

(Ţârlea et al., 2010). This is happens due to convenient thermodynamics 

properties of R717. 

Fig. 3 – Coefficient of performance versus evaporation temperature. 

Fig. 4 presents the variation of TEWI factor versus evaporation 

temperature. As shown, the TEWI factor for R717 refrigerant is 60% smaller 

than R404A and R507A. This implies that R717 refrigerant is the most suitable 

for environment from the global warming point of view.


Fig. 4 – TEWI factor versus evaporation temperature. 

Fig. 5 shows the annual energy consumption versus the evaporation 

temperature. It can be observed that the annual energy consumption for R717 

refrigerant is 15% smaller than R404A and R507A. 

Fig. 5 – Annual energy consumption versus evaporation temperature. 


1. The theoretical study analysed a single stage refrigeration system with 

R 404A as refrigerant. 

2. To improve the eco-efficiency, the HFC refrigerant was replaced with 

an ecological refrigerant R717. 

3. The comparative study of these facilities was based on the coefficient 

of performance of a refrigeration system and the TEWI factor.


4. The refrigerant R717 is more eco-efficient than R404A and R507A 

with a 15% lower electrical power consumption, 6% higher coefficient of 

performance, 60% lower TEWI factor and 15% lower annual energy 

consumption. 

5. Retrofitting the single stage refrigeration system by replacing R404A 

refrigerant with R717 it is proven that the savings are higher from the ecoefficiency 

point of view and the costs are lower (more equipment in the case of 

R717 than R404A). 

REFERENCES 

Ţărlea G.M., Vinceriuc M., Zabet I., Ţărlea A., Ecological Alternative for R404A 

Refrigerant. The 41st HVAC&R Congress KGH 2010, Tome “Heating, 

Refrigerating and Air-Conditioning”, 3-5 December, Belgrad, Serbia, 2010, pp. 

27-33. 

Ţărlea G.M.,Vinceriuc M., Zabet I., Ţărlea A., Theoretically Study of Ecological 

Alternative for R404A, R507A and R22. The 42nd International Congress and 

Exhibition Heating, Refrigeration and Air-Conditioning November 30 – December 

2, Belgrade, 2011. 

Ţărlea G.M., Vinceriuc M., Zabet I., Ammonia as a Very Eco-efficient Romanian 

Refrigerant. International Congress, 5-7 May COFRET, Iaşi, 2010. 

Zabet I., Contribution Regarding the Study of the Eco-efficiency in Refrigeration 

Systems. PhD Thesis, 2011. 

STUDIU TEORETIC COMPARATIV PRIVIND ÎMBUNĂTĂŢIREA DIN PUNCT 

DE VEDERE AL ECO-EFICIENŢEI UNUI SISTEM FRIGORIFIC CA URMARE A 

ÎNLOCUIRII HIDROCARBURILOR ŞI AMESTECURILOR DE HFC CU 

AMONIAC 

(Rezumat) 

Criza energetică majoră precum şi încălzirea globală care afectează în prezent 

economia mondială şi viitorul societăţii umane, impun creşterea performanţelor 

energetice şi ecologice ale echipamentelor şi sistemelor frigorifice şi de aer condiţionat. 

În acest scop pe plan mondial există un efort susţinut pentru reducerea emisiilor de 

dioxid de carbon rezultate din arderea combustibililor fosili şi a altor emisii de gaze cu 

efect de seră. 

Se prezintă un sistem frigorific într-o treaptă de comprimare cu vapori 

funcţionând cu agentul frigorific R404A, instalaţie care a fost optimizata prin înlocuirea 

cu agentul frigorific R717. 

Structura aleasa la redactarea lucrării prezente este următoarea: în prima parte s-a 

făcut o scurta descriere a condiţiilor de lucru, în partea a doua s-au calculat parametrii 

teoretici de eco-eficienţă (TEWI, COP, puterea electrică consumată şi consumul anual 

de energie electrică) şi în a treia parte s-au prezentat rezultatele obţinute în urma 

studiului teoretic.


Conform celor arătate în Fig. 2 consumul electric al sistemului frigorific 

funcţionând cu agentul frigorific R717 este cu 15% mai mic decât în cazul utilizării 

agentului frigorific R404A si R507A. Aceasta reducere aduce un câştig semnificativ din 

punct de vedere al costurilor precum şi al impactului ecologic asupra mediului 

înconjurător. 

Conform celor prezentate în Fig. 3 se poate observa un coeficient de performanţă 

mai mare în cazul agentului frigorific R717: cu 6% faţă de R404A şi cu 7% faţă de 

R507A. 

În concluzie se poate afirma că optimizarea sistemului frigorific cu vapori într-o 

treaptă de comprimare prin înlocuirea agentului frigorific R404A cu agentul frigorific 

R717 aduce un câştig din punct de vedere al eficienţei energetice, costurilor şi un 

avantaj din punct de vedere al mediului înconjurător prin reducerea gradului de poluare.


Publicat de 



Secţia 


INFLUENCE OF ENVIRONMENTAL AND CONSTRUCTIVE 

FACTORS ON THE VAPORIZERS PERFORMANCE 



BY 

NICOLAE BARA 1 and MARIN BICA ∗2 

1 Frigotehnica SA Bucharest 

2 University of Craiova 

Abstract. Correct design and choice of vaporizers are very important for 

refrigeration systems proper operation and for their efficiency. A wrong sized 

vaporizer may produce an excessive decrease of the vaporization temperature, 

and reducing it with every degree corresponds to a reduction in cooling power 

with approx. 3…4%. This is why the vaporizer can not be dissociated from its 

liquid supply system. In practice, often a proper rebound refrigerant corresponds 

to each type of vaporizer. The measurements for vaporizers with the rib step 

between 2 mm and 3.5 mm were performed on the Thermotechnics Laboratory 

stand from the Faculty of Mechanics of Craiova. The most pronounced air 

cooling occurs for the rib step value equal to 2 mm. So, it is confirmed the 

statement that the rib step has to be chosen for each type of heat exchanger, after 

carefully testing on laboratory stands. 

Key words: vaporizer, refrigeration system, rib step. 


Vaporization is the thermodynamic process through which the refrigerant 

changes its state of aggregation from liquid to vapors, absorbing heat from the 

cold source represented by the cooled environment. In any refrigeration 

machine the vaporizer is the unit absorbing heat, achieving the machine useful 

effect. 

∗ Corresponding author: e-mail: marinbica52@gmail.com

44 Nicolae Bara and Marin Bica 

There are many types of vaporizers, depending on their destination: vaporizers 

for air cooling; vaporizers for liquids cooling. 

Correct design and choice of vaporizers are very important for 

refrigeration systems proper operation and for their efficiency. A wrong sized 

vaporizer may produce an excessive decrease of the vaporization temperature, 

and reducing it with every degree corresponds to a reduction in cooling power 

with approx. 3…4%. This is why the vaporizer can not be dissociated from its 

liquid supply system. In practice, often a proper rebound refrigerant corresponds 

to each type of vaporizer. 

2. Experimental Research 

For the chosen vaporizer represented in Fig. 1, measurements were made 

on the air flow at different velocities. The results are presented in Table 1. In 

this case the air flow velocity variation was performed with the help of the car 

speed variator. In this case the air flow velocity variation was performed with 

the help of the car speed variator. 

Fig. 1 – The air conditioner vaporizer from a baby-car. 

Table 1 

Variation of the cooled air temperature depending on the flow (tair: 28.4 ° C and 22 o C) 

Nr. mair [kg/s] t 2 

′ [°C] t 2 

′ [°C] Number of plates Channel width 

1 0.081 28.4 21.1 14 0.002 

2 0.110 28.4 18.2 14 0.002 

3 0.127 28.4 17.1 14 0.002 

4 0.142 28.4 15.4 14 0.002 

5 0.089 22 18 14 0.002 

6 0.117 22 16.2 14 0.002 

7 0.132 22 14.3 14 0.002 

8 0.147 22 12.8 14 0.002 

where t 2 – ambient temperature, 

′ 

t 2 

′ – air temperature at vaporizer outlet. 

The graph shown in Fig. 2 highlights the known decreasing trend of the 

temperature at vaporizer outlet, when air velocity (air flow) increases. There are 

confirmed the conclusions from the specialty literature that the flow becomes 

turbulent when velocity increases, causing an intensification of heat transfer on


the air. Taking into account the large differences of values between αair and 

αlf coefficients and the lowest values obtained on the air flow, research has to be 

focused on increasing their values in order to increase the overall heat transfer 

coefficient. The research performed on the vaporizer turns out that an air flow 

velocity increasing is a solution up to a certain value and then the increasing 

slope of the heat exchange is negligible. 

Fig. 2 – Cooled air temperature variation depending on flow. 

The two curves were built for different ambient conditions, respectively 

28.4 and 22 o C. Measurements to determine the air flow were performed using a 

Kimo-type thermo-anemometer provided by ICMET Craiova. There were 

measured air temperatures at the vaporizer entry and exit, and the air flow


velocity. There were calculated the volume and mass air flow, and the air 

density at average temperature was determine using the tables. The obtained 

values are listed in Table 2. 

Nr. 

S, 

m 2 

ρair 

kg/m 3 

Table 2 

Values measured to determine the air flow 

t′ air 

o 

C 

t′′ air , 

o 

C 

tair, m 

o C 

wair 

m/s 

Vair 

m 3 /s 

mair 

kg/s 

1 0.0279 1.205 22 18 20 2.64 0.07385 0.089 

2 0.0279 1.208 22 16.2 19.1 3.47 0.09685 0.117 

3 0.0279 1.212 22 14.3 18.15 3.90 0.10891 0.132 

4 0.0279 1.215 22 12.8 17.4 4.34 0.12128 0.147 

The parameters obtained for the step of 0.002 were measured in order to 

highlight the influence of the rib step width dimensions, and then they were 

calculated for other values. The measurements and calculations results were 

listed in Table 3. 

No. 

Table 3 

Values determined by calculus for different dimensions of the air flow channel 

Channel 

width 

t′ air 

°C 

t′′ air 

°C 

t air. m 

o C 

ρ air 

kg/m 3 

w air 

m/s 

m air 

kg/s 

1. 0.001 28.4 22.3 25.35 1.183 2.68 0.0792 0.025 

2. 0.001 28.4 20.2 24.3 1.188 3.86 0.1146 0.025 

3. 0.001 28.4 18.1 23.25 1.192 4.22 0.1257 0.025 

4. 0.001 28.4 16.8 22.6 1.194 4.84 0.1444 0.025 

5. 0.001 28.4 16.2 22.3 1.195 5.02 0.1499 0.025 

6. 0.001 28.4 15.8 22.1 1.196 5.18 0.1548 0.025 

7. 0.001 28.4 15.2 21.8 1.197 5.45 0.1630 0.025 

8. 0.001 28.4 14.8 21.6 1.1986 5.89 0.1764 0.025 

9. 0.0015 28.4 20.1 24.25 1.188 2.68 0.0865 0.0264 

10. 0.0015 28.4 19.2 23.8 1.189 3.86 0.124 0.0264 

11. 0.0015 28.4 18.4 23.4 1.191 4.22 0.1364 0.0264 

12. 0.0015 28.4 17.6 23.00 1.193 4.84 0.1568 0.0264 

13. 0.0015 28.4 17.1 22.75 1.194 5.02 0.162 0.0264 

14. 0.0015 28.4 16.4 22.4 1.1954 5.18 0.185 0.0264 

15. 0.0015 28.4 15.6 22.00 1.197 5.45 0.190 0.0264 

16. 0.0015 28.4 14.2 21.3 1.1998 5.89 0.1944 0.0264 

17. 0.002 28.4 19.2 23.8 1.189 2.68 0.0866 0.0272 

S 

m 2


18. 0.002 28.4 17.6 23.00 1.193 3.86 0.1252 0.0272 

19. 0.002 28.4 15.8 22.10 1.196 4.22 0.1372 0.0272 

20. 0.002 28.4 14.4 21.40 1.199 4.84 0.578 0.0272 

21. 0.002 28.4 14.00 21.2 1.2002 5.02 1.6388 0.0272 

22. 0.002 28.4 13.8 21.1 1.2006 5.18 1.1691 0.0272 

23. 0.002 28.4 13.4 20.9 1.2024 5.45 1.1782 0.0272 

24. 0.002 28.4 13.00 20.7 1.2026 5.89 1.1926 0.0272 

25. 0.0025 28.4 20.4 24.40 1.187 2.68 0.0881 0.0277 

26. 0.0025 28.4 18.8 23.60 1.190 3.86 0.1272 0.0277 

27. 0.0025 28.4 17.1 22.75 1.194 4.22 0.1395 0.0277 

28. 0.0025 28.4 15.8 22.10 1.196 4.84 0.1603 0.0277 

29. 0.0025 28.4 15.3 21.85 1.1976 5.02 0.1665 0.0277 

30. 0.0025 28.4 14.9 21.65 1.1984 5.18 0.1719 0.0277 

31. 0.0025 28.4 14.2 21.30 1.1998 5.45 0.1811 0.0277 

32. 0.0025 28.4 13.9 21.15 1.2004 5.89 0.1958 0.0277 

33. 0.003 28.4 21.1 24.75 1.186 2.68 0.0893 0.0281 

34. 0.003 28.4 18.2 23.3 1.1918 3.86 0.1292 0.0281 

35. 0.003 28.4 17.1 22.75 1.194 4.22 0.1415 0.0281 

36. 0.003 28.4 15.4 21.9 1.1974 4.84 0.1628 0.0281 

37. 0.003 28.4 15.2 21.8 1.1978 5.02 0.1689 0.0281 

38. 0.003 28.4 14.9 21.65 1.1984 5.18 0.1744 0.0281 

39. 0.003 28.4 14.4 21.4 1.1994 5.45 0.1836 0.0281 

40. 0.003 28.4 14.00 21.2 1.2002 5.89 0.1986 0.0281 

41. 0.0035 28.4 21.2 24.8 1.185 2.68 0.0898 0.0283 

42. 0.0035 28.4 19.1 23.75 1.19 3.86 0.1299 0.0283 

43. 0.0035 28.4 18.4 23.4 1.191 4.22 0.1422 0.0283 

44. 0.0035 28.4 16.2 22.3 1.195 4.84 0.1636 0.0283 

45. 0.0035 28.4 15.8 22.1 1.1966 5.02 0.1699 0.0283 

46. 0.0035 28.4 15.3 21.85 1.1976 5.18 0.1755 0.0283 

47. 0.0035 28.4 14.9 21.65 1.1985 5.45 0.1848 0.0283 

48. 0.0035 28.4 14.00 21.2 1.2002 5.89 0.2000 0.0283 

The measurements for a vaporizer with 2 mm rib step were performed on 

the Thermotechnics Laboratory stand from the Faculty of Mechanics of 

Craiova. Thermal efficiency was calculated for ribs having the step of 1 mm, 

respectively 3.5 mm, using the thermal model adopted by the specialty 

literature. The most pronounced air cooling occurs for the 2 mm step. 

So, it is confirmed the statement that the rib step has to be chosen for 

each type of heat exchanger, after carefully testing on laboratory stands. 

Graphical results are presented in Fig. 3.


Fig. 3 –Cooling efficiency depending on the rib step. 


1. For surfaces with sinusoidal ribs, specialty literature data regarding the 

behavior of these heat transfer surfaces is poor. Some recent data are obtained 

on the computer and it substantially deviate from the values obtained by 

experimental research, showing that theoretical models are still improvable. 

2. Theories about the entry effect and the effect of the boundary layer 

release from the sinusoidal rib demonstrated by complex experimental research 

open opportunities for thermally superior devices in relation to existing devices. 

These results have led to research for the development of exchangers with the 

two fluids in counter current flow. 

3. Research on new constructive solutions obtaining for heat exchangers 

in refrigeration systems has highlighted the importance of ribs surface quality 

and of the step got in their fabrication. 

4. Obtaining much better thermal efficiency leads to vehicles total mass 

decreasing and fuel consumption reducing with direct implications in limiting 

environmental pollution.


REFERENCES 

* * * Thermal Management for Heavy Vehicles (Class 7 -8 Trucks). Workshop Report 

and Multiyear Program Plan (Draft), U.S. Department of Energy, Office of Heavy 

Vehicles Technologies, 2000. 

Bică M., Nagi M., Transfer de masă şi căldură. Universitaria, Craiova, 1999. 

Chiriac F., Chiriac V., Cristea A., An Alternative Method for the Cooling of Power 

Microelectronics Using Classical Refrigeration. Thermal Issues in Emerging 

Technologies, ThETA 1, Cairo, Egypt, Jan 3-6, 2007. 

Frank P., Incropera F., DeWitt D., Fundamentals of Heat and Mass Transfer. Wiley & 

Sons, 2002. 

Gray A., The Growth of Aluminum in Automotive Heat Exchangers. Innoval 

Technology Limited, Banbury, OX16 1TQ, UK, 2005. 

Natarajan V., Convective Heat Transfer from a Stacked Electronic Package. ThETA 1, 

Cairo, Egypt, 2007. 

INFLUENŢA FACTORILOR DE MEDIU ŞI CONSTRUCTIVI ASUPRA 

PERFORMANŢELOR VAPORIZATOARELOR 

(Rezumat) 

Proiectarea şi alegerea corectă a vaporizatoarelor are o importanţă mare pentru 

funcţionarea corectă a instalaţiilor frigorifice şi pentru eficenţa acestora. Un vaporizator 

greşit dimensionat poate să producă o scădere excesivă a temperaturii de vaporizare, iar 

la reducerea acesteia cu fiecare grad, corespunde şi o reducere a puterii frigorifice cu 

cca. 3…4%. Acesta este şi motivul pentru care nu se poate disocia vaporizatorul de 

sistemul său de alimentare cu lichid. În practică, adesea fiecărui tip de vaporizator îi 

corespunde un sistem propriu de destindere a agentului frigorific. Pe standul din 

Laboratorul de Termotehnică de la Facultatea de Mecanică din Craiova s-au efectuat 

măsurători pentru un vaporizatoare cu pasul nervurii cuprins între 2 mm şi 3,5 mm. 

Răcirea cea mai pronunţată a aerului se produce la pasul nervurii de 2 mm. Se confirmă 

afirmaţia că pasul nervurii trebuie ales pentru fiecare tip de schimbător căldură în parte, 

în urma unor încercări minuţioase pe standurile de probe.


Publicat de 



Secţia 


AN IMPROVED METHOD FOR HEAT PUMPS 

REFRIGERANT CHOICE 

BY 

IONEL OPREA ∗ 

"Dunărea de Jos" University of Galaţi 

Department of Thermal Systems and Environmental Engineering 

Received: October 28, 2012 


Abstract. This paper summarizes the impact of thermo-physical properties 

on refrigerant selection for HP: R600, R404a, R407c,R410a, R134a, R507, 

R134a and R717, which have zero ODP. Impact factors are not sufficient for a 

complete evaluation, whereas the multicriteria approach allows for an effective 

initial assessment of the refrigerant. Additional parameters such as efficiency and 

safety issues will be included in the detailed analysis. 

Key words: coefficient of performance; heat of vaporization; heat of 

condensation; condensing pressure; evaporation pressure; compressor work; 

volume flow. 


The choice of a Refrigerant implies compromises between conflicting 

desirable thermophysical properties. A refrigerant must satisfy many 

requirements, some of which do not directly relate to its ability to transfer heat. 

Chemical stability under the conditions of use is an essential requirement. 

Safety codes may demand for a nonflammable refrigerant of low toxicity for 

some applications. The environmental impact of refrigerant leaks must also be 

considered. Cost, availability, efficiency, and compatibility with compressor 

lubricants and equipment materials are other concerns. 

Transport properties (e.g., thermal conductivity and viscosity) affect the 

performance of heat exchangers and piping system. High thermal conductivity 

∗ e-mail: ioprea@ugal.ro

52 Ionel Oprea 

and low viscosity are desirable. No single fluid satisfies all the attributes desired 

of a refrigerant; consequently, various refrigerants are used. 

Refrigerant 

Table 1. 

Refrigerant Atmospheric Lifetime, ODP and GWP values 

Atmospheric 

Lifetime, 

years a 

ODP b GWP c 100 

R-600 0.018 d 0 ~20 d 

R-717 0.01 d 0


system efficiencies should be expected to be equal to, or higher than, those of 

R404a, R717, and R134a systems. The risks represented by the flammability of 

the hydrocarbons must be seriously taken into account. The risks can be reduced 

by designing the systems for a minimum charge of refrigerant, careful leak 

detection during production, hermetic design with a minimum number of 

connections, the use of spark-proof electric components and ventilation of 

confined spaces. 

Fig. 1 – Representation of refrigerant cycle for a zeotropic refrigerant. 

The following matters are taken into account in this multicriteria analysis: 

1.COP - coefficient of performance: refrigerants with a high critical 

temperature give the best COP; the natural refrigerants are superior in this 

respect. 

2. qk: in HP cycles involving condensation, a refrigerant must be chosen 

so that this change of state will occur at a temperature somewhat below the 

critical temperature; 

3. q0: since the evaporation of the liquid is the only step in the HP cycle 

which produces heating, the latent heat of a refrigerant should be as high as 

possible. Considering condensation and evaporation, ammonia (R717) is a 

better heat conductor, compared with the synthetic refrigerants. 

4. V - swept volume flow: compressors are often some of the most critical 

and expensive systems at a production facility, and deserve special attention. 

5. W - compressor work: for the efficiency of the process the compressor 

work is also of interest. The compressor work of isentropic compression from 

the temperature of 0°C to 70°C, 60°C, 52°C, 45°C to the condensing 

temperature, is given for the different fluids.


COP PT 

Puterea specifica a condensatorului [kJ/kg] 

6 

5 

4 

3 

2 

1 

0 

1400 

1200 

1000 

800 

600 

400 

200 

0 

R407c 

R134a 

R407c 

R134a 

Regim I Regim II Regim II 

R404a R407c R410a R507 R600 R717 R134a 

Fig. 2 – Comparative analysis of COP systems. 

R717 R717 R717 

R600 R600 

R407c 

R134a 

R600 

R134a R134a 

R404a 

R134a 

R404a 

R404a 

Regim I Regim II Regim II 

R404a R407c R410a R507 R600 R717 R134a 

Fig. 3 – Comparative analysis of the specific powers of condensation. 

6. H – compression ratio: when operating between two specific 

temperatures, fluids with low vapour pressures (high normal boiling point) will


have larger compression ratios than fluids with high vapour pressures. 

However, there are exceptions to this general rule and this can be clearly seen in 

the comparison. 

Table 2 

Numerical results for comparative analysis 

Agent COP 

*R404a 

*R407c 

*R410a 

*R507 

**R600 

**R717 

**R134a 

qk 

[kJ/kg] 

q0 

[kJ/kg] 

V 

[m 3 /kg] 

W [kW] 

H 

[/] 

5.132 135.605 109.18 144.13 3.012 3.392 

4.188 124.703 94.93 144.13 3.691 3.984 

3.288 109.812 76.41 144.13 4.702 4.754 

5.57 198.252 162.71 226.42 2.774 3.822 

4.697 189.393 149.07 226.42 3.291 4.55 

3.905 177.753 132.24 226.42 3.959 5.522 

5.24 193.921 156.96 145.43 2.946 3.382 

4.347 182.679 140.66 145.46 3.556 3.975 

3.499 166.665 119.03 145.46 4.418 4.747 

5.38 142.541 116.05 136.02 2.873 3.373 

4.463 133.971 103.96 136.02 3.463 3.97 

3.56 120.683 86.83 136.02 4.336 3.97 

5.22 347.6 254.17 1540.18 3.028 4.606 

4.473 335.574 232.78 1540.18 3.561 5.572 

3.758 319.233 213.47 1540.18 4.062 6.534 

5.11 1311.43 1054.9 1239.01 3.024 4.985 

4.44 1310.118 1015.1 1239.01 3.481 6.088 

3.813 1306.421 979.41 1239.01 3.882 7.712 

4.916 171.256 136.41 296.99 3.145 4.73 

4.154 162.769 123.58 296.99 3.722 5.74 

3.4 150.937 111.85 296.99 4.278 6.754 

*Cycle: 0/15/45; 0/15/52; 0/15/60; **Cycle: 0/15/52; 0/15/60; 0/15/70 

ODP GWP 

Price per 

unit 

$ 

0 3900 234 

0 1800 299 

0 2100 219 

0 4000 255 

0 20 172 

0 1 86 

0 1430 129 

7. ODP ozone depleting potential: the fact that these ‘‘harmless’’ 

substances were found to have a very unexpected and harmful impact on the 

global environment, raised doubts also about the use of other manmade 

substances, not present in the natural environment. Later, these doubts have 

been confirmed by the fact that the emissions of these refrigerants contributed 

by more than 20% to the global release of CO2 - equivalents during some years 

before the ban of the CFCs (IPCC/TEAP, 2005)


8. GWP - global warming potential: the refrigerant selection based on a 

simple approach of ‘zero ODP’ will have to pay high costs to both global 

warming and energy efficiency. Using this single criterion is no longer 

environmentally acceptable today. 

9. Price per Unit of refrigerant, $. 

Agent 

R134a 

R717 

R600 

R507 

R410a 

R407c 

R404a 

1 

20 

1430 

1800 

2100 

0 500 1000 1500 2000 2500 3000 3500 4000 4500 

GWP 

Fig. 4 – Comparative analysis of price. 

Fig. 5 – Comparative analysis (70°C, 60°C, 52°C, 45°C condensing temperature and 

0°C evaporating temperature). 

3900 

4000



1. The evaporation heat content of R404a is significantly lower (90% 

less) than the baseline R717 value, while that of R410a is slightly higher. 

2. The COP cycle of R134a is about 17% lower than the baseline value, 

indicating that this refrigerant does not match the baseline R407c cycle 

performance. These characteristics imply that both R404a and R410a 

refrigerants have weaker heat transfer ability and lower cooling capacity on 

equal heat pump load base, as compared to R717. 

3. Increased vapor density for R507, R410a and R404a may also lead to 

the use of a smaller compressor and coil tubing that could result in less system 

power consumption and more efficient component design. 

4. Overall, however, a comparable COP system is possible for R404a, and 

even a moderate system efficiency improvement might be expected for R410a 

for a certain heat pump load. 

5. The COP system is also influenced by compressor volumetric 

efficiency, which is a function of compression ratio H (pcond/pevap) and 

compressor isentropic efficiency, which could be affected by other transport 

properties, as well as system design. 

REFERENCES 

*** Report of the Refrigeration, Air Conditioning and Heat Pumps. UNEP, Technical 

Options Committee, Nairobi, 1998. 

*** Livre blanc sur les fluides frigorigènes. Conseil National du Froid, Paris, AFF, 

2001, 51 pages. 

*** Report of the Technology and Economic Assessment Panel April 2000. UNEP. 

2000, UNEP, Nairobi, 193 pages. 

Harnisch J., Hendriks C., Economic Evaluation of Emission Reductions of HFCs. PFCs 

and SF in Europe, Ecofys Energy and Environment, 2001. 

Chang Y.S., Kim M.S., Ro S.T., Performance and Heat Transfer Characteristics of 

Hydrocarbon Refrigerants in a Heat Pump System. International Journal of 

Refrigeration, 23 (3), 232-242 (2000). 

Pelletier O., Propaneas Refrigerant in Residential Heat Pumps. Licentiate thesis. Royal 

Institute of Technology, Stockholm, Sweden, 1998. 

METODA MULTICRITERIALĂ DE ALEGERE A AGENŢILOR FRIGORIFICI 

PENTRU POMPELE TERMICE 

Lucrarea propune o evaluare multicriterială pentru pompele termice care lucrează 

cu agenţi frigorifici cu ODP zero (R600, R404a, R407c, R410a, R134a, R507, R134a şi 

R717). Proprietăţilor termo-fizice ale agenţilor nu sunt suficiente pentru o evaluare 

completă, iar o abordare multicriterială permite o apreciere completă. Parametrii


suplimentari, cum ar fi eficienţa, siguranţa sistemului şi aspectele economice sunt 

incluse în analiza detaliată.


Publicat de 



Secţia 


THERMAL PROPERTIES OF POLYSTYRENE/CARBON 

NANOTUBES COMPOSITES PREPARED BY SHEAR CASTING 

BY 

RĂZVAN FLORIN BARZIC ∗1 , IULIANA STOICA 2 , 

ANDREEA IRINA BARZIC 2 and GHEORGHE DUMITRAŞCU 1 

1 “Gheorghe Asachi” Technical University of Iaşi, 

Faculty of Mechanics 

2 “Petru Poni”Institute of Macromolecular Chemistry, Iaşi 



Abstract. The polystyrene/carbon nanotubes (PS/CNT) composite films are 

obtained by a simple solution dispersion procedure. The carbon-based nanofillers 

are dispersed in methylene chloride, and PS/CNT composites are prepared by 

mixing PS/methylene chloride solution with CNT/methylene chloride dispersion, 

homogenized by ultrasonication. The corresponding films are characterized by 

atomic force microscopy (AFM) technique in order to evaluate the morphology 

developed by the resulted system. The thermal conductivity of these materials is 

estimated by considering it as an additive function of the PS and CNT 

compositions. The obtained results indicate that the studied nanocomposites 

exhibit high heat transfer – suitable for protecting chips of overheating in power 

electronics. 

Key words: thermal conductivity, polymer, carbon nanotubes. 


Thermally conductive polymer composites offer new possibilities for 

replacing metal parts in several applications, including power electronics, 

electric motors and generators, heat exchangers, thanks to the polymer 

advantages such as light weight, corrosion resistance and ease of processing. 

∗ Corresponding author: e-mail: barzicrazvan@tuiasi.ro, gdum@tuiasi.ro

60 Răzvan Florin Barzic et al. 

Current interest to improve the thermal conductivity of polymers is focused on 

the selective addition of nanofillers with high thermal conductivity. Unusually 

high thermal conductivity makes carbon nanotubes (CNT) the best promising 

candidate material for thermally conductive composites (Han & Fina, 2011). 

For an enhanced thermal conductivity the CNT must be well dispersed and 

oriented in the polymer matrix (Ke et al., 2007 and Riggs et al., 2000). 

The aim of the present study is to prepare some nanocomposites from 

commercial polymer (polystyrene) and CNT, the latter being dispersed by 

ultrasonication. The orientation of the used nanofillers was achieved by a new 

method which implies a shear field. The morphology of resulted 

nanocomposites is evaluated by atomic force microscopy. Thermal conductivity 

of these materials is estimated by considering it as an additive function of the 

PS and CNT compositions. All these aspects are presented in the context of the 

current state of research on thermal and morphological characterization of 

polymer composites in heat engineering-oriented applications, marking the own 

research and future directions. 

2. Experimental 

2.1. Materials 

Polystyrene (PS) was purchased from Sigma Aldrich having a molecular 

weight of 40 000 g/mol. Also, carbon nanotubes (CNT) were achieved from SC 

ICEFS COM SRL Savinesti having a diameter of 8…15 nm. To get a better 

dispersion and orientation of carbon nanotubes into the matrix of polystyrene 

(PS) a new method has been employed, which consists in shearing the 

composite starting from the solution phase. For this purpose, we used a solvent 

with high diffusion rate, i.e. methylene chloride, to maintain the organization 

imposed shear. 

Thus, the solution is deposited onto a device similar to that illustrated in 

Fig. 1, which has a moving blade, inducing shear controlled orientation of 

macromolecular chains and implicitly the carbon nanotubes. To investigate to 

what extent this technique leads to the desired result morphological studies were 

made on the nanocomposite films prepared as described previously. 

2.2. Characterization 

Atomic force microscopy (AFM) data were recorded with a camera Pro- 

M SPM Solver tool. Peak use is the type NSG10/Au Silicon, having a radius of 

curvature of 10 nm and an average oscillation frequency of 255 kHz. The 

images were scanned in semi-contact, their resolution is 256 × 256. 

The glass transition temperature of PS was determined with a Mettler 

differential scanning calorimeter.


3. Results and Discussion 

3.1. Morphology 

The detailed morphology of both investigated polymer and the 

corresponding nanocomposites subjected to shearing was examined by atomic 

force microscopy. The topography of the PS film obtained by shearing its 

solution in methylene chloride is shown in Fig. 2. AFM image of multiwalled 

carbon nanotubes introduced into the PS matrix is illustrated in Fig. 3. AFM 

phase contrast image of the PS/CNT nanocomposites films is presented in Fig. 

4, and their morphology in Fig. 5. 

From AFM data the following observations can be extracted: 

i) casting and shearing polystyrene solution leads to macroscopically 

oriented films, as shown in Fig. 2; 

ii) AFM phase contrast image shows that carbon nanotubes are oriented 

in the matrix of PS, with the length perpendicular to the composite film surface; 

iii) organization of carbon nanotubes in PS is ideal for improving the 

thermal conductivity, resulting in copolymers with materials designed to protect 

electrical components overheating circuit. 

3.2. Thermal Properties 

Thermal properties of PS/CNT composites are theoretically evaluated 

using a formalism based on connectivity indices developed by Bicerano 

(Bicerano, 1996). The glass transition temperature, Tg 

, is calculated using an 

equation with terms related the cohesive energy and van der Waals volume. The 

estimated value of Tg 

is 108.85°C, which is in agreement with experimental 

value obtained by differential calorimetry data namely of 103.2°C. In the case 

of nanocomposite films the glass transition temperature is high, i.e. 116°C 

revealing a good thermal stability induced by the presence of CNT. 

The thermal conductivity, λ , of these materials is estimated by 

considering it as an additive function of the PS and CNT compositions. First, 

the thermal conductivity of PS at room temperature was predicted using 

1 

0.135614 0.126611 0.108563( 0 0.125 ) 

BB 

+ χ 

N N + N − NH 

λ = + 

. (1) 

N N 

1 BB 

where χ denotes the portion of the first –order connectivity index contributed 

by the bounds between pairs of backbone atoms. 

1 

BB 

For polystyrene the value of χ is 0.8165. The thermal conductivity 

of PS calculated from Eq. (1) is 0.135 W/m K. Secondly, considering the 

thermal conductivity of the multiwalled carbon nanotubes, measured by the


3- ω method, as being 830 W/m K (Choi & Poulikakos, 2005), the λ of the 

nanocomposites was estimated from 

λnanocomposite = λPS wPS + λCNT wCNT 

. (2) 

where wPS is the volume fraction of PS and wCNT 

is the volume fraction of 

CNT. 

For a percent of 5% CNT in the PS matrix the thermal conductivity is 

increased up to 41.49 W/m K. The obtained results indicate that the studied 

nanocomposites could exhibit high heat transfer – suitable for protecting chips 

of overheating in power electronics. 

polystyrene (PS) 

Double-walled carbon 

nanotubes 

The orientation of the 

polymer matrix 

by shearing 

PS /CNT nanocomposite 

Orientation of carbon 

nanotubes in the PS matrix by 

shearing 

Fig. 1 – Scheme for preparation of nanocomposites PS/carbon nanotubes orientated by 

shear. 

Fig. 2 – AFM image of PS film obtained from a sheared solution in methylene chloride.


Fig. 3 – AFM image of CNT in ultrasonicated aqueous 

suspension 

deposited on silicon substrate. 

Fig. 4 – AFM phase contrast image of the nanocomposite PS/ carbon 

nanotubes film. 

Fig. 5 – 2D and 3D AFM images of nanocomposite PS/carbon nanotubes film. 


1. This study presents a new method for orientation 

of carbon nanotubes 

in polymer 

matrix using shear field. 

2. The AFM data show a good 

organization of the macromolecular chains 

and also of the CNT.


3. Theoretical values of thermal conductivity are high recommending the 

investigated nanocomposites 

as good thermal conductor materials for power 

electronics. 

Acknowledgements. This paper was realized with the support of POSDRU 

CUANTUMDOC 

“Doctoral Studies for European Performances in Research and 

Innovation” ID79407 project funded by the European Social Fund and Romanian 

Government. 

REFERENCES 

Bicerano 

J., Prediction of the Properties of Polymers from their Structures. J.M.S. – 

Rev. Macromol. Chem. Phys., C36, 161-196 (1996). 

Choi T. Y., Poulikakos D., Measurement of Thermal Conductivity of Idividual 

Multiwalled Carbon Nanotubes by the 3-ω Method. Appl. Phys. Lett., 87 (2005). 

Han Z., Fina A., Thermal Conductivity of Carbon Nanotubes and their Polymer 

Nanocomposites: A Review. Prog. Polym. Sci., 36, 914-944 (2011). 

Ke G., Guan W. C., Tang C. Y., Hu Z., Guan W. J., Zeng D. L., Deng F., Covalent 

Modification of Multiwalled Carbon Nanotubes with a Low Molecular Weight 

Chitosan. Chin. Chem. Lett., 18, 361-364 (2007). 

Riggs J. E., Guo Z., Carroll D. L., Sun Y. P., Strong Luminescence of Solubilized 

Carbon Nanotubes. J. Am. Chem. Soc., 122, 5879-5880 (2000). 

PROPRIETĂŢI TERMICE ALE NANOCOMPOZITELOR 

POLISTIREN/NANOTUBURI DE CARBON OBŢINUTE PRIN METODA 

FORFECĂRII 

(Rezumat) 

Pentru a obţine o orientare mai bună a nanotuburilor 

de carbon inserate în 

matricea de polistiren (PS) s-a utilizat o metodă nouă, 

şi anume forfecarea compozitului 

încă din faza de soluţie. 

În acest scop, s-a utilizat un solvent cu rată de difuzie mare, şi 

anume clorura de metilen, pentru a menţine organizarea impusă prin forfecare. Astfel, 

soluţia se depune pe suportul unui dispozitiv, care prezintă o lamă care se deplasează cu 

viteză controlată inducând prin forfecare orientarea catenelor macromoleculare şi 

implicit a nanotuburilor de carbon. Pentru a investiga în ce masură această tehnică 

conduce la rezultatul dorit s-au făcut studii morfologice a filmelor astfel preparate. 

Morfologia filmelor astfel preparate a fost investigată prin microscopie de forţă atomică 

(AFM) arată că nanotuburile de carbon sunt orientate în matricea de PS, având lungimea 

perpendiculară pe suprafaţa filmului compozit. Organizarea nanotuburilor de carbon în 

PS este ideală pentru îmbunătăţirea conductivităţii termice, rezultând materiale polimere 

compozite cu rolul de protejare a componentelor circuitului electric de supraîncălzire. 

Acest aspect este susţinut şi de calculele teoretice privind conductivitatea termică care 

pentru un procent de 5% CNT în matricea de PS creşte până la 41.49 W/m K.


Publicat de 



Secţia 


MODELLING OF A PLATE CROSS FLOW 

HEAT EXCHANGER 

BY 

IONEL IVANCU, DANIEL DRAGOMIR STANCIU ∗ , IONUŢ CRÎŞMARU, 

DAN TEODOR BĂLĂNESCU and GEORGE OVIDIU RĂU 

“Gheorghe Asachi” Technical University of Iaşi, 

Department of Mechanical and Automotive Engineering 



Abstract. The paper presents heat transfer modelling in a plate heat 

exchanger. Inside the heat exchanger, heat is transferred from the flue gas (of a 

hot water boiler) to air. The purpose of the paper is to analyse variation of the air 

parameters when air passes one channel of the heat exchanger. 

Key words: heat transfer, plate heat exchanger, modelling. 


Heat exchangers are devices used for heat transfer between two or more 

fluids with different temperature levels. They are essential elements for many 

thermal systems. For example, they make possible the heat transfer from a 

primary to a secondary circuit in a solar thermal system. 

Due to their great advantages (intensive heat transfer, low price, high heat 

recovery efficiency, compactness, large heat transfer surface, reduced pressure 

loss, easy operation) plate heat exchangers are widely spread and studied 

(Durmus et al., 2009), (Ghosh et al., 2006), (Chittur et al., 1990). 

The purpose of the developed study is the analysis of the heat transfer 

from flue gas (delivered by a hot water boiler) to air in a plate heat exchanger. 

The first stage of the study was the CFD analysis of the heat exchanger while 

∗ Corresponding author: e-mail: ddragomir03@yahoo.com

66 Ionel Ivancu et al. 

the second stage consists in the experimental analysis of the heat exchanger. 

The main objective of the study is the mutual validation of both analysis 

methods, theoretical and experimental. 

The paper presents the results obtained in the first stage of the study. 

2. The Experimental Heat Exchanger 

In order to set out the geometry of the plate heat exchanger, the 

information from (Jorge et al., 2004) was used. Two pictures of the heat 

exchanger are shown in Fig.1a and Fig.1b. The exchanger size is 300 x 300 

mm, with a number of 18 channels for air flow and 17 channels for flue gas 

flow. The distance between plates is 5 mm. The clamping of the plates is 

realized with threaded rods. The plates were made of galvanized steel. 

a b 

Fig. 1 – The experimental plate heat exchanger. 

In order to develop the experimental study, an experimental stand was 

conceived and developed. The experimental study is not the subject of the 

present paper and will be accomplished in the next stage of the research. 

3. Modelling of Heat Transfer 

The CAD modelling of both the separating wall and flow channel was 

made with CATIA V5 R19 software. Helpful information in this sense comes 

from (Haghshenas et al., 2011), (Gut & Pinto, 2003), (Gut et al., 2004), 

(Galeazzo et al., 2011). In order to achieve the temperature distribution in the 

separating wall of the exchanger, the Steady-Thermal module of ANSYS v13 

was used. To achieve the transfer, forced convective conditions were imposed 

for walls in contact with fluids. 

For the separating wall there were used hexahedron finite elements. The 

mesh was composed of 2548 of elements with 19088 nodes. For the modelling 

of the air flow channel, tetrahedral finite elements were used. The mesh was 

composed of 28579 elements with 9136 nodes. The mesh for the model is 

presented in Fig. 2.


The input data required by heat exchanger model are: the pressure – p1 

[N/m 2 ], the velocity – c1 [m/s] and the temperature – t1 [°C] of the fluid in the 

inlet section of the channel. The parameters have the following values: 

p1 = 101378 N/m 2 , c1 = 1.025 m/s, t1 = 18 °C. 

Fig. 2 –The mesh for the air flow channel. 

The air pressure drop Δp was considered constant and was measured on 

the experimental stand. In order to obtain the temperature on the inner surface 

of the air flow channel was necessary to accomplish the temperatures 

distribution in the separating wall. 

Fig. 3 –Distribution of temperature in the separating wall. 

The forced convective heat transfer from flue gas to air takes place in 

laminar conditions, over a plane plate. Proper calculation methods for this case 

are presented in (Popescu, 2003), (Badea & Necula, 2000), (Leca et al., 1998).


a b 

Fig. 4 –The experimental plate heat exchanger: 

a – distribution of the temperature; b – isothermal lines. 

a b 

Fig. 5 –Distribution of pressure: 

a – total pressure; b – static pressure. 

Nusselt and Reynolds criteria are calculated as follows 

cLc 

ℜ e = , 

(1) 

υ 

αx 

12 13 

N u= = 0.332 ℜe Pr 

. 

(2) 

λ 

The convective heat transfer coefficient results from Eq. (2). Variation of 

temperature in plate material is presented in Fig. 3. 

The model was made in the laminar flow hypothesis. CFD analysis of the 

model predicts the pressure and temperature distribution in the air channel.


The temperature field is shown in Fig. 4a while the isothermal lines are 

presented in Fig. 4b. 

The distribution of the total and static pressure are represented in Fig. 5a 

and Fig. 5b, respectively. 


1. The results of the analysis offer useful information regarding the 

variation of air parameters inside the heat exchanger. It can be observed that the 

mounting rods placed in the median line of the channel induce perturbations. 

2. Because of the rods placed in the inlet and outlet sections of the flow 

channel, the air velocity decreases. Consequently, the static pressure of air 

increases upstream the rods while the downstream static pressure of air 

decreases. 

3. The temperatures distribution is also influenced by the rods. Because 

the specific heat of rods material is higher than the specific heat of the air, the 

rods behave as thermal accumulators. Consequently, the air temperature around 

the rods is higher. 

Acknowledgements. This paper was realized with the support of POSDRU 

CUANTUMDOC “Doctoral Studies for European Performances in Research and 

Innovation” ID79407 project funded by the European Social Fund and Romanian 

Government. 

REFERENCES 

Durmus A., Benli H., Kurtbas I., Gül H., Investigation of Heat Transfer and Pressure 

Drop in Plate Heat Exchangers Having Different Surface Profiles. International 

Journal of Heat and Mass Transfer, 52, 1451-1457 (2009). 

Ghosh I., Sarangi S.K., Das P.K., An Alternate Algorithm for the Analysis of 

Multistream Plate Fin Heat Exchangers. International Journal of Heat and Mass 

Transfer, 49, 2889-2902 (2006). 

Chittur C.L., Owen E.P., Dynamic Simulation of Plate Heat Exchangers. International 

Journal of Heat and Mass Transfer, 33, 995-1002 (1990). 

Jorge A.W., Gut J., Pinto M., Optimal Configuration Design for Plate Heat 

Exchangers. International Journal of Heat and Mass Transfer, 47, 4833-4848 

(2004). 

Popescu A., Elemente fundamentale de transfer de căldură. Ed. Eurobit, Timişoara, 

2003. 

Badea A., Necula H., Schimbătoare de căldură. Ed. AGIR, Bucureşti, 2000. 

Leca A., Mladin E.C., Stan M., Transfer de căldură şi masă. Ed. Tehnică, Bucureşti, 

1998. 

Haghshenas Fard M., Talaie M.R., Nasr S., Numerical and Experimental Investigation 

of Heat Transfer of ZnO/Water Nanofluid in the Concentric Tube and Plate Heat 

Exchangers. Thermal Science, 15, 1, 183-194 (2011).


Gut J.A.W., Pinto J.M., Modelling of Plate Heat Exchangers with Generalized 

Configurations, International Journal of Heat and Mass Transfer, 46, 14, 2571- 

2585 (2003). 

Gut J.A.W. et al., Thermal Model Validation of Plate Heat Exchangers with 

Generalized Configuration. Chemical Engineering Science, 59, 21, 4591-4600 

(2004). 

Galeazzo F.C.C. et al., Experimental and Numerical Heat Transfer in a Plate Heat 

Exchanger. Chemical Engineering Science, 61, 21, 7133-7138 (2006). 

MODELAREA UNUI SCHIMBĂTOR DE CĂLDURĂ 

CU PLĂCI ÎN CURENT ÎNCRUCIŞAT 

(Rezumat) 

În lucrare este prezentat un schimbator de căldură cu plăci, în curent încrucişat, 

ce aparţine instalaţiei experimentale, cu dimensiunile şi materialul din care este realizat. 

Fluidele care trec prin schimbător sunt aerul şi gazele de ardere. 

Ca date iniţiale pentru modelarea schimbului de căldură s-au utilizat presiunea 

p1 , viteza c1 şi temperatura t1 

fluidului în regim de intrare. Modelele au fost realizate 

cu ajutorul programului CATIA iar pentru realizarea distribuţiei de temperaturi a fost 

folosit modulul STEADY-THERMAL din programul ANSYS. Modelarea a fost făcută 

în ipoteza curgerii laminare. 

În urma modelării s-au obţinut distribuţiile de temperatură şi presiune. 

Modelarea oferă informaţii utile privind variaţia parametrilor aerului la curgerea prin 

schimbătorul de căldură.


Publicat de 



Secţia 


EXPERIMENTAL RESULTS ON A VAPOUR INJECTION 

SCROLL COMPRESSOR 

BY 

ION ZABET 1 and GRAŢIELA-MARIA ŢÂRLEA ∗2 

1 Romanian General Association of Refrigeration of Bucharest, 




Abstract. This paper presents experimental results from a vapour injection 

scroll compressor system. The purpose of this paper is to test a vapour injection 

scroll compressor for a heat pump. For the measurements, the inputs were given 

by different pressure values (evaporation and condensation process). On this test 

rig we mount a pressure sensor on the evaporator supply and a new system of 

expansion valves. This new system of the expansion valves allow us to extend 

the tests matrix to a wide range of capacities and a higher control over the 

system. The expansion valves are mounted three on the evaporator and two on 

the injection evaporator. The work is conducted in two phases: the first phase of 

the rig consists in the set-up of a test rig, its instrumentation, the calibration of 

the measurements devices, the set-up of the data acquisition system and a few 

verification tests and the second phase consists in the achievement of the 100 

tests. The last part of the paper presents the results obtained by the experimental 

model. These results are: performance coefficient, power consumption, 

exhausted compressor temperature and refrigerant charge of the refrigerant 

R404A in different working conditions. 

Key words: vapour injection scroll compressor, exhausted compressor 

temperature, refrigerant charge and heat exchanger. 


The purpose of this study is to test a vapour injection scroll compressor 

and find a simplify model for a heat pump (Lemort, 2008), (Winandy & Lebrun, 


72 Ion Zabet and Graţiela-Maria Ţârlea 

2002). In the same time we will try to define the point where injection process 

take place on the: isentropic process, isochoric process or between them. The 

work is conducted in two phases: the first phase of the rig consists in the set-up 

of a test rig, its instrumentation, the calibration of the measurements devices, 

the set-up of the data acquisition system and a few verification tests; the second 

phase consists in the achievement of the 100 tests. On this test rig we mount a 

pressure sensor on the evaporator supply and a new system of expansion valves. 

This new system of the expansion valves allow us to extend the tests matrix to a 

wide range of capacities and a higher control over the system. The expansion 

valves are mounted three on the evaporator and two on the injection evaporator. 

2. Description of Experimental Model 

The schematic representation of the test rig is given in Fig. 1. The 

refrigerant circuit (Zabet, 2011) comprises a main evaporator, an injection 

evaporator, the compressor, a condenser and different expansion valves. The 

refrigerant is R407C. 

The main evaporator and injection evaporator are fed by a glycol water 

closed loop. 

Glycol water is an aqueous solution 50% of propylene glycol. The glycol 

water loop is composed of a pump, an electrical boiler, an expansion tank, a bypass 

loop and different manual control valves (at the inlet and outlet of the main 

evaporator and the injection evaporator). The condenser is cooled by tap water. 

Two type of test was performed: with injection and without injection. The 

compressor characteristics are: type scroll; injection mode vapour; swept 

volume 166 cm 3 ; N= 2900 rpm. The main evaporator is a helical heat 

exchanger. 

The second evaporator is a helical heat exchanger which is used for heat 

the refrigerant liquid to be injected in the compressor. The boiler is built by 

using two concentric cylinders. The internal cylinder is completely closed and 

filled with gaseous nitrogen. Glycol water circulates upwards between the 

external and the internal cylinders. Glycol water is heated by 10 electrical 

resistances (6 resistances of 9 kW and 4 resistances of 6 kW). A flow sensor is 

installed at the boiler exhaust, to protect the boiler if no water movement is 

detected. The nominal flow rate of the pump is 40m 3 /h. 

The boiler is equipped with a temperature regulation system, which 

allows maintaining a stable temperature set point at the boiler exhaust (by 

switching on and off some resistances). On the second evaporator it was 

mounted two expansion valves and two different orifice types in order to 

varying the injected refrigerant mass flow. On the main evaporator it was 

mounted three expansion valves with different orifice size. The expansion 

valves are mounted in parallel for both types of evaporator. The condenser is 

shell and tubes heat exchanger and use tap water for refrigerant condensation.


Fig. 1 – Schematic representation of the test rig. 

3. 3. Experimental Results Analyses 

Table 1 presents the results of the experimental test carried out without 

vapour injection. Each test point is characterized by the supply pressure (Psu), 

exhausted pressure (Pex), superheating temperature (ΔTSI) and subcooling 

temperature (ΔTsub). 

Table 2 presents the results of the experimental test carried out with 

vapour injection. Each test point is characterized by the supply pressure (Psu), 

exhausted pressure (Pex), superheating temperature (ΔTSI) and subcooling 

temperature (ΔTsub). Let be other types of situations. 

Figs. 2–8 (Zabet, 2011) present the cooling capacity (Φ0), heating 

capacity (Φcd), compressor power ( W ), coefficient of performance for heating 

& 

cp 

(COPh) and cooling (COPc), compressor exhausted temperature (tex;cp) versus 

pressure ratio (π) in the cases with/without injection. As can be seen from the 

figures, vapour injection method is better than the case without injection for 

pressure ratio between 5 and 7.5. 

For the vapour injection tests the injection mass flow rate ( inj 

M& ) varied 

from 0 to 0.027 kg/s, which corresponds to an injection ratio INJR from 0 to 

9.6%. The injection ratio is defined by injection mass flow rate divided by 

supply mass flow ratio.


Nr. 

Test 

Table 1 

Results of the test without vapour injection 

Set points Experimental results without injection 

P su ΔT SI Pex ΔT sub π t ex;cp 

M cd 

& cp W& Φcd Φ0 COPheat 

bar K bar K C kg/s kW kW kW 

1 2 5 17 2 8.50 111.00 0.044 5.32 10.60 5.78 1.99 

2 2 6 20 2 10.00 95.18 0.053 6.22 26.69 6.79 3.94 

3 3 9 12 3 4.00 115.00 0.096 5.31 12.20 17.61 1.96 

4 4 5 12 2 3.00 73.97 0.100 5.38 30.20 18.10 4.22 

5 5 3 14 2 2.80 121.20 0.131 6.38 12.41 23.21 1.72 

6 5 5 17 3 3.40 75.00 0.149 7.15 22.28 23.11 4.18 

7 6 11 19 4 3.17 109.70 0.171 8.17 28.79 27.03 2.86 

Nr. 

Test 

Table 1 (continued) 

Results of the test without vapour injection 

Set points Experimental results 

without injection 

P su ΔT SI Pex ΔT sub π COP cool ε s ε v 

bar K bar K 

1 2 5 17 2 8.50 1.09 0.65 0.76 

2 2 6 20 2 10.00 2.96 0.71 0.78 

3 3 9 12 3 4.00 1.09 0.66 0.87 

4 4 5 12 2 3.00 3.23 0.71 0.86 

5 5 3 14 2 2.80 0.91 0.63 0.86 

6 5 5 17 3 3.40 3.33 0.69 0.86 

7 6 11 19 4 3.17 1.96 0.72 0.86 

In Fig. 9 is presented the compressor power ( W ), the cooling capacity 

& 

(Φ0) and the coefficient of performance (COP) versus the injection ratio (INJR). 

As can be seen from Fig. 9 (Zabet, 2011) the cooling capacity has a slight 

increase for injection ratio up to 9%. 

After this level, the cooling capacity decreases. In the same time with 

cooling capacity increases, the compressor power increases in a similar 

proportion which gives a quite constant coefficient of performance in the final. 

cp

Nr. 

test 


Table 2 

Results of the test with vapour injection 


P su ΔT SI Pex ΔT sub π t ex;cp 

M cd 

& cp W& Φcd Φ0 COPheat 

bar K bar K C kg/s kW kW kW 

1 2 5 17 2 8.50 102.20 0.070 6.67 16.23 10.2 2.43 

2 2 6 20 2 10.00 107.60 0.080 7.51 17.72 10.9 2.36 

3 3 9 12 3 4.00 81.14 0.087 5.60 21.04 15.6 3.76 

4 4 5 12 2 3.00 67.22 0.119 5.77 26.69 21 4.62 

5 5 3 14 2 2.80 65.89 0.140 6.38 29.90 23.8 4.69 

6 5 5 17 3 3.40 74.23 0.161 7.56 32.78 25.7 4.33 

7 6 11 19 4 3.17 84.88 0.181 8.63 36.84 28.8 4.27 

Nr. 

Test 

Table 2 (continued) 

Results of the test with vapour injection 


Psu ΔTSI Pex ΔTsub π inj M& Pinj tinj;cp COPcool εs εv bar K bar K kg/s bar C 

1 2 5 17 2 8.50 0.027 4.4 2.9 1.53 0.52 0.76 

2 2 6 20 2 10.00 0.024 4.71 7.7 1.44 0.55 0.78 

3 3 9 12 3 4.00 0.02 4.82 20 2.79 0.56 0.81 

4 4 5 12 2 3.00 0.01 5.69 8.7 3.63 0.58 0.85 

5 5 3 14 2 2.80 0.011 6.68 15 3.73 0.58 0.84 

6 5 5 17 3 3.40 0.012 7.58 19 3.39 0.61 0.86 

7 6 11 19 4 3.17 0.015 8.81 26 3.34 0.62 0.87 

In the above Tables, P, p – pressure (bar), T – temperature (K), t – temperature 

(ºC), COP –performance coefficient (-), M& – mass flow rate (kg/s), W & – 

compressor power (kW), Φ – capacity (kW), ε – compressor efficiency (-), ΔTSI 

– superheating temperature difference (K), ΔTsub – subcooling temperature 

difference (K), π – pressure ratio (-), su –supply, ex – exhaust, inj – injection 

point, cp –compressor, s – isentropic compression, v – isochor compression, cd 

– condensing point, 0 – evaporating point. 

The same phenomenon happens for the compressor exhausted 

temperature. The discharge temperature increases very slightly up to 9%. After 

this point the discharge temperature starts to decrease, as it can be seen from


Fig. 10 (Zabet, 2011). This trend is due to two counteracting phenomenon: first 

the increase of the pressure inside the scrolls due to vapour injection and second 

the decrease of the enthalpy at injection point as the injection ratio increases. 

Φ0 [kW] 

20,00 

16,00 

12,00 

8,00 

Without Injection 

With Injection 

5,00 5,50 6,00 6,50 7,00 7,50 

π [−] 

Fig. 2 – Cooling capacity versus pressure ratio. 

W cp [kW] 

7,80 

7,20 

6,60 

6,00 

5,40 

4,80 



4,20 

5,00 5,50 6,00 6,50 7,00 7,50 

π [−] 

Fig. 3 – Compressor power versus pressure ratio. 

M cd [kg/s] 

0,126 

0,112 

0,098 

0,084 

0,070 

0,056 



0,042 

5,00 5,50 6,00 

π [−] 

6,50 7,00 7,50 

Fig. 4 – Total mass flow rate versus pressure ratio. 

As can be seen from Fig. 9 the cooling capacity has a slight increase for 

injection ratio up to 9%. After this level, the cooling capacity decreases. In the 

same time with cooling capacity increases, the compressor power increases in 

a similar proportion which gives a quite constant coefficient of performance in 

the final.

t ex;cp [°C] 


130 

120 

110 

100 

90 

80 

70 



60 

5,00 5,50 6,00 6,50 7,00 7,50 

π [−] 

Fig. 5 – Compressor exhausted temperature versus pressure ratio. 

Φcd [kW] 

COP h [-] 

30 

25 

20 

15 

10 



5 

5,00 5,50 6,00 6,50 7,00 7,50 

π [−] 

Fig. 6 – Heating capacity versus pressure ratio. 

4,00 

3,50 

3,00 

2,50 

2,00 



1,50 

5,00 5,50 6,00 6,50 7,00 7,50 

π [−] 

Fig. 7 – Heating COP versus pressure ratio. 

The same phenomenon is happens for the compressor exhausted 

temperature. The discharge temperature increases very slightly up to 9%. After 

this point the discharge temperature starts to decrease, as it can be seen from 

Fig. 10.


COP c [-] 

2,50 

2,00 

1,50 

1,00 

0,50 



0,00 

5,00 5,50 6,00 6,50 7,00 7,50 

π [-] 

Fig. 8 – Cooling COP versus pressure ratio. 

W cp [kW], COP [-], Φ 0 [kW] 

35 

28 

21 

14 

7 

0 

Wcp COP 

0,08 0,084 0,088 0,092 0,096 

INJR=M inj;cp / Msu;cp [-] 

Fig. 9 – Compressor power, cooling capacity and COP versus vapour injection ratio. 

This trend is due to two counteracting phenomenon: first the increase of 

the pressure inside the scrolls due to vapour injection and second the decrease 

of the enthalpy at injection point as the injection ratio increases. 

t ex;cp [°C] 

128 

112 

96 

80 

64 

48 

Φ0 

t ex;cp 

0,08 0,084 0,088 0,092 0,096 

INJR=M inj;cp / Msu;cp [-] 

Fig. 10 – Compressor discharge temperature versus vapour injection ratio. 


1. The experimental analysis shown in this paper allows us to compare 

the compressor behaviour with two different configurations: the first one 

without injection and the second one with vapour injection.


2. It is shown how the vapour injection permits one to increase slightly 

the cooling capacity for injection ratio up to 9%. After this level, the cooling 

capacity decreases. 

3. In the same time with cooling capacity increases, the compressor 

power increases in a similar proportion which gives a quite constant coefficient 

of performance. 

Acknowledgements. The authors would like to thank COPELAND S.A. Europe 

and Thermodynamic Laboratory from University of Liege. 

REFERENCES 

Winandy E., Lebrun J., Scroll Compressors Using Gas and Liquid Injection: 

Experimental Analysis and Modelling. Int. J. of Refrigeration, 25, 1143-1156 

(2002). 



Lemort V., Contribution to the Characterization of Scroll Machines in Compressor and 

Expander Modes. PhD Thesis, Liège, Belgium, 2008. 

REZULTATELE EXPERIMENTALE ALE PRIVIND UN COMPRESOR SCROLL 

CU INJECŢIE DE VAPORI 

(Rezumat) 

Criza energetică majoră care afectează în prezent economia mondială şi 

încălzirea globală, impun creşterea performanţelor energetice şi ecologice ale 

echipamentelor şi sistemelor frigorifice şi de aer condiţionat. În acest sens, utilizarea 

compresoarelor scroll cu injecţie de vapori şi compresoarelor scroll standard şi a 

agenţilor frigorifici ecologici deschid aria unor multiple aplicaţii în domeniul energetic 

şi al ingineriei mecanice. 

În acest sens s-a realizat un studiu experimental având o matrice de testare 

compusă din 100 puncte, caracterizate prin diferite presiuni de vaporizare şi condensare. 

Realizarea testelor s-a efectuat în două etape: prima fază în realizarea modelului 

experimental (setarea, instalarea, calibrarea instrumentelor de măsurat, setarea 

sistemului de achiziţie şi realizarea testelor de verificare a standului experimental) şi a 

două fază realizarea experimentelor şi analizarea performanţelor compresorului scroll 

cu injecţie de vapori. Sistemul de control al modelului experimental s-a realizat prin 

intermediul următoarelor reglaje: presiunea la aspiraţia compresorului a fost ajustată 

prin controlul debitului masic de agent intermediar care circulă prin vaporizatorul 

principal, supraîncălzirea vaporilor la aspiraţia compresorului fiind dată de ventilele de 

laminare poziţionate la intrarea în vaporizatorul principal, presiunea la injecţia în 

compresor a fost reglată prin controlul debitului masic de agent intermediar care 

parcurge vaporizatorul aferent procesului de injecţie, supraîncălzirea vaporilor la 

injecţie fiind controlată cu ajutorul ventilelor de laminare la intrarea în vaporizatorului 

secundar (pentru injecţie) si presiunea la procesul de condensare a fost controlată şi 

setată prin reglarea debitului masic de apă care parcurge condensatorul.


Măsurătorile au fost făcute la secundă, mediate pe un interval de timp de 5 

minute după atingerea echilibrului termodinamic. 

In urma cercetărilor efectuate s-a deschis perspectiva unor noi direcţii de studiu 

în domeniul optimizării sistemelor frigorifice în procesul de injecţie cu vapori la 

compresoarele scroll cum ar fi: simplificarea modelelor de calcul care să conţină un 

număr cât mai redus de parametri şi ecuaţii, continuarea studiului cu aplicaţii în 

domeniul pompelor de căldura, analizarea performanţelor sistemelor frigorifice şi de aer 

condiţionat funcţionând cu compresoare scroll pentru diverşi agenţi frigorifici ecologici 

(cu ODP zero şi GWP redus).


Publicat de 



Secţia 


A STUDY ON PERFORMANCE OF A HYDROGEN FUELLED 

SPARK IGNITION ENGINE 

BY 

VICTOR PANTILE ∗ , CONSTANTIN PANĂ and NICULAE NEGURESCU 

“Politehnica” University of Bucureşti, 

Department of Mechanical Engineering and Mechatronics 

Received: 26 March 2012 

Accepted for publication: 14 April 2012 

Abstract. The use of alternative fuels represents a reliable alternative to 

decrease the pollutant emissions and it can also increase the independence 

regarding the fossil fuels. The combustion properties of hydrogen make this fuel 

one of the best solution to replace the classical fuels. Many researches using 

hydrogen were made and most of them had good results. In addition to 

experimental research, the modelling of the thermo-gas-dynamics processes 

inside the cylinder represents an effective investigation tool of engine operation, 

reducing the time duration and cost of experimental investigation. The paper 

presents results of the thermo-gas-dynamics processes modelling inside the 

engine fuelled with hydrogen. The results obtained show that hydrogen is an 

excellent fuel, allowing engine operation with the best energetically and 

pollution performance. At the use of hydrogen the engine load qualitative control 

strategy can be adopted together with the quantitative adjustment. 

Key words: internal combustion engines, hydrogen, modelling, 

performance. 


Considering the pollution problems and the declining fuel reserves today, 

investigations have been concentrating on increasing efficiency and lowering 

the pollutant emissions by using alternative fuels. 

Due to its properties, hydrogen is considered an ideal alternative fuel and 

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

82 Victor Pantilie et al. 

many researchers have studied the effect of using hydrogen as a fuel (pure or 

mixed with another fuel) on pollutant emissions and engine performance (Das et 

al., 2000), (Al-Baghdadi & Al-Janabi, 2000), (Yamin et al., 2000), (Maher et 

al., 2003). 

The combustion of hydrogen does not create pollutant emissions such as 

CO2, CO, HC or PM because hydrogen does not contain any carbon. From the 

oil combustion, some fractions of unburned hydrocarbons are present inside the 

engine combustion chamber but these are negligible (Pantile, 2010). The massspecific 

lower heating value of hydrogen is almost three times as high as that of 

gasoline thus the brake specific fuel consumption is reduced ant the thermal 

efficiency of the engine is increased. Hydrogen has antiknock quality due to the 

self-ignition temperature of the hydrogen/air mixture which is greater than that 

of gasoline. A combustible mixture is produced by a small amount of hydrogen 

mixed with air, which can be burned in a conventional spark ignition engine at 

an equivalence ratio below the lean flammability limit of gasoline/air mixture. 

The resulting ultra-lean combustion produces low flame temperature and leads 

directly to lower heat transfer to the walls, higher engine efficiency and lower 

NOx emissions in the exhaust (Al-Baghdadi & Al-Janabi, 2003). The laminar 

flame speed of a hydrogen-air mixture at stoichiometric conditions is almost 6 

times higher than that of gasoline. At lean dosage conditions (λ=2) the laminar 

flame speed of hydrogen is approximately equivalent to that of a stoichiometric 

gasoline-air mixture (Verhelst & Wallner, 2009). The hydrogen flammability 

limits are wider compared to gasoline. This means that the load of the spark 

ignition engine can be controlled by the excess air-fuel ratio – the load engine 

qualitative adjustment. Furthermore, the compression ratio of the engine can be 

increased due to the higher octane number. The main disadvantage comes from 

the low density of hydrogen which influences the storage of hydrogen and the 

power output which is lower compared to liquid fuels (when comparing engines 

with external mixture formation), (Pantile et al., 2011). 

The U.S. Department of Energy and Department of Transportation have 

taken initiatives to shift towards a hydrogen-based transportation system. The 

aim of these researches is to develop and commercialize hydrogen fuel cell 

vehicles in an economic manner. However, hydrogen can be used as a fuel for 

transportation in internal combustion engines rather than in fuel cells at least for 

some decades (Romm, 2006), (Sukumaran & Kong, 2010). 

The lack of efficient and economical hydrogen infrastructure represents 

one of the major issues in exploiting hydrogen energy. Internal combustion 

engines can be considered as a transition technology for achieving economical 

hydrogen infrastructure (Safaria et al., 2009). Developing hydrogen programs 

are conducted by several automotive companies especially by BMW Group and 

Ford Motor Company. BMW has introduced its hydrogen engine in the 7 series; 

6.0 L V12 bi-fuel (gasoline – hydrogen) engine which operates with liquid 

hydrogen stored on board (Kiesgen et al., 2006). Also, Ford has optimized


Zetec 2.0 L, I4 engine fuelled with hydrogen dedicatedly (Stockhausen et al., 

2002), (Tang et al., 2002). 

Table 1 

Hydrogen and gasoline properties 

Property Hydrogen Gasoline 

Molecular mass, kg/kmol 2.016 114 

Density, kg/m 3 0.089 750 

Theoretical air-fuel ratio, kg/kg comb 34.32 14.5 

Flame velocity in air, m/s 2.37 0.12 

Octane number >130 90…98 

Auto-ignition temperature, K 850 750 

Lower heating value, KJ/kg 119600 42690 

Minimum ignition energy, mJ 0.018 0.25 

Flammability limits in air, % 4…75 1…7.6 

Laminar burning velocity in air, m/s 2…2.3 0.37…0.43 

Normal boiling point temperature, K 20 310…478 

Stoichiometric composition in air, % 29.5 1.65 

Adiabatic flame temperature, K 2384 2270 

Although fuel cells have better efficiency than hydrogen engines, the cost 

of a hydrogen internal combustion engine is much less than a fuel cell and the 

power system of a vehicle. To use hydrogen in an internal combustion engine, 

some modifications have to be made but the working process is basically the 

same. Some difficulties should be overcome before the engines go into a 

common use, such as detonation, backfire and abnormal combustion (Jorach, 

1997). By using a suitable mathematical model the performance of a hydrogen 

engine can be simulated therefore it is possible to determine a beneficial 

working range for optimizing and predicting the engine’s performance (Jie et 

al., 2003). 

Nowadays, the development of an internal combustion engine is based on 

a close link between experimental engine testing and numerical simulation. 

Multi-dimensional and one dimensional thermo-fluid dynamic models are 

commonly used to optimize the engine design through the prediction of the 

unsteady flows in the intake and exhaust systems (Onorati et al., 2007), 

(Winterbone & Pearson, 2000), (Morel et al., 2003), (Montenegro et al., 2005), 

(Onorati et al., 2004 a), (Onorati et al., 2004 b), and of the combustion and 

emission formation processes in the cylinder (D’Errico et al., 2002), (D’Errico 

& Lucchini, 2005), (D’Errico et al., 2008). 

2. Processes Modelling 

There are many numerical simulation models used and because it is less 

expensive than experimental engine testing it is widely used in academic and


industrial research. The simulation program used in the present paper is AVL 

Boost v2011. 

The modelling was made for the DOHC Daewoo engine (aspirated 

engine and supercharged engine) with the characteristics from Table 2. 

The combustion law used was the Vibe law for two zones and the heat 

transfer coefficient used was Woschni. 

Table 2 

Engine specifications 

Engine type DOHC 

Displacement 1.5, litres 

Stroke 81.5, mm 

Bore 76.5, mm 

Cylinders 4 

Number of valves 16 

Connecting rod length 120, mm 

Compression ratio 9.2 

Maximum power engine speed 4800, rpm 

Engine operating regimes: full load and 3500 rpm with the next 

supercharge pressure and excess air-fuel ratio: 

gasoline with normal intake at λ=0.8, 0.85, 0.9, 1; 

gasoline with supercharge pressure 1.3 at λ=0.8, 0.85, 0.9; 

gasoline with supercharge pressure 1.5 λ=0.8, 0.85, 0.9; 

hydrogen with normal intake at λ=1, 1.2, 1.3, 1.5, 2, 2.5; 

hydrogen with supercharge pressure 1.3 at λ=1, 1.3, 1.4, 1.5, 2, 2.5; 

hydrogen with supercharge pressure 1.5 at λ=1, 1.4, 1.5, 2, 2.5; 

For each simulation were made 50 cycles for better results. 

Supercharging the engine prevent backfire, cools the cylinder walls and 

increases the engine performance, allowing hydrogen to be used in safe 

conditions and with lower pollutant emissions. To protect the engine against 

damage the supercharge pressure was limited to 1.5 bar. 

3. Results 

For the engine running on gasoline with normal intake the maximum 

pressure is 48 bar. When the engine is supercharged the pressure increases 

significantly reaching 87 bar with a supercharge pressure of 1.5 bar. Using 

hydrogen as a fuel, the maximum pressure with normal intake is higher at 

stoichiometric excess air-fuel ratio and as the dosage was leaner the pressure 

drops at 43 bar at excess air-fuel ratio equal to 2.5. To achieve similar 

performance as with gasoline (normal intake), the engine running on hydrogen 

is supercharged and for the reduction of the NOx emissions, the values for the 

excess air-fuel ratio should be greater than 1.8. The supercharge pressure used is


1.3 bar and 1.5 bar. Maximum pressure reached when using supercharged 

hydrogen is around 89 bar at stoichiometric excess air-fuel ratio (Fig. 1). 

Fig. 1 – Maximum pressure diagram. 

The maximum temperatures when using gasoline presents a slight 

increase as the engine was supercharged and as the excess air-fuel ratio 

increased from 0.8 to 0.9 (Fig. 2). 

On hydrogen the excess air-fuel ratio has a major impact so the maximum 

temperatures reached 2900 K when the engine is supercharged at 1.5 bar at 

stoichiometric excess air-fuel ratio and dropped below 1900 K at leaner dosage 

(λ=2.5). 

Fig. 2 – Maximum temperature diagram. 

The brake specific fuel consumption is much lower at the use of 

hydrogen due to good hydrogen burning proprieties and heat losses reduction. 

A slight impact on BSFC has also the supercharge pressure, and it can be 

seen in the figure below that when the engine is supercharged the BSFC 

decreases on gasoline as well as on hydrogen.


Another influence factor is the excess air-fuel ratio. On gasoline BSFC 

decreases when λ increased from 0.8 to 0.9. On hydrogen BSFC decreases when 

the excess air-fuel ratio increased from 1 to 1.5 and when the dosage is leaner 

the BSFC has a slight increase (Fig. 3). 

Fig. 3 – Brake specific fuel consumption diagram. 

The NOx emissions level on gasoline increases with the increasing excess 

air-fuel ratio from 0.8 to 0.9, but decreases when the engine is supercharged due 

to the timing adjustment which decreased to avoid detonation (Fig. 4). 

Fig. 4 – Gasoline NOx emissions diagram. 

When hydrogen was used the NOx emissions level had a different 

behaviour (Fig. 5). When normal intake was used the maximum was reached at 

λ=1.3 with more than 5000 ppm. 

When the engine was supercharged with 1.3 bar supercharge pressure the 

maximum emissions level was reached at λ=1.4 and when the engine was


supercharged with 1.5 bar supercharge pressure the maximum emissions level 

was reached at λ=1.5. From this point the emissions level dropped significantly 

and as the dosage was leaner the emissions levels were very low. 

Fig. 5 – Hydrogen NOx emissions diagram. 

Fig. 6 – NOx emissions vs. power and excess air-fuel ratio diagram. 

The hydrocarbons levels are significantly reduced and there are no carbon 

monoxide emissions when using hydrogen as a fuel.


At the engine running on gasoline calculated around 2000 ppm NOx 

emissions at stoichiometric excess air-fuel ratio. At the hydrogen fuelled engine 

running to limit this NOx level must be used the excess air-fuel ratio values 

greater than 1.8 (Fig. 6). 

At this dosage the power output on hydrogen with supercharge pressure 

of 1.5 bar is greater than the power output on gasoline with normal intake and at 

excess air-fuel ratio greater than 2 the power is similar but the NOx emissions 

level is significantly reduced (less than 50 ppm). 


1. This paper confirms that hydrogen can be used in internal combustion 

engines and by supercharging a spark ignition engine it can obtain similar 

performance with lower emissions. 

2. Using hydrogen without supercharging it is possible but with a major 

impact on the power of the engine if hydrogen is admitted in the intake. 

3. The NOx emissions level dropped below 50 ppm at λ=2 and if the 

dosage is leaner it can achieve near zero NOx emissions. 

4. The high engine efficiency when using hydrogen makes this fuel a 

perfect fuel for urban cycles due to combustion improvement. 

Acknowledgements. The work has been funded by the Sectoral Operational 

Programme Human Resources Development 2007-2013 of the Romanian Ministry of 

Labour, Family and Social Protection through the Financial Agreement 

POSDRU/88/1.5/S/61178. 

The authors would like to thank AVL GMBH Graz Austria for providing the 

software AVL Boost v2011 used in this study. 

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paper 2007-01-0495, Detroit, 2007. 

Onorati A., Ferrari G., D’Errico G., Secondary Air Injection in the Exhaust after 

Treatment System of S.I. Engines: 1D Fluid Dynamic Modelling and 

Experimental Investigation. SAE Transactions, Journal of Engines, 112–113, 

544-55 (2004). 

Onorati A., Ferrari G., Montenegro G., Caraceni A., Pallotti P., Prediction of S.I. 

Engine Emissions During an ECE Driving Cycle via Integrated Thermo-fluid 

Dynamic Simulation. SAE Int. Congress & Exp. Detroit, Michigan, paper 2004- 

01-1001, March 8–11 2004, Detroit (MI) 2004. 

Pantile V., Rusu E., Pana C., Aspects of the Hydrogen Use at the SI Engine. Conference 

Proceedings of the Academy of Romanian Scientists PRODUCTICA Scientific 

Session, 2011. 

Pantile V., Regarding to the Use of Hydrogen in SI Engine. CONAT, S007, 47-54 

(2010). 

Romm J., The Car and Fuel of the Future. Energy Policy, 34, 2609-2614 (2006). 

Safaria H., Jazayeri S.A., Ebrahimi R., Potentials of NOX Emission Reduction Methods 

in SI hydrogen Engines: Simulation Study. Int. J. Hydrogen Energy, 34, 1015- 

1025 (2009). 

Stockhausen W.F., Natkin R.J., Kabat D.M., Reams L, Tang X, Hashemi S., Ford 

P2000 Hydrogen Engine Design and Vehicle Development Program. SAE Paper 

2002-01-0240 (2002). 

Sujith Sukumaran, Song-Charng Kong., Numerical Study on Mixture Formation 

Characteristics in a Direct-injection Hydrogen Engine. Int. J Hydrogen Energy 

35, 7991-8007 (2010).


Tang X., Kabat D.M., Natkin R.J., Stockhausen W.F., Heffel J., Ford P2000 Hydrogen 

Engine Dynamometer Development. SAE Paper 2002-01-0242 (2002). 

Verhelst S., Wallner T., Hydrogen-fueled Internal Combustion Engines. Progress in 

Energy and Combustion Science, 35, 1, 490-527 (2009). 

Winterbone D.E., Pearson R.J., Theory of Engine Manifolds Design. London: 

Professional Engineering Publishing, London, 2000. 

Yamin J.A.A., Gupta H.N., Bansal B.B., Srivastava O.N., Effect of Combustion 

Duration on the Performance and Emission Characteristics of a Spark Ignition 

Engine Using Hydrogen as a Fuel. Int. J. Hydrogen Energy, 25, 6, 581 (2000). 

STUDIU ASUPRA PERFORMANŢELOR MOTORULUI CU APRINDERE PRIN 

SCÂNTEIE ALIMENTAT CU HIDROGEN 

(Rezumat) 

Lucrarea prezintă rezultatele modelării termo-gazo-dinamice la folosirea 

hidrogenului în motorul cu aprindere prin scânteie. Rezultatele obţinute arată că 

hidrogenul este un combustibil excelent, care permite funcţionarea motorului cu 

performanţe energetice ridicate şi emisii poluante reduse atunci când se foloseşte 

supraalimentarea şi valori ale coeficientului de exces de aer λ>1.8.


Publicat de 



Secţia 


ASPECTS REGARDING THE USE OF BIOETHANOL IN SPARK 

IGNITION ENGINES 

BY 

ALEXANDRU RADU ∗ , CONSTANTIN PANĂ and NICULAE NEGURESCU 

“Politehnica” University Bucharest, 

Department of Mechanical Engineering and Mechatronics 



Abstract. During the last years the spark ignition engines have been 

further improved in order to reduce carbon dioxide emissions (and thus the 

greenhouse effect) and fuel consumption, and increasing the engine performance. 

The use of alternative fuels represents an efficiency solution for these objectives. 

This explains the intense interest world-wide to using bioethanol (originating 

from renewable sources) as an automotive fuel, especially in blending with 

gasoline. This paper presents results of simulating to thermo-gas-dynamics 

processes inside the cylinder of an SI engine fuelled with ethanol-gasoline blends 

with different percentages of ethanol at different engine operating conditions. 

The main topics presented are the effects of using blends on: engine 

performance, emissions and fuel consumption. The results showed that an 

improved engine performance and lower emissions can be obtained with higher 

ethanol percentage in the blend, due to the better combustion properties ethanol. 

Keywords: spark ignition engine, bioethanol, fuel properties 

characteristics, simulation, NOx emissions. 


Considering the energy crisis and pollution problems, our investigations 

have been concentrating on reducing the fuel consumption and lowering the 

concentration of toxic components in combustion products by opting for 

alternative fuels. Ethanol is considered as an ideal alternative fuel. Ethanol has 

been recognized as an alternate fuel because of highly desirable properties. 

∗ Corresponding author: e-mail: radu_alex_85@yahoo.com

92 Alexandru Radu et al. 

Furthermore the ethanol could be a renewable source of energy if is 

produced from biomass and since its characteristics are fairly similar to the 

gasoline ones, the conversion of power units from one fuel to another looks 

quite promising in terms of efforts and costs (Turner et al., 2007), (Ku et al., 

2000), (Nakata & Utsumi, 2006), (Baeta et al., 2005). 

Recently more scientists have recognized ethanol as the best choice 

among alcohols for use in SI engines. Their studies have evaluated the effect of 

fuel properties and characteristics, and demonstrated the effects of ethanol on 

engine performance and on the environment. For example, Kumar et al. (Kumar 

et al., 2009) studied three different ethanol-gasoline blends i.e. 10, 30 and 70 % 

ethanol blended with gasoline (E10, E30 and E70 respectively), in a onecylinder 

engine, 500cc, water-cooled, AVL made a research engine with a CR 

10:1. The research engine was coupled with 80 kW, water cooled, eddy current 

dynamometer. The net effect on emission was significant reduction in CO and 

increase in NOx emission with increasing ethanol content in the blend. However 

no significant change in HC emission was observed. Operating with E70 blend, 

CO reduces by approximately 60% and NOx increases by approximately 50- 

60% at all operating speed under WOT condition. 

Wen Dai et al. (Dai et al., 2003) have developed an ethanol model in an engine 

cycle simulation tool, GESIM (General Engine Simulation program), for the 

simulation of spark-ignition engines using ethanol and ethanol-gasoline blends 

as fuels (E22 and E85). The developed model has demonstrated that CO 

emissions and unburned hydrocarbons generally decrease, while fuel 

consumption increases because of its low lower heating value. 

M. Bahattin Celik studied mixtures of E0, E25, E50, E75 and E100 

fuels in a single-cylinder four-stroke small engine whose original compression 

ratio was 6/1, at a constant load and speed (Celik, 2008). The compression ratio 

was raised from 6/1 to 10/1. The experimental results have shown that the most 

suitable fuel in terms of performance and emissions was E50. Therefore, engine 

power increased by about 29% when running with E50 fuel compared to the 

running with E0 fuel and the specific fuel consumption, and CO, CO2, HC and 

NOx emissions were reduced by about 3%, 53%, 10%, 12% and 19%, 

respectively. 

Gogos and al. (Gogos et al., 2008) tested the fuels E0, E10, E20 and E50 

in a 1300cc old technology vehicle without a catalytic converter. The results 

have shown that increasing the ethanol percentage in the blend has decreased 

the CO and HC emissions but increased the NOx emissions. For fuels E10 and 

E20 an increase on the engine’s brake torque and power along with a decrease 

in fuel consumption were observed and for E50, both brake torque and power 

were reduced. 

In another study, Nakama et al. (Nakama et al., 2008) ethanol– 

gasoline blended fuels (E3, E10, E20, E40, E60, E80 and E100) were tested 

in a single-cylinder engine with a displacement of 0.325 litres that has been


constructed based on the Suzuki M13A. Tests were conducted under conditions 

of ε=9.5, 11.0, 13.9,and 15.0 at engine speeds of 1500, 2000, 3000, and 4000 

rpm, and with full load and half load. It was confirmed that increasing the 

compression ratio together with increasing ethanol content is effective to 

overcome the shortcomings of poor fuel economy caused by the low calorific 

value of ethanol. 

2. Properties of Ethanol 

To study the effects of ethanol/gasoline mixtures on the performance of 

an internal combustion engine one has to consider some of ethanol’s properties. 

The most important are the following: 

Ethanol has a higher octane rating (RON=106) than gasoline and because 

of its antiknock properties, engines with a higher compression ratio, and 

subsequently more power, can be designed. 

A major chemical difference in the fuels is the inclusion of oxygen in the 

ethanol (Table 1). This allows the use of a poorer air to fuel mixture, and as a 

result a better combustion is achieved. This fact further causes fewer emissions 

of CO and unburned HC (Radu & Pană, 2007). 

Table 1 

Physical and chemical properties of gasoline and ethanol 

Properties Gasoline Ethanol 

Density at 15, kg/m 3 735..760 792 

Boiling temperature (at 1.013 bar), o 30..190 78 

In flammability limits: λs..λi 0,4..1,4 0,3..1,56 

Reid pressure, daN/cm 2 0.8..0.9 0.14 

Auto-ignition temperature, o 257..327 420 

Lower heating value, kJ/kg 43500 26800 

Fuel in water, % negligible 100 

Heat of vaporization, kJ/kg 290..380 904 

Octane number MON/RON 90/98 87/106 

Composition : C/H/O, % mass 85/15/0 52/13/35 

Dynamic viscosity at 0 o , mPa s 0.72..0.74 0.796 

Stoichiometric burning air, kg/kg 14.5 9 

As shown in Table 1, the heat of vaporization of ethanol is approximately 

three times higher than that of gasoline and requires more energy to vaporize 

the fuel. This property can contribute to the increase of engine power and 

efficiency because of the resulting higher density of the mixture.


Spark timing is important because of ignition delay issues. Ethanol 

actually has a slightly higher flame speed in smooth combustion than does 

gasoline. However, it has a much longer delay between application of the spark 

and a fully-formed flame (the “ignition delay”). This means that the spark 

timing may need to be advanced. 

Ethanol flammability range is much wider for air-ethanol mixtures 

comparative to gasoline (0, 3...1, 56 versus 0, 4...1, 4) providing engine run 

stability in the area of lean mixtures (Pană et al., 2007). 

The enthalpy of evaporation of ethanol is 842-930 kJ/kg that is higher 

than the value of 330-440 kJ/kg for gasoline. This property can contribute to the 

increase of engine power and efficiency because of the resulting higher density 

of the mixture, but this also causes ignition problems at low temperatures 

(Cowart et al., 1996). 

3. Modelling the Engine Cylinder Process 

The main reason for the growth in engine modelling activities arises from 

the economic benefits and to reduce the time. Real engine testing cannot be 

replaced by models, but they are able to provide good estimates of performance 

changes resulting from possible engine modifications. 

There are now available comprehensive “commercial” models, which 

have a wide purpose of use with refined inputs and outputs to facilitate their use 

by engineers. Also, many universities have produced their own thermodynamic 

models, of varying degrees of complexity, scope and ease of use. 

Such software is AVL BOOST v.2011.6 with which we have developed a 

model for the engine running simulation with gasoline and ethanol-gasoline 

blends and with normal intake and supercharged admission. The engine 

numerical model was validated by comparing the obtained values with the 

experimental data when using gasoline and E20 as fuels, on a 1.5L DOHC 

engine (76.5 mm bore, 81.5 mm stroke, 9.2 compression ratio). 

The theoretical researches were carried at engine speed 2500 rpm, full 

load and different dosage values (λ=0.8; 0.85; 0.9 for gasoline and λ=1.0; 1.1; 

1.2 for blend) for two fuels: gasoline and E20 (80% gasoline blend with 20% 

ethanol) and with different intake pressures for the normal admission engine 

and supercharged engine with 1.2 bar. For E20 the tests were carried out for 

dosages λ>1 in order to emphasize the main advantages of using ethanol in the 

lean mixtures field. 

The present investigation specifically concerns the engine performance 

and reduction of pollutant emissions from the exhaust of engines using ethanol 

fuel as a substitute for gasoline. 

3.1. Results 

Fig. 1 shows that peak cylinder pressure (pmax) and indicated mean 

effective pressure (IMEP) increases for E20 gasoline-ethanol blend comparative


with gasoline in both cases: normal intake and supercharged. Values of the two 

parameters have decreased for leaner mixtures in the field of λ=1…1,2. 

Increased peak cylinder pressure is due to higher burning rate of ethanol 

determinates a reduced of the combustion duration. 

Fig. 1 – Effect of ethanol addition on peak cylinder pressure. 

In Fig. 2 can be observed that the peak cylinder temperature is higher in 

case of the E20 fuel comparative with gasoline especially to supercharged 

engine. In order to reduce the NOx emissions at the engine fuelled with E20 the 

dosage must be leaner (e.g. λ=1.2). 

Fig. 2 - Effect of ethanol addition on peak cylinder temperature.


Fig. 3 – Brake specific fuel consumption vs. indicated mean effective pressure. 

The tests have shown an important decrease in emissions of CO and HC for 

operation with E20 gasoline-ethanol blend (Fig. 4 and Fig. 5). This is mainly due to 

better combustion properties. The level of these emissions increases at the enrichment 

of the dosage for both fuels and engines admission type (normal intake and 

supercharged). 

Fig. 4 – Comparison of specific HC emission for gasoline and E20. 

The NOx emissions presents a significant increase for E20 fuelled engine 

mainly due to higher combustion temperature. It can be seen that NOx emission 

level is higher for the supercharged engine model at the excesses air-fuel ratio 

value and decreases at poorer dosages.


Fig. 5 – Comparison of specific CO emission for gasoline and E20. 

Fig. 6 – Comparison of specific NOx emission for gasoline and E20. 


1. Results have shown that the model developed in AVL BOOST has 

successfully predicted the trends of engine performance, fuel consumption and 

various exhaust emissions for gasoline and E20 fuels. 

2. An improved engine performance was obtained at fuelling with 

ethanol-gasoline blend due to the ethanol properties which allowed engine 

operation at leaner dosages than in the case of engine working with gasoline. 

3. The use of ethanol-gasoline blends leads to an important decrease of 

HC and CO emissions while NOx emissions level increased but at use of the 

leaner dosages (λ>1.17) the NOx emissions level is smaller than standard engine.


4. The best results of the investigated parameters were obtained on the 

turbocharged engine. 

Acknowledgements. The work has been funded by the Sectorial Operational 

Programme Human Resources Development 2007-2013 of the Romanian Ministry of 

Labour, Family and Social Protection through the Financial Agreement 

POSDRU/88/1.5/S/60203. The authors would like to thank AVL GMBH Graz Austria 

for providing the software AVL Boost v.2011.6 used in this study. 

REFERENCES 

*** http://www.txideafarm.com/ethanol_fuel_properties_and_data 

Alok Kumar, Khatri D.S., Babu M.K.G., An Investigation of Potential and Challenges 

with Higher Ethanol-gasoline Blend on a Single Cylinder Spark Ignition Research 

Engine. SAE Paper, 2009-01-0137 (2009) 

Bahattin Celik M., Experimental Determination of Suitable Ethanol–gasoline Blend 

Rateat High Compression Ratio for Gasoline Engine. Applied Thermal 

Engineering, 28, 396-404 (2008). 

Coelho Baeta J.G., Amorim R.J., Molina Valle R.J., Mautone Barros E., Bahia de 

Carvalho R.D., Multi-Fuel Spark Ignition Engine – Optimization Performance 

Analysis, SAE Paper 2005-01-4145 (2005). 

Cowart J.S., Boruta W.E., et al., Powertrain Development of the 1996 Ford Flexible 

Fuel Taurus, SAE Paper 952751 (1996). 

Heywood John B., Internal Combustion Engine Fundamentals. 1998. 

Kenjiro Nakama, Jin Kusaka, Yasuhiro Daisho, Effect of Ethanol on Knock in Spark 

Ignition Gasoline Engines. SAE Paper 2008-32-0020 (2008). 

Ku J., Huang Y., Hollowell B., Belle S., Matthews R., Hall M., Conversion of a 1999 

Silverado to Dedicated E85 with Emphasis on Cold Start and Cold Driveability. 

SAE Paper 2000-01-0590 (2000). 

Merkourios G., Savvidis D., Triandafyllis J., Study of the Effects of Ethanol Use on a 

Ford Escort Fitted with an Old Technology Engine. SAE Paper 2008-01-2608 

(2008). 

Nakata K., Utsumi S., The Effect of Ethanol Fuel on Spark Ignition Engine. SAE Paper 

2006-01-3380 (2006) 

Pană C., Negurescu N., Popa M. G., Cernat A., Soare D., Some Experimental Aspects of 

the Ethanol Use in SI Engine. Automobile, Environment and Farm Machinery, 

The 1th International Congress, Tehnical University of Cluj-Napoca (2007). 

Radu A., Pana C., Theoretical Research of Bioethanol Use in S.I. Engine. Conference 

Proceedings of the Academy of Romanian Scientists, 3, 1, 177 (2011). 

Turner J. W. G., Peck A., Pearson R. J., Flex-Fuel Vehicle Development to Promote 

Synthetic Alcohols as the Basis of a Potential Negative CO2 Energy Economy, 

SAE Paper 2007-01-3618 (2007). 

Wen Dai, Sreeni Cheemalamarri, Curtis E.W., Riadh Boussarsar, Morton R.K., Engine 

Cycle Simulation of Ethanol and Gasoline Blends. SAE Paper, 2003-01-3093 

(2003).


ASPECTE PRIVIND UTILIZAREA BIOETANOLULUI ÎN MOTOARE CU 

APRINDERE PRIN SCÂNTEIE 

(Rezumat) 

În ultimii ani, motoarele cu aprindere prin scânteie au fost îmbunătățite 

încontinuu în vederea reducerii emisiilor de dioxid de carbon (şi totodată, efectul de 

seră), a consumului de combustibil şi creşterea performanţelor motorului. Utilizarea 

combustibililor alternativi reprezintă o soluţie eficientă pentru aceste obiective. Astfel 

se explică interesul mare la nivel mondial de a utiliza bioetanolul (provenind din surse 

regenerabile), drept combustibil auto, în special în amestec cu benzina. Lucrarea de faţă 

prezintă rezultatele simulării proceselor termo-gazo-dinamice din interiorul cilindrului 

unui motor cu aprindere prin scânteie alimentat cu amestecuri benzină-etanol la diferite 

condiţii de funcţionare ale motorului. Principalele subiecte abordate sunt efectele 

utilizării amestecurilor asupra performanţei motorului, emisiilor şi consumului de 

combustibil. Rezultatele au arătat că o performanţă îmbunătăţită a motorului şi emisii 

mai reduse pot fi obţinute prin supraalimentarea motorului şi utilizarea amestecurilor 

benzină-etanol, datorită proprietăţilor de ardere mai bune ale etanolului.


Publicat de 



Secţia 


EMISSION LEVEL FROM INTERNAL COMBUSTION ENGINE 

USING FOSIL AND ALTERNATIVE FUELS 

BY 

IOAN HITICAS ∗ , LIVIU MIHON, DANILA IORGA and WALTER SVOBODA 



Politehnica University of Timişoara, 

MMUT Department 

Abstract. The start of the propulsion engines, coming as a result of 

technological evolution, revealed quickly different problems like comfort, noise, 

performances and, not least, exhaust gases. Appearance of compression ignition 

engines and spark ignition engine, technology that replaces the animal traction, 

have been received very well by engineering society of that period, but the 

problem of exhaust gases was and remains an issue to be solved. The vehicle 

emit different pollutant elements through the tailpipe, like carbon dioxide (CO2), 

carbon monoxide (CO), hydrocarbon (HC), nitrogen oxides (NOx), volatile 

organic compounds (VOC’s), particulate matter (PM) and other elements, as a 

consequence of burning the fossil fuels inside the combustion chamber. The 

atmosphere is continuously changing due to all this pollutant elements, 

producing damages on human and earth health. The paper present the studies and 

experimental research concerning the nature of exhaust gases, for vehicles 

equipped with internal combustion engine, realized in “Politehnica” University 

of Timisoara, Faculty of Mechanics, taking into account the figures of exhaust 

gases for vehicle functioning with diesel as fuel, and vehicle with spark ignition 

engine using gasoline and alternative fuel as energy source. The conclusion we 

reached are: diesel engines emit less light hydrocarbons than gasoline engines, 

and the spark ignition engine working with alternative fuels – namely isobutene 

(IP50), emit less nitrogen oxides than gasoline as fuel. 

Key words: CO2, pollutant emission, spark ignition engine, diesel engine, 

alternative fuels. 

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

102 Ioan Hiticas et al. 


Internal combustion engine remains a subject who must be treated due to 

the elements involved in the combustion process. In this paper are treated the 

exhaust problems, analysing the processes in diesel engine and spark ignition 

engine. The oil crisis around the 1970, when along the USA many fuel stations 

were closed due to this crisis, conducted to less pollutant emission by vehicle 

fleet, and also conducted the engineering to find other sources solution for 

vehicle, like alternative fuels or electric vehicles. 

The CO2 emissions over the world are structured after the quantities 

emitted by”departments”, like vehicles, industry, manufacturing, agriculture, 

and other. The first three countries group on IEA report (International Energy 

Agency) on CO2 emission, in billions of metric tons, for year 2010, was: 

Europe, USA and Australia – over 15 billions of metric tons, China – over 8.1 

billions of metric tons, and India – over 1.8 billions of metric tons. The 

reference was made for year 2010 because this year was a record for greenhouse 

gas emission. Globally, CO2 emission from burning fossil fuel achieved a 

record of 30.6 billion metric tons. UNFCCC (United Nations Framework 

Convention on Climate Change) presented a national emission report on the 

Europe greenhouse gases and the data for Romania are: CO2 emission on Fuel 

Combustion sector on year 2010 was 75499.84204 Giga tonnes and CO2 

emission for Road Transportation sector was 13498.11400 Giga tonnes. This 

numbers must prevent us that the global problems of greenhouse effect have to 

be solved and should not be delayed. 

Due to the high level of CO2 emission around the world, presented in Fig. 

1, situation made for year 2009, and also harmful emission, has made the 

countries to applied legislation to reduce them, like Euro I to Euro VI. 

Fig. 1 – World CO2 emission by sector (IEA, 2011). 

The experimental researches conducted in University Politehnica of 

Timisoara and presented in this paper, confirm the danger, on human and earth 

health, of CO2 emission level, and harmful emission, produced by thermal 

engine.


2. Technical Data 

The internal combustion engine, as we know, transforms the heat 

produced by fuels burned inside the combustion chamber, in work. But the 

process is not ending here. How we obtain the work and how the problem of 

exhaust gases is somehow solved, we won’t treat in this paper. The oil resources 

will end and the renewable energy will take place the fossil fuels. Until than we 

present technical data about the evolution of exhaust gases from diesel engine 

and spark ignition engine. 

We can calculate the CO2 emission from the combustion of fossil fuels 

along one year or more years, following 

PE = ∑ FC 

Coef , 

FC, j, y i, j, y i, y 

i 

where: PEFC,j,y [tonnes CO2/year] – CO2 emission from fossil fuels burned in j 

process during the year y; FCi,j,y [mass/year] – quantities of fuel type burned in j 

process during the year y; Coefi,y [CO2 tonnes/mass] – CO2 emission coefficient 

of fuel type i in year y; i – The fuel type burned in process j during the year y. 

Taking into account the urban transportation in Timişoara, treated also in 

this paper for the exhaust gases aspects, we can obtain the traffic capacity by 

calculus, using the velocity map, υ, 

⎛ F ⎞ 

υ =f( γ,ξ,δ) and γ= g ⎜ ⎟, 

(2) 

⎝ K( φ) 

⎠ 

where: γ – capacity utilization; F – circulation; K(φ) – effective capacity; ξ – 

percentage of freight vehicles. 

This calculus allowed estimating the traffic capacity, which is a reference 

to us to estimate the CO2 emission. 

Following, we present a model of calculating the CO2 emission for one 

vehicle, the Opel/Vauxhal vehicle, with the following technical data: engine 

capacity 1248 cm 3 (1.3 CDTi), 5 door, MPV, manual transmission M5, 

manufacturing year 2008, fuel type – diesel, 120 g CO2/km. This calculation 

was made for a yearly normal exploitation of 20000 km. Thus 

20000 · 120/10 6 = 2.4 tonnes CO2. 

After this simple calculus, we can estimate the mass of CO2 emission for 

vehicles working with fossil fuel, for a region, country or continent (Höglund & 

Niittymäki, 1999). Taking into account the city Timişoara, with over 300000 

inhabitants and with almost the same number of vehicles, we can obtain the 

estimate value for CO2 emission, keeping the yearly normal exploitation. 

2.4 tonnes CO2 · 300000 vehicles = 720000 tonnes CO2. 

(1)


3. Experimental Research 

The experimental research was conducted in vehicle dynamics laboratory 

from Faculty of Mechanics, Timişoara. For the experiments we analysed two 

engine types: first type was the diesel engine, which equipped the urban 

vehicles, and a second engine, a spark ignition engine, which equipped a 

passenger car. 

In this paper we present the real situation of diesel engine which 

equipped the vehicle in urban transportation (Koerner & Klopatek, 2002), and 

we monitored the exhaust gases to identify the principal harmful elements, with 

adverse health effect. 

In compression ignition engines case, we realised measurement for 

particulate matter and opacity. Vehicles used for experimental research was an 

urban bus, Mercedes, model which fit almost the entire urban city fleet. The 

vehicles are equipped with diesel engines (Wallington T.J. et al., 2008); model 

MO45LhA, 11 litres displacement, and supercharged, nominal power 183 

kW/2100 rpm, and emission standard were Euro III. Usually a diesel engine 

emitted through tailpipe different harmful species, like HC, NOx, and others, 

including particulate matter which composed the soot. Another reference on this 

engine is the degree of smoke and the opacity, which is in correlation with 

linear absorption coefficient, k, measured through opacimeter. 

Tested vehicle have run a travel distance of 3500 km, divided in five 

segments. First segment was 1323 km distance, followed by other 4 segments, 

presented in Table 1. For the first segment, we make measurements with the 

opacimeter, having k = 0.23 m -1 , or Hartridge Standard Unit, HSU = 0.90 %, 

resulting PT (particulate matter) = 0.05 [g/m 3 ], for this first segment. 

Table 1 

Evolution of particulate matter and opacity 

Nr. Travel distance, km k, m -1 HSU, % PT, g/m 3 

1 1323 0.23 0.90 0.05 

2 1741 0.31 0.86 0.07 

3 2322 0.44 0.81 0.09 

4 2936 0.26 0.88 0.06 

5 3465 0.38 0.83 0.08 

The opacity is expressed through linear absorption coefficient, K, through 

the equation 

1 1−N 

k = ln , 

(3) 

L 100 

where: k – Linear absorption coefficient; L – Effective length of a column of 

homogeneous gas, m.


Below, in Fig. 2 and Fig 3, are presented the evolution of particulate 

matter and the opacity measured for the Mercedes vehicles, made in University 

Politehnica of Timişoara and the graphics present us values which ensure a 

function of the vehicles in good parameters, heaving no problems with pollution 

standards. 

Fig. 2 – Particulate matter evolution. 

Fig. 3 – Opacity evolution. 

In the second case, for the spark ignition engine, we monitor the 

evolution of exhausts, using as energy sources an alternative fuel mixture, 

namely isobutene, in proportion: 50% gasoline and 50% isobutene, called 

further IP50. We choose this alternative fuel due to his advantages: is a 

superior fuel comparing with gasoline, due to his thermal efficiency, applied 

especially for high compression engine, but not only. The experimental research 

was made for an Opel Omega A, the data given in Table 1: 

Maximum 

power, kW 

85 at 5200 

rpm 

Maximum 

speed [rpm] 

6400 

Table 1 

Opel Omega A data [Hiticas et al., 2012] 

Maximum Displacement 

torque, Nm cm 3 

Diameter x Compression 

Stroke mm ratio 

170 at 260 

rpm 

1998 86 X86 9.2 

Supply Fuel pressure 

bar 

Firing order Exhaust control 

Bosch 

Oxygen probe, 

Motronic 1.5 2.5…3 on the 1-3-4-2 catalytic 

Injection ramp injector 

convertor 

The tests were obtained on the Maha LPS 3000 dyno, in the Faculty of 

Mechanical Engineering, involving several elements, like user interface, the 

remote control, roller, temperature, pressure and humidity module, and another 

module for simulating the air flow. The AVL DiCom 4000 gas analyser was


used to monitor the exhaust species and amounts during the measurements on 

Opel Omega vehicle, using IP50 as fuel mixture. The elements monitored were 

CO, CO2, O2, NOx, HC, and, based on this elements (Crass, 2008), we can 

calculate the excess air ratio, λ. 

After the necessary adjustments made to prepare the vehicle for 

measurements (Bosch, 2004; Atkins, 2009) we analysed the exhaust gases in 

two cases: for partial load and for full load. The graphics present the following 

curves: the reference curves, which are for 100% gasoline used as fuel, the 

second curves are for engine adaptation, and the last curves for engine 

functioning with fuel mixture, IP50. 

Fig. 4 – CO emission at full load. 

Fig. 5 – HC emission at full load 

[Hiticas et al., 2012]. 

Fig. 5 – CO2 emission at partial load. 

Fig. 6 – NOx evolution at full load 

[Hiticas et al., 2012]. 

It can be observed from graphics that the fuel mixture present advantages 

and also disadvantages. After our experimental research we conclude that this


alternative fuel mixture can be a future solution applied to vehicle production, 

because the hydrocarbon presents very good values, but also, the nitrogen 

oxides must be corrected. 


1. The conclusion we reached after the studies and experimental research 

involves following elements: greenhouse gas emission from thermal engine 

remain a problem. The vehicle equipped with diesel or spark ignition engine 

emitted harmful elements with consequences on human and nature health. 

2. In first case, the diesel engine, we monitor the particulate matter and 

the opacity along a period and distance, we conclude that the urban Mercedes 

bus, which is part of the urban fleet, present normal values, comparing the 

values of the same parameters for other vehicle, also present in urban fleet. 

3. In the second case, the spark ignition engine, we monitor the exhaust 

gases evolution when the engine functioning with fuel mixture, IP50. We 

achieved that the engine perform well, and HC present acceptable values, 

comparing with functioning with gasoline, but the NOx must be corrected, due 

to the higher values than functioning with gasoline. 

4. As future research, the function of spark ignition engine with fuel 

mixture trough alternative fuel, must present better values for nitrogen oxide. 

Acknowledgements. This work was partially supported by the strategic grant 

POSDRU/88/1.5/S/50783, Project ID 50783 (2009) co-financed by the European Social 

Fund – Investing in People, within the sectarial Operational Programme Human 

Resources Development 2007-2013. 

This work was also partially supported by the strategic grant 

POSDRU/21/1.5/G/13798, inside POSDRU Romania 2007-2013, co-financed by the 

European Social Fund – Investing in People. 

REFERENCES 

Koerner B., Klopatek J., Anthropogenic and Natural CO2 Emission Sources in an Arid 

Urban Environment. Elsevier, Environmental Pollution, 116, S45–S51 (2002). 

Wallington T.J., Sullivan J. L., Hurley M. D., Emissions of CO2, CO, NOx, HC, PM, 

HFC-134a, N2 O and CH4 from the Global Light Duty Vehicle Fleet. 

Meteorologische Zeitschrift, 17, 2, 109-116 (2008). 

Hiticas I., Marin D., Mihon L., Experimental Research Concerning the Pollution of an 

Internal Combustion Engine with Injection of Gasoline, in Conditions of 

Changing the Fuel. IN-TECH, Croatia, 2012 (in print). 

Höglund P.G., Niittymäki J., Estimating Vehicle Emissions and Air Pollution related to 

Driving Patterns and Traffic Calming. Urban Transport Systems, Lund, Sweden, 

1999.


Khan S. et al, Idle Emissions from Heavy-Duty Diesel Vehicles: Review and Recent 

Data. Journal of the Air & Waste Management Association, 56, 1404-1419 

(2006). 

*** Tracking Progress in Carbon Capture and Storage, International Energy Agency, 

(April 2012). 

*** Tool to Calculate Project or Leakage CO2 Emissions from Fossil Fuel Combustion. 

Methodological tool, UNFCCC / CCNUCC Report (August 2008) 

*** CO2 Emission from Fuel Combustion. Highlight, International Energy, (2011). 

Crass M., Reducing CO2 Emissions from Urban Travel: Local Policies and National 

Plans. Proceedings of OECD International Conference, Competitive Cities and 

Climate Change, Milan, Italy, October 2008. 

Stoicescu A.P., Proiectarea performanţelor de tracţiune şi de consum ale 

automobilelor. Ed. Tehnică, Bucureşti, 2007. 

Blair G. P., Design and Simulation of Four-Stroke Engines. SAE International, USA, 

1999. 

Raţiu S., Mihon L., Motoare cu ardere internă pentru autovehicule rutiere – Procese şi 

caracteristici. Ed. Mirton, Timişoara, 2008. 

Bosch R., Gasoline-Engine Management, Handbook, 2 nd Edition, Germany, 2004. 

Atkins R. D., An Introduction to Engine Testing and Development. SAE International, 

USA, 2009. 

Tokar A., Cercetări privind interacţiunea dintre automobilul echipat cu motor cu 

ardere internă şi mediu. Ed. Politehnica, Timişoara, 2009. 

NIVELUL EMISIILOR POLUANTE ALE MOTOARELOR CU ARDERE INTERNĂ 

CARE UTILIZEAZĂ COMBUSTIBILI CLASICI ŞI ALTERNATIVI 

(Rezumat) 

Acest articol prezintă studii şi cercetări experimentale privind autovehiculele 

echipate cu motoare cu aprindere prin scânteie şi prin comprimare. Experimentele au 

fost realizate în cadrul Universităţii Politehnica din Timişoara, Facultatea de Mecanică, 

analizând evoluţia gazelor de evacuare la funcţionarea cu motorină, pentru motoarele 

diesel, cât şi la funcţionarea cu benzină, dar şi cu un combustibil alternativ (isobutanol), 

în amestec în proporţie de 50% cu 50% benzină, în cazul motoarelor cu aprindere prin 

scânteie. Concluziile la care s-a ajuns sunt: pentru motoarele monitorizate de către 

autori şi care echipează o bună parte din parcul auto al transportului urban din 

Timişoara, emit mai puţine hidrocarburi decât motoarele cu aprindere prin scânteie, iar 

acestea din urmă, la funcţionarea cu amestecul alternativ, emit mai puţini oxizi de azot 

decât la funcţionarea numai cu benzină.


Publicat de 



Secţia 


THEORETICAL AND EXPERIMENTAL INVESTIGATIONS OF 

THE SI ENGINE TURBOCHARGING 

BY 

COSTIN DRAGOMIR, CONSTANTIN PANĂ * , NICULAE NEGURESCU 

and ALEXANDRU CERNAT 



“Politehnica” University of Bucharest 

Abstract. The spark ignition engine turbocharging is a efficient method of 

great actuality for their brake mean effective pressure and of the specific power 

increasing (downsizing concept) and for the pollution decreasing. The method 

can be applied at the SI engines with an efficient control of their operation for to 

avoid the knock appearance. The theoretical and experimental investigations 

were performed on a turbocharged spark ignition engine with 1.5 l displacement 

and fuel MP injection. 

The paper presents the influences air supercharging pressure upon the 

engine performances. In paper is determined a optimum correlation between 

dosage, spark ignition advance, air supercharging pressure, exhaust gas 

temperature, brake mean effective pressure, brake specific fuel consumption. 

Key words: engine, supercharging, downsizing, downspeeding. 


Significant increasing in power per litter for spark ignition engine can be 

assured by supercharging, which allows the increases of indicated mean 

effective pressure. Downsizing and downspeeding application for spark ignition 

engines represent modern concepts for lower displacement engines development 

(compact gauge and low costs), with a lower speed for maximum power/ 

maximum torque comparative to aspirated engines (for the same or higher 

power/torque values), with favourable effects on engine thermal and mechanical 

stress, efficiency, pollutant emissions and wear. Supercharging was considered 

______________________________ 

*Corresponding author: e-mail: constantin.pana@mail.upb.ro

110 Costin Dragomir et al. 

a common method used for IMEP increasing only for diesel engine. The main 

issues of spark ignition engine supercharging are represented by: knocking 

phenomena may appears; exhaust gases temperature increases; engine thermalmechanical 

stress increases 

Nowadays modern management of SIE running allows the supercharging 

also for spark ignition engines. Thus, an optimal correlation between 

supercharging pressure -compressor exhaust air temperature- compression ratiospark 

ignition timing-dosage-exhaust gases temperature can assure the spark 

ignition engine operation without knocking combustion and with remarkable 

energetically and polluting performances. During the last years the 

supercharging of SIE was the most efficient method of increasing performances 

from the point of view of energetically and polluting terms. 

Thus, in Table 1 some of supercharged spark ignition engines 

characteristics are presented. 

Engine 

Displacement 

Table 1 

Compression 

ratio 

Pmax/nPmax 

[kW/rpm] 

Mmax/nMmax [Nm/rpm] 

Uni Melb 

WATTARD 

0,43 

9-13:1 

variable 

53/9000 65 / 7000 

Audi 1.8 9.8 :1 125/5 900 225/1950-5000 

VW 1.984 10.5 :1 147/5.700 280/1800-4700 

VW 1.4 10 :1 90/5.000-5.500 200/1.500-4.000 

Audi Q3 1.8 9.6 :1 125/4300-6000 280/1700-4200 

Skoda FW 1.2 40/4750 105/3000 

BMW 2.979 10.2 :1 225/5800 400/1200-5000 

VW 2.0 FSI 1.984 10.5 :1 147/5700 280/1800-4700 

VW1.2MPI 1.2 10.3 :1 44/5200 108/3000 

Renault 1.2 10 :1 85/4500 190/2000-4000 

MAHLE 

Downsizing Engine 

1.2 9.75:1 144/6500 285/2500-3000 

2. Engine In-cylinder Thermo-gas Dynamic Processes Simulation 

Engine in-cylinder processes simulation was developed for 1.5 litter 

DAEWOO engine. In order to define the energetically performances and the 

cycle performances parameters, a zero dimensional physic-mathematical model 

was developed. The model uses a Vibe combustion formal law and takes into 

consideration the heat transferred to the walls. 

For knock avoiding, the combustion duration was established shorter than 

end-gas auto ignition delay, parameters being evaluated by Douaud and Evzat 

equation. For program calibration the experimental investigation results were 

used and the considered model hypotheses were verified. 

In order to determinate the polluting emission level for the simulated 

regimes the AVL Boost program was used. Modelling process was developed


for the aspirated engine and for the supercharged engine designed at built in the 

laboratories of the Department of thermotechnics, engines, thermal equipments 

and refrigeration installations from University Politehnica of Bucharest. 

Different supercharging pressures, ps- absolute pressure and temperatures of the 

blower exhaust cooled air were taken into consideration. For knock avoiding the 

reach dosages in the area of 0.8….0.9 and cooling of the inlet air were used. 

For each regime, the spark ignition timing was set up for efficiency and knock 

avoiding. Modelling results for full load and speed of 4800 rpm are shown in 

Fig. 1-6. 

p max [bar] 

100 

80 

60 

40 

calculated 

20 

1 1.2 1.4 1.6 1.8 2 

ps [bar] 

measured 

Fig.1 – Maximum pressure versus ps. 

p e [bar] 

NOx [ppm] 

20 

15 

10 

5 

1 1.2 1.4 1.6 1.8 2 

500 

400 

300 

200 

100 

calculated 

measured 

ps [bar] 

Fig. 3 – BMEP versus ps. 

calculated 

0 

1 1.2 1.4 1.6 1.8 2 

p s [bar] 

measured 

Fig.5 –NOx versus ps. 

c e [g /kW h ] 

CO [%] 

400 

350 

300 

250 

HC [ppm ] 

10 

8 

6 

4 

2 

measured 

calculated 

1 1.2 1.4 1.6 1.8 2 

ps [bar] 

Fig.2 – BSFC(ci) versus ps. 

calculated 

0 

1 1.2 1.4 1.6 1.8 2 

400 

300 

200 

p s [bar] 

measured 

Fig.4 –CO versus ps. 

measured 

calculated 

100 

1 1.2 1.4 1.6 1.8 2 

p s [bar] 

Fig. 6 –HC versus ps. 

3. Experimental Investigations and Results 

The experimental researches were carried on an automotive SI engine, 

type DAEWOO –1.5 l, at different engine operating regimens. The engine was 

mounted on a test bench (Fig. 7) equipped with the next necessary instruments


for measuring operations: AVL ALPHA 160 eddy current dynamometer 

equipped with throttle actuator that work in parallel with the dyno in order to 

operates the control lever of the injection pump, real time AVL data acquisition 

system for processing and storage of measured data, AVL in-cylinder pressure 

transducer line, AVL gas analyzer, Khrone Optimass mass flow meter, engine 

inlet air flow meter, thermo resistances for engine cooling liquid temperature, 

engine oil and air intake temperatures and thermocouples for exhaust gas 

temperature, manometer for air pressure from engine intake manifold. All 

instrumentation was calibrated prior to engine testing. Experimental research’s 

carried out to obtain fuel consumption characteristics for different speeds and 

engine full load and to determinate the energetically and pollutant engine 

performance. Based on fuel consumption characteristics were obtained some 

graphic representations for different parameters such as: brake mean effective 

pressure (BMEP), brake specific fuel consumption (BSFC), pollutants 

emissions level (HC, CO2, and NOx) versus to absolute pressure ps. 

Fig. 7- Test bed scheme: AGE exhaust gas analyzer ; AS charge amplifier; B battery; ca 

intake manifold; CAD data acquisition computer; ce intake manifold; CF dyno power 

cell; CG fuelling system computer; CIG injectors actuation; ct three way catalyst; DA 

air flowmeter; DC fuel flowmeter; EGR exhaust gas recirculation valve; F eddy current 

dyno; FC fuel filter; IND Indimodul 621 data acquisition unit; IT temperature 

indicators; M Daewoo 1.5 spark ignition engine; MP supercharging pressure 

manometer; ORS throttle; PCF dyno command panel ; R engine cooler; RI intercooler; 

RC fuel reservoir; SA power supply; SCP throttle actuator servomotor; SRAF dyno 

cooling system; st gas analyzer speed sensor; TC turbocompressor; tf dyno cooling 

water temperature sensor; TF dyno speed transducer; tp cylinder pressure transducer; 

TPU angle encoder; UCP principal electronic control unit; UCS secondary electronic 

control unit; UPF dyno power unit; UPS throttle actuator servomotor power unit; VE 

cooling electric fan for intercooler; x electronic emitter-receptor.


4. Theoretical and Experimental Investigations Results 

The results of theoretical and experimental investigations presented in 

Fig. 1-6, show good correlation between them. 

Supercharging pressures used are in the range of 1.4…1.8 bar. 

Comparative to the aspirated engine, the maximum pressure increases with 

almost 100 % for a supercharging pressure (ps) of 1.8 bar, Fig. 1. High value of 

maximum pressure leads to the limitation of supercharging pressure at 1.4 bar 

when the increase of maximum pressure value (~ with 50 % comparative to 

classic solution) is acceptable for the engine reliability. 

In order to avoid the knocking and to limit the maximum pressure, the 

ignition angle value optimization was achieved in such a way that the BSFC is 

maintained almost constant for different supercharging pressures (Fig. 2). 

For 1.4 bar supercharge pressure the increase BMEP is significant 

(~22%) and increase is more at much higher supercharging pressure because of 

the influence of ignition timing value (Fig. 3). Also from this reason the 

limitation of the supercharging pressure at 1.4 bar is justified. 

Due to combustion process improvement the level of CO and HC 

emissions decreased (~24% for CO and ~11% for HC) (Fig. 4 and Fig. 6). The 

NOx emission level decreases significant for 1.4 bar supercharging pressure 

(Fig. 5) fact also influenced by the reduction of spark ignition timing. 

Supercharging represents an efficient method of engine downsizing. 

Thus, in order to obtain the same power as the reference engine, the engine 

displacement must be reduced by 1.5 times for 1.4 supercharging pressure. This 

engine displacement reduction can be assured by the limitation of the engine 

cylinders number, which also leads to the increase of the engine mechanical 

efficiency. Another advantage of supercharging is the reduction of the 

maximum power/maximum torque speeds – downspeeding concept. Thus, the 

same power of 66 kW at 4800 rpm of the normal aspirated engine was obtained 

for supercharging at 1.4 bar at 3900 rpm, and the maximum torque of 137 

Nm/3600 rpm was reached at the speed of 2500 rpm. The reduction of the fuel 

consumption and engine wear are the main advantages of this concept. 

Spark ignition timing modification directly affects the combustion 

process variability evaluated by coefficient of variability in maximum pressure. 

For 4800 rpm the coefficient of variability in maximum pressure reach the value 

of (COV) = 6.574% for λ= 0.85 and 1.4 bar supercharging pressure. 

p max 

The influences of combustion process on engine running are reflected by 

the values of the coefficients of cycle variability in indicated mean effective 

pressure. The values of variability coefficient in IMEP are (COV)IMEP=1.986% 

for λ= 0.85 at ps = 1.4 bar.



1. The theoretical and experimental investigations results allow 

formulating the following conclusions: 

2. SI engine supercharging is a method to obtain efficiency and specific 

power/torque performance increasing. 

3. The pollutant emissions level decreases due to the improvement of the 

combustion processes. 

4. Knock is the most important limiting factor of the supercharging 

engine. An optimum correlation establish between dosage, spark ignition 

advance, air boost pressure, air boost temperature, exhaust gas temperature, 

brake mean effective pressure, brake specific fuel consumption leads to the 

avoiding of knocking phenomena. 

5. Supercharging represents an efficient method of engine downsizing 

and downspeeding. 

Acknowledgments. The authors would like to thank to AVL List GmbH Graz, 

Austria, for providing the possibility to use the Simulation Software AVL BOOST 

REFERENCES 

Attard W., Watson H.C., Konidaris S., Khan M.A., Comparing the Performance and 

Limitations of a Downsized Formula SAE Engine in Normally Aspirated, 

Supercharged and Turbocharged Modes. University of Melbourne, SAE Paper 

2006-32-0072, in SAE International (2006). 

Attard W.A., Toulson E., Hamori F., Watson H.C., Combustion System Development 

and Analysis of a Downsized Highly Turbocharged PFI Small Engine. SAE Paper 

2009-32-0185/20097185, in SAE Japan and SAE International (2009). 

Attard W.P., Konidaris S., Toulson E., Watson W.C., The Feasibility of Downsizing a 

1.25 Liter Normally Aspirated Engine to a 0.43 Liter Highly Turbocharged 

Engine, University of Melbourne, SAE Paper 2007-24-0083, in SAE International 

(2007). 

Basshuysen R., Gasoline Engine with Direct Injection, Processes, Systems, 

Development, Potential. Vieweg+Teubener, GWV Fachverlage Gmbh, 

Wiesbaden, Germany, 2008. 

Basshuysen R., Ottomotor mit Direkteinspritzung, Verfahren, Systeme, Entwicklung, 

Potenzial. 2, Auflage. Vieweg+Teubener, GWV Fachverlage Gmbh, Weisbaden 

Germany, 2008. 

Becker N., Der Neue 1,0-L- Dreizilinder-MPI- Motor Für Den UP! .ATZ Extra 

September 2011, Germany, pp.36-43. 

Blaxill H., The Role Of IC Engines In Future Energy Use. MAHLE Powertrain, 

Germany, 2011. 

Eiser A., Deblaze M., Freidmann K., Kerschenlohr A., Schachner M.(2011), Der 

Antrieb, Der neue Audi Q3. ATZ Extra Juli 2011, Germany, pp. 24-31. 

Halder J., Dreizylindermotoren von Volkswagen. MTZ 05/2009, Germany, 2009, pp. 

354- 360.


Heywood J.B., Symposium University of Madison, Wisconsin, Engine Research 

Center, 2005. 

Heywood J.B., Internal Combustion Engines Fundamentals, McGraw-Hill Book 

Company, 1988. 

Kirwan J.E., Mark Shost M., Roth G., Zizelman J., 3-Cylinder Turbocharged Gasoline 

Direct Injection: A High Value Solution for Low CO2 and NOx Emission,s SAE 

Paper 2010-01-0590, Publishe 04/12/2010, in SAE International (2010). 

Klauer N., Kretschmer J., Unger H., Der Antreb des BMW 535i Gran Turismo. ATZ 

09/September 2009, Germany, 2009, pp. 610-617. 

Korte V., Hancock D., Blaxill H., Downsizing-Motor von Mahle als 

Technologiedemonstrator, konzept, Auslegung und Konstruktion, MTZ 01/2009, 

Germany, 2009, pp. 10 -19. 

Navrátil J., Polášek M., Vítek O., Macek J.,Baumruk P., Simulation of Supercharged 

and Turbocharged Small Spark-Ignition Engine. 3 th. International Colloquium 

MECCA, Czech, 2003, pp. 27-33. 

Pană C., Negurescu N., Popa M.G., Cernat A., Investigations Regarding the Use of the 

Turbocharging and Bioethanol at SI Engines. Bul. Inst. Polit. Iasi, LXI(LX), 4, s. 

Construcţii de maşini (2010). 

Pfalzgraf B., Fitzen M.,Siebler J., Erdmann H.-D., First ULEV Turbo Gasoline Engine- 

The Audi 1.8 l 125kW 5-Valve Turbo. SAE Paper 2001-01-1350, SAE 

International. 

Stephenson M., MAHLE Powertrain, Engine Downsizing - An Analysis Perspective. 

SIMULIA Conference, Germany, 2009. 

INVESTIGAŢII TEORETICE ŞI EXPERIMENTALE ALE MOTORULUI 

CU APRINDERE PRIN SCÂNTEIE SUPRAALIMENTAT 

(Rezumat) 

Supraalimentarea motoarelor cu aprindere prin scânteie este o metodă eficientă 

de mare actualitate pentru creşterea presiunii medii efective şi a puterii litrice a lor 

(conceptul downsizing) şi pentru reducerea emisiilor poluante. Metoda poate fi aplicată 

la motoarele cu aprindere prin scânteie cu un management eficient a funcţionării lor în 

vederea evitării apariţiei fenomenului de ardere cu detonaţie. 

Investigaţiile teoretice şi experimentale au fost efectuate pe un motor cu 

aprindere prin scânteie cu cilindreea de 1,5 l şi cu injecţie multipunct transformat dintrun 

motor de serie cu admisie normală. 

Lucrarea prezintă influenţe ale presiunii de supraalimentare asupra 

performanţelor motorului. 

În lucrare a fost stabilită o corelaţie optimă între dozaj, avans la declanşarea 

scânteii electrice, presiune de supraalimentare, temperatura gazelor de evacuare, 

presiune medie efectivă şi consum specific efectiv de combustibil pentru un control 

eficient al funcţionării motorului cu aprindere prin scânteie supraalimentat.


Publicat de 



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INVESTIGATION THE EFFECT OF HYDROGEN ADDITION IN 

A SPARK IGNITION ENGINE 

BY 

EUGEN RUSU ∗ , CONSTANTIN PANĂ and NICULAE NEGURESCU 

“Politehnica” University Bucureşti, 

Faculty of Mechanic Engineering and Mecatronics 


Accepted for publication: 13 March 2012 

Abstract. The main concern of the automotive industry at this moment is to 

reduce pollutant emissions, fuel consumption, and also increasing the engine 

performance. Due to its excellent combustion properties hydrogen can be used as 

addition for the improving combustion in the gasoline-fuelled spark ignition (SI) 

engines. Beside experimental research thermo-gas-dynamic simulation models 

of spark ignition (SI) engines are one of the most effective tools for the analysis 

of engine performance, parametric examinations and assistance to new 

developments. 

The paper presents results of modelling thermo-gas-dynamic processes of 

the spark ignition engine fuelled with hydrogen in addition. The research has 

focused on emissions (NOx, CO, CO2), fuel consumption, and engine 

performance. 

Key words: internal combustion engine, hydrogen addition, simulation, 

fuel properties, performance 


During the last years SI-engines have been further improved. Engine 

development has focused on the reduction of tailpipe emissions, better fuel 

economy and higher engine performance as well as reduction of system costs. 

The legislation for Particulate Matter is particularly challenging, both in 

terms of measurements and in ensuring conformity to legislation. The restriction 

∗ Corresponding author: e-mail: eugen_rusu21@yahoo.com

118 Eugen Rusu et al. 

in PM emission applied to diesel engine might also be applied to gasoline 

engines in the near future. In addition, PM has been found to have adverse 

effects on human health, especially for the smaller particles, since they have 

higher deposition efficiency in the human respiratory system. SI engines have 

been found to be the main sources for fine or ultrafine particles. 

In order to meet new requirements for emission reduction and fuel 

economy a variety of concepts are available for gasoline engines. In the recent 

past new ways have been found using alternative fuels and fuel combinations to 

reduce costs of engine function and fuel consumption. 

The presented concept for a SI-engine consists of combined injection of 

gasoline and hydrogen. A hydrogen enriched gas mixture is being injected 

additionally to gasoline into the engine. 

2. Literature Review of Past Numerical Simulation 

Computer simulation has been always a way to overcome the costs of an 

experimental study. With the help of the simulation researchers could facilitate 

the development of the engines. For the simulation of the thermodynamic 

processes inside the cylinder with hydrogen enrichment not many studies have 

been done. 

G. Fontana from the University of Cassino has conducted a numerical 

investigation using a computational model, suitably developed in order to 

predict the combustion process of dual fuel mixtures. This model is based on 

the KIVA3-V code (Amsden et al., 1989), (Amsden, 1997) and was tested for 

simulating the behavior of the engine fueled with gasoline-hydrogen mixtures. 

To predict the thermal NO formation the Zeldovich kinetic model has been 

used. Tests were done at a wide-open throttle with several equivalence ratios of 

both hydrogen-air mixtures and gasoline-air mixtures (Fontana et al., 2002). 

The computational analysis has marked the possibility of operating with high air 

excess (lean and ultra lean mixtures) without a performance decrease but with 

great advantages on pollutant emissions and fuel consumption. The results have 

shown a reduction of NOx and CO2 emissions. The utilization of hydrogen as an 

additive to a gasoline engine permits reducing the global gasoline consumption, 

measured on the reference engine, to the gasoline consumption calculated 

considering both gasoline injection and hydrogen production. 

Another work regarding the numerical simulation was made by Maher 

Abdul and Haroun Abdul (Abdul-Resul & Abdul-Kadim, 1998) from the 

University of Babylon. They used a detailed model to simulate a four stroke 

cycle of a spark ignition engine fueled with hydrogen-gasoline. The engine 

performance parameters have been improved when operating the gasoline S.I.E. 

with dual addition. The results of the study were promising. It has been found 

that 4% of hydrogen causes a 30% reduction in CO emissions, a 27% reduction 

in NOx, emissions a 34% reduction in specific fuel consumption and increase in 

the thermal efficiency and output power by 5 and 4%.


3. Hydrogen Properties 

Hydrogen has many attractive intrinsic properties that make it a 

promising fuel. 

Table 1 

Properties of gasoline amd hydrogen 

Properties Gasoline Hydrogen 

Molecular weight (g/mol) 2.015 110 

Stoichiometric fuel to air ratio (F/A) 34.3 14.6 

Minimum ignition energy (mJ) 0.02 0.24 

Ignition temperature (K) 858 530 

Adiabatic flame temperature (K) 2384 2270 

Flame speed at 20 0 C (cm/s) 237 41.5 

Limits of flammability (vol % in air) 4.1/75 1.5/7.6 

Quenching gap (cm) 0.06 0.2 

Lower heating value (MJ/kg) 120 44 

Diffusion coefficient at stoichiometric conditions (cm 2 /s) 0.61 0.05 

Main elements that make hydrogen a promising fuel are: 

a) The lower minimum ignition energy of hydrogen ensures a more stable 

ignition and eases cold start engine operation, but raises the danger of abnormal 

combustion. 

b) Its high laminar burning velocity (around five times faster than that of 

gasoline) is expected to increase the indicated efficiency and reduce the cycleto-cycle 

variations in combustion. 

c) The higher diffusion coefficient of hydrogen may enhance the mixing 

process, which also can increase the engine efficiency and help to produce less 

soot and unburned hydrocarbons. 

d) Hydrogen also has a smaller quenching distance compared to gasoline, 

so that the flame can travel further into crevices to ensure more complete 

combustion. 

e) Moreover, adding hydrogen will extend the lean limit due to its lower 

flammability limit in air. However, the lower net energy density of hydrogen 

compared with gasoline for a unit volume of stoichiometric mixture with air 

may reduce the power output. 

f) Adding hydrogen also produces a higher adiabatic flame temperature in 

air, which raises concerns over NOx emissions. 

4. Modelling 

This simulation was done with the help of computer assisted program. In 

this paper AVL BOOST was used. 

BOOST simulates a wide variety of engines, 4-stroke or 2-stroke, spark 

or auto-ignited. Applications range from small capacity engines for motorcycles 

or industrial purposes up to large engines for marine propulsion. BOOST can


also be used to simulate the characteristics of pneumatic systems (AVL Boost 

User Guide). 

In the drawing bellow it is presented the main window in AVL Boost 

program and the engine setup on which the simulation was done. 

Fig.1 – AVL main window. 

4.1. Initial Data 

All tests were done on a gasoline engine with the following 

characteristics 

Table 2 

Engine characteristics 

Type 4 cycle, Gasoline 

Fuel delivery Direct injection 

Bore x Stroke 76.5 x 81.5 

Displacement 1489 dm 3 

Maximum power (kW/CP) 65.6/88 @4800 

Cylinders 4 in line 

Compression ratio 9.2 

4.2. Test Conditions 

Tests were operated at 3000 rpm at WOT. The gasoline tests were done at 

a air fuel ratios of 0.8, 0.85, 0.9, 1 while the test when hydrogen was added to 

the mixture were done at a air fuel ratio of 1, 1.5, 1.7, 2.


Hydrogen percentage was kept at a maximum of 20%. Also tests were 

done at 5%, 10% and 15% of hydrogen percentage. For this test a supercharging 

pressure of 1.3 bar was used. 

5. Results and Discussions 

5.1. Maximum Pressure 

Fig. 2 indicates the profiles of peak cylinder pressure against excess air 

ratio. It can be seen that the peak cylinder pressure increases with the increase 

of hydrogen addition fraction. This is due to the fast burning characteristics of 

hydrogen that make the gasoline–hydrogen mixture be burnt in a short time. 

Past literature presents that the peak cylinder pressure is related to the fuel 

energy flow rate. When the engine fuel air ratio it is increased, the total fuel 

(gasoline and gasoline H2) energy flow rate gently decreases, because of that the 

peak cylinder pressure decreases with the increase of excess air ratio. 

Fig. 2 – Maximum pressure versus air fuel ratio. 

5.2. Maximum Temperature 

Fig. 3 displays the variations of peak cylinder temperature with excess air 

ratio at hydrogen volume fractions of 0%, 5%, 10%, 15%, 20%. NOx emissions 

are related to the peak cylinder temperature so higher temperatures higher NOx 

emissions (Heywood, 1988). Fig. 2 demonstrates that peak cylinder temperature 

drops with the increase of excess air ratio for both the original engine and the 

hydrogen enriched engine due to the reduced fuel energy flow rate. Peak 

cylinder temperature increase with the increase of hydrogen addition fraction, 

this is because hydrogen has a very small energy density.


Fig. 3 – Maximum temperature versus air fuel ratio. 

5.3. BSFC 

Fig. 4 displays break specific fuel consumtion for gasoline and hydrogen. 

Adding hydrogen will reduce the fuel consumtion in both cases when the engine 

was normally aspirated and supercharced with p=1.3bar. When air fuel ratio 

reaches 1.5 bsfc remains at a constant value until air fuel ratio is 2. 

Fig. 4 – BSFC versus air fuel ratio.


5.4 NOx Emissions 

Fig. 5 shows the variations of NOx emissions with excess air ratio at 

different hydrogen volume fractions. NOx emissions depend on temperature and 

oxygen concentration in the cylinder. NOx emissions increases with the increase 

of hydrogen addition fraction. Although the hydrogen-enriched engine produces 

more NOx emissions when the excess air ratio is around stoichiometric 

conditions, NOx emissions for all hydrogen enrichment levels drop to 

acceptable value when the air fuel ratio is around 2. 

Fig. 5 – NOx emisions versus air fuel ratio. 

5.5. HC Emissions 

HC emissions are mainly caused by the unburnt hydrocarbons in IC 

engines. Fig. 6 displays the profile of HC emissions at 3000 rpm. As it is shown 

in Fig. 6, although the λ increases, the decreased HC emissions go with the 

increase of hydrogen addition level, reflecting the capability of hydrogen 

enrichment on improving engine combustion. The explanation is that the 

increased oxygen concentration would help the fuel be completely burnt, and 

thus benefit reducing HC emissions at lean conditions (Rankin, 2008). The high 

flame speed of hydrogen also benefits the hydrogen–gasoline fuel mixture to be 

fully burnt and results in less HC emissions. Moreover, the short quenching 

distance of hydrogen permits the flame of the hydrogen–gasoline mixture to 

propagate much closer to the crevices such as the gaps between piston and 

cylinder wall than that of the pure gasoline, and thereby reduces HC emissions 

(Kahraman et al., 2009). To conclude, HC emissions at lean burn limit are 

gradually reduced with the increase of hydrogen addition fraction for the 

HHGE.


Fig. 6 – HC emissions versus air fuel ratio. 

Comparing NOx emissions result with power output we can conclude that 

the best way to work with hydrogen in addition regardless the percentage of 

hydrogen (1%...20%), best air fuel ratio is 1.62…2. At this air fuel ratio power 

output is similar to that of gasoline, also NOx and HC emissions are lower 

comparative when using gasoline, (Fig. 7). 

Fig. 7– Effective power versus air fuel ratio. 


As a result of the simulation next conclusions were drawn: 

a) The addition of hydrogen improves engine operating at lean 

conditions. 

b) The maximum temperature increases when adds more hydrogen with 

negative effect on the NOx emission level. When using leaner dosages NOx 

emission level is reduced.


c) The maximum pressure increases when adds more hydrogen. Using 

leaner the maximum pressure will decrease. 

d) HC and NOx emissions level are obviously reduced at the increase of 

hydrogen addition and at leaner mixtures use. 

e) The solution of the hydrogen fuelling system, designed and applied, 

has proven its real functionality and offers flexibility in adjusting the parameters 

which assures the hydrogen flow control. 

Acknowledgements. The authors would like to thank AVL GMBH Graz Austria 

for providing the software AVL Boost v.2011.6 used in this study. 

The work has been funded by the Sectorial Operational Programme Human 

Resources Development 2007-2013 of the Romanian Ministry of Labour, Family and 

Social Protection through the Financial Agreement POSDRU/88/1.5/S/60203. 

REFERENCES 

Das L.M., Hydrogen–Oxygen Reaction Mechanism and its Implication to Hydrogen 

Engine Combustion. International Journal of Hydrogen Energy, 21, 703-15 (1996). 

Amsden A.A., O'Rourke P. J., Butler T. D., KIVA-II: A Computer Program for 

Chemically Reactive Flows with Sprays. Los Alamos National Laboratory Report 

LA-11560-MS (May 1989). 

Amsden A.A., KIVA-3V: A Block-Structured KIVA Program for Engines with Vertical 

or Canted Valves. Los Alamos National Laboratory Report LA-13313-MS (July 

1997). 

Fontana G., Galloni E., Jannelli E., Minutilo M., Performance and Fuel Consumption 

Estimation of a Hydrogen Enriched Gasoline Engine at Part-Load Operation. 

SAE Paper 2002-01-2196 (2002). 

Abdul-Resul S., Al-Baghdadi M., Abdul-Kadim S., Al-Janabi H., Improvement of 

Performance and Reduction of Pollutant Emission of a Four Stroke Spark Ignition 

Engine Fueled with Hydrogen-gasoline Fuel Mixture. AVL Boost User Guide, 

1998. 

Heywood J.B., Internal Combustion Engine Fundamentals. McGraw-Hill Book Co., 

1998. 

Rankin D.D., Lean Combustion Technology and Control. Elsevier, 2008. 

Kahraman N., Ceper B., Akansu S.O., Aydin K., Investigation of Combustion 

Characteristics and Emissions in a Spark-ignition Engine Fueled with Natural 

Gas–hydrogen Blends. International Journal of Hydrogen Energy, 34, 1026-1034 

(2009) 

INVESTIGAREA EFECTULUI ADĂUGĂRII HIDROGENULUI ÎN ADAOS LA UN 

MOTOR ALIMENTAT CU BENZINĂ 

(Rezumat) 

Principala preocupare a industriei de automobile în acest moment este de a 

reduce emisiile poluante, consumul de combustibil, şi de asemenea creşterea 

performanţelor motorului. Datorită proprietăţilor sale excelente de ardere hidrogenul 

poate fi folosit în adaos pentru îmbunătăţirea arderii.


Pe lângă cercetarea experimentală simularea proceselor termo-dinamice cu 

ajutorul programelor asistate de calculator este una din cele mai eficiente modalităţi de 

cercetare. 

Lucrarea prezintă rezultatele modelării termo-dinamice ale proceselor pe motorul 

cu aprindere prin scânteie alimentat cu hidrogen în adaos. 

Cercetarea s-a concentrat pe emisiile de (NOx, CO, CO2), consumul de 

combustibil, şi performanţele motorului.


Publicat de 



Secţia 


ANALYSIS OF DIESEL ENGINE OPERATION ON METHANE, 

METHANOL AND DIESEL FUELS 

BY 

ADRIAN SABĂU ∗ , CONSTANTIN DUMITRACHE 

and MIHAELA BARHALESCU 

Received: March 23, 2012 

Accepted for publication: April 15, 2012 

Constanţa Maritime University, 

Department of Mechanical Engineering 

Abstract. In this paper is presented the results of the second-law analysis 

of engine operation with diesel fuel compared with the results of a similar 

analysis for cases where a light, gaseous (CH4) and an oxygenated (CH3OH) 

fuel are used. The overall energy and availability balance during an engine cycle 

are studied analytically. It is shown theoretically that the decomposition of 

lighter molecules leads to less entropy generation compared to heavier fuels. 

Key words: key irreversibility, equilibrium, availability, second-law 

efficiency, combustion. 


Recent studies (Dunbar&Lior, 1994) show that almost 1/3 of the energy 

of fuel is destroyed during the combustion process in power generation. This 

has caused a renewed interest in exergy analyses, since effective management 

and optimization of thermal systems is emerging as a major modern technical 

problem (Bejan et al., 1996). For internal combustion engines, early work 

(Abraham et al., 1994) on the evaluation of the global engine operation using of 

second-law techniques was followed by detailed availability and irreversibility 

calculations during the engine cycle (Rakopoulos et al.,1993). Second-law 

∗ Corresponding author: e-mail: ady.sabau@gmail.com

128 Adrian Sabău et al. 

arguments have been used to evaluate novel engine concepts (Flynn et 

al.,1984), to investigate the effect of operating parameters on efficiency 

(Rakopoulos et al.,1993). The overall energy and availability balance during an 

engine cycle are studied analytically in Refs. (Bedran&Beretta, 1985). This 

article, present a generalization of the method developed in Ref. (Sabau et al., 

2001) for the second-law analysis of the operation with diesel fuel and use it to 

analyze the operation with alternative fuels. Specifically, the cases of a light 

gaseous fuel (CH4) and an oxygenated (CH3OH) fuel were studied. 

2. Experimental Data 

The engine used to study the operation with diesel fuel was a T684 (Euro 

II) made by Tractorul Plant Braşov, direct-injection diesel engine. This was a 

four-stroke, naturally-aspirated, water-cooled engine with a “bowl-piston” 

combustion chamber having a bore of 102 mm, a stroke of 115 mm and rod 

length 182 mm. The compression ratio was 17.5 and the nominal speed range 

was between 800 and 2400 rpm. A five-hole injector nozzle (hole diameter of 

240 µm) was used for diesel fuel injection. It was located near the combustion 

chamber centre with an opening pressure of 210 bar. The full load conditions 

correspond to equivalence ratio of φ=0.625, injection timing of 18°CA BTDC 

and injection duration of 24°CA. Measurements were taken for equivalence 

ratios 0.600 and 0.470 and injection timings of 16, 18 and 20°CA BTDC. 

3. Modelling Approach 

A single-zone model was used to simulate the engine operation. The most 

important assumptions were the following: 

i) The working medium was considered, in general, to be a mixture of 10 

species (O2, N2, CO2, H2O, H2, OH, NO, CO, O, H) and fuel vapour. 

ii) All 10 species were considered as ideal gases. 

iii) Blow by was ignored. 

The residual gas fraction was taken equal to 3% independent of 

conditions of operation, which was a reasonable assumption for the operating 

parameters described above (Rakopoulos et al.,1993). 

Experimental data were available only for diesel fuel injection. 

Computationally, the cases of methane (CH4) and methanol (CH3OH) injection 

were also investigated. Proper processing of the experimentally acquired 

indicator diagrams can yield the preparation and reaction rates of the fuel, as 

described in detail in Refs. (Bedran & Beretta, 1985) and (Abraham et al., 1994). 

The preparation stage involved vaporization of the fuel, its superheating to the 

temperature of cylinder gases and mixing with the oxidizer so that a flammable 

mixture forms in certain regions of the combustion chamber. The intuitive 

assumption that preparation times for gaseous (CH4) injection must be much 

shorter than the ones for liquid injection is not correct (Abraham et al., 1994).


Summarizing the data processing algorithm here for the purpose of 

completeness, we can state that it correlates the differential change in mixture 

composition to the differential of the natural logarithm of pressure in the 

combustion chamber. Similar but significantly simpler approaches to the 

calculation of an effective “burning law” from pressure data are presented in 

Refs. (Sabau et al., 2001). The reacted and prepared fuel quantities computed 

are used as input to the equilibrium code described below. The fitting of a 

Wibbe function (Sabau et al., 2001) to the reacted fuel distribution makes this 

analysis much simpler. As Fig. 1 shows, the Wibbe function offers a flexible 

and accurate approximation to the reacted fuel quantity. The four coefficients of 

the function are calculated using a least squares fit. 

Fig. 1 – Approximation of the 

reacted fuel 

quantity by a Wibbe 

function. 

It should be mentioned that the crank angle at the start of combustion 

must be located for such a formulation to work. The onset of combustion could, 

in principle, be calculated. However, this complicates the model significantly 

without adding any value to the present study. The angle of onset of combustion 

is treated as a parameter of the Wibbe function determined by the least square 

algorithm. Once the molar quantities Npr and Nre of prepared and reacted fuel, 

respectively, are known, determination of the thermodynamic condition of the 

working medium requires calculation of pressure P, temperature T and the 

molar quantities Ni of the aforementioned ten species. This calculation was 

achieved through the solution of 12 equations consisting of: 

The ideal gas equation of state for the mixture 

11 

PV = RT∑ N . 

i= 

1 

i 

(1)


The energy balance for a closed system of variable volume: 

11 11 

∑ ∑ . 

dQ+ d N ( h −h ) − pdV = udN + dT N c −dN 

h 

pr fl fv i i i vi pr fv 

i= 1 i= 

1 

The first two terms on the left-hand side are the heat influx. The third 

term is the work output of the system. The first two terms on the right hand side 

are the change in internal energy of the working medium due to the change of 

its composition and temperature respectively. The third term is the enthalpy 

influx to the working medium due to fuel vaporization. The volume is 

calculated analytically as a function of the crank angle. The thermodynamic 

properties are computed as a function of T by polynomial fittings to the data of 

JANAF Thermo-chemical Tables. Heat transfer through the system boundary is 

calculated as a function of T, the thermodynamic properties of the working 

medium and the wall temperature using Annand’s formula, as described in Ref. 

(Sabau et al., 2001). 

The molar quantities Ni are calculated using a chemical equilibrium 

assumption. Specifically, we assume that the following six reactions among the 

above 10 species were in equilibrium 

H2↔2H, O2↔2O, H2 + O2 ↔2OH, 

N + O ↔2NO, 

(3) 

2 2 

2H + O ↔2HO, 

2 2 2 

2CO + O ↔2CO 

. 

2 2 

At each crank angle, equilibrium equations for these six reactions as well 

as atom number conservation equations for the atoms of C, H, O and N, 

constitute a 10x10 system of equations which can be solved for Ni. This is a 

generalization of the procedure described in Refs. (Rakopoulos et al.,1993) and 

(Sabau et al., 2001) for diesel fuel, where the validity of the chemical 

equilibrium assumption is discussed. In general, the chemical equilibrium 

assumption is not valid for the formation of pollutants. Given the fact that the 

quantities of pollutants are small, the main conclusions of work are not affected. 

4. Second-law Analysis 

Availability of a system is defined as the maximum work that can be 

produced from the system through interaction with its surroundings during a 

reversible transition to a state of thermal, mechanical, and chemical equilibrium 

with its environment, and while heat is exchanged during the transition only 

(2)

Fig. 2 – The closed system the 

related availability streams. 


with this environment. This state of 

equilibrium is defined as the dead state of 

the system and it depends on the pressure 

P0, the temperature T0 and the composition 

of the environment. In the present study, 

the environment is dry atmospheric air 

with P0=1 bar and T0=298 K. Thermal 

equilibrium is achieved when there is no 

heat exchange between the system and the 

environment, i.e. when the system is at 

temperature T0. Similarly, mechanical 

equilibrium is achieved when there is no 

work exchange between the system and its 

environment. Chemical equilibrium is 

achieved when no system component can 

react with the environment. For the present 

case, this means that in the dead state all 

the 11 species of the working medium have been either oxidized or reduced to 

N2, O2, CO2, and H2O. It is shown in Refs. (Bejan, 1988) that the availability of 

a closed system is equal to 

Φ = U + P V −T 

S − G . (4) 

0 

A direct consequence of the above definition is that when a differential 

amount of reversible work dW=(P-P0)dV is extracted from a system, its 

availability is reduced by exactly that amount. Similarly, when a differential 

amount of heat dQ is supplied to the system at temperature T, the system 

availability is increased by an amount dQ(1-T0/T). 

The system under consideration is shown in Fig. 2. It is a closed system, 

since only the part of the cycle from inlet valve closing to exhaust valve 

opening is considered. The only mass influx to the system is the prepared fuel, 

which takes place after injection. 

The differential availability fluxes can also be seen in Fig. 2. On the basis 

of the availability balance, the differential variation of the working medium 

availability is given by 

d Φ=−( P− P)d V + [dQ+ d N ( h −h )](1 − T / T) + dN a −dI 

. (5) 

0 

0 pr fl fv 0 

pr fv 

On the right hand side, the first term is availability out flux from the 

system in the form of work. The second term is availability input through heat 

flux and the third term is availability added to the system by the fuel vapor. The 

term dI makes Eq. (5) different from a conservation equation. 

During combustion working medium availability is destroyed and a 

differential amount dI of combustion irreversibility is produced. Similarly, 

0


during fuel preparation, there is availability destruction due to mixing, which is 

included in the calculation of afr. From Eqs. (4) and (5), it can be seen that in 

order to calculate the combustion irreversibility, one has to determine the dead 

state and calculate afr. 

For the system to reach the dead state at P0=1 bar and T0=298 K, all of 

the working medium constituents have to be either oxidized or reduced to N2, 

O2, CO2 and H2O. Knowledge of the precise mechanism with which this 

happens is not necessary for determination of the molar quantity of each of the 

four species in the dead state. 

This schematic representation has no meaning as a chemical reaction; it 

simply shows the relation of molar quantities between the working medium 

species and the molar quantities of N2 and O2 in the dead state. A similar 

relation for the complete oxidation of an oxygenated fuel vapor gives 

⎛ β γ ⎞ 

β 

CHO α β γ + ⎜α+ − O2 αCO2 

4 2 

⎟ → + . (6) 

⎝ ⎠ 

2H O 

Finally, similar considerations for all species yield the following relations 

for the composition of the dead state 

0 NNO 

NN = N 

2 N + , 

2 2 

0 

NCO2 = NCO2 + αN 

fv + N, 

0 ⎛ β γ ⎞ 

NO = N 

2 O − 

2 ⎜α+ − ⎟N 

fv + 

⎝ 4 2⎠ 

. 

NO + NNO −NH −N 2 CO 

+ 

2 

NOH − N H + , 

4 

0 β 

NHO= N 

2 HO+ N N 

2 fv + H2 

2 

NH + NOH 

+ . 

2 

The differential variation of the G0 term is then given by 

4 ⎡ ⎛ 0 

0 

N ⎞⎤ 

0 = ∑ i ⎢ i 0 0 + 0 ⎜ ⎥ 

⎜ 0 ⎟ 

i= 

1 N ⎟ 

1 

i 

dG d N g ( T , P) RT ln . (8) 

⎢⎣ ⎝∑⎠⎥⎦ The second term of this sum corresponds to the entropy of mixing in the 

dead state. 

2 

(7)


The fuel vapor availability consists of two parts. The first part, usually 

termed “thermal” or “thermo-mechanical” part, is equal to 

a = h −T s −g 

. (9) 

fv,termomechanical fv 0 fv 

0 

fv 

This is the amount of availability introduced by the injected gas. A 

second part, usually termed “thermo-chemical” availability, is due to the ability 

of fuel to react and produce reactants of lower availability. For the calculation 

of the thermo-chemical availability, we consider the reaction of complete 

oxidation of the fuel in atmospheric air. 

It is shown (Bejan, 1988), that the thermo-chemical availability is equal 

to the difference in Gibbs free energy between reactants and products at 

temperature T0 and pressure P0. It follows that 

β ⎛ β γ⎞ 

a g αg g α g 

2 ⎝ 4 2⎠ 

0 0 0 0 

fv,termochemical = fv − CO − 

2 H2O+ ⎜ + − ⎟ O2 

− 

β /2 

⎡ α ⎛ β ⎞ ⎤ ε 

⎢ α ⎜ ⎟ ε ⎥ 

2 

− RT ln ln ⎢ ⎝ ⎠ ⎥, 

α+ β/4 −γ/2 

⎢⎛ β γ⎞ 

⎥ 

ζ 

⎢⎜α+ − ⎟ ζ ⎥ 

⎣⎝ 4 2⎠ 

⎦ 

(10) 

where the last term of the sum accounts for the difference in entropy of mixing 

between products and reactants, and ε and ζ are defined as follows 

β γ 

ε= α+ − , 

4 2 

β 

ζ = 0.79ε+ 

α+ 

. 

2 

(11) 

Adding Eqs. (9) and (10) we get for the total fuel availability 

β ⎛ β γ⎞ 

a = h −T s −αg − g + α + − g 

2 ⎝ 4 2⎠ 

0 0 0 

fv fv 0 fv CO2 H2O⎜ ⎟ O2 

β/2 

⎡ α⎛ β ⎞ ⎤ ε 

⎢ α ⎜ 

2 

⎟ ε ⎥ 

− RT ln ⎢ ⎝ ⎠ ⎥. 

α+ β/4 −γ/2 

⎢⎛ β γ⎞ 

⎥ 

ζ 

⎢⎜α+ − ⎟ ζ ⎥ 

⎣⎝ 4 2⎠ 

⎦ 

− 

. 

(12) 

If we substitute Eqs. (4) and (8) in Eq. (5) and subtract Eq. (2) multiplied 

by T0/T, we get


11 T ⎛ 

0 

N ⎞ 

i 

dI = d Nprafv − [dQ+ d Npr( hfl − hfv )] + t0∑dNi⎜si−Rln ⎟− 

T ⎜ 

i= 1 

N ⎟ 

⎝ ∑ i ⎠ 

T T 

− − + + 

11 4 

0 0 

0 

dNprhfv VdP∑dNicpidT∑d Nig0i. T T i= 1 i= 

1 

. 

(13) 

The energy balance for the system can be reformulated in enthalpy terms as 

dQ+ d N ( h − h ) + VdP= hdN + 

pr fl fv i i 

fv 

10 ⎛ 

⎜∑ i= 

1 

i pi fv 

i= 

1 

⎞ 

pfv ⎟d 

d prhfv + h + N c + N c T − N 

⎝ ⎠ 

10 

∑ 

(14) 

Eq. (14) is the expression of the same energy balance as Eq. (2) 

reformulated in terms of enthalpy. If Eq. (14) is multiplied by T0/T and 

subtracted from Eq. (13), keeping in mind that the expression for gi is: 

it is shown that 

Ni 

gi = hi − Tsi + RTln 

∑ Ni 

, (15) 

T 11 

4 

0 

dI = − ∑ dNi 

gi 

+ ∑ dNi 

g0i 

+ dN pr fv fv 

T 

[ a − h 1 T / T ] 

0 ( − 0 

i= 

1 i= 

1 

. (16) 

This proves that the differential availability destruction is only a function 

of the differential change in mixture composition. It is interesting to note that 

one of the main mechanisms of irreversibility production, namely heat transfer, 

affects the final result only indirectly since dI is not an explicit function of dQ. 

Irreversibility production due to fluid viscosity, as discussed in (Dunbar & 

Lior, 1994), cannot be calculated within the framework of the current singlezone 

analysis. 

5. Results and Analysis 

The “availability balance” during the engine cycle is presented for the 

diesel fuel injection case in Fig. 3. Only the “closed” part of the engine cycle is 

presented. This is a balance between five terms: working medium availability, 

availability transfer through work, availability transfer through heat, input of 

availability with the injected fuel, and destruction of availability (production of 

irreversibility). All five terms are calculated independently as described earlier 

use a computer code writes in Matlab. The “deficit” in the balance caused by 

the employed numerical scheme is of the order of 1% of the availability of the 

injected fuel. Since the availability input with the injected fuel starts and ends 

abruptly with the start and end of injection, respectively, it is expected that 

discontinuities in the derivative of the corresponding curve should appear.


Fig. 3 – Availability balance for 

operation with diesel fuel. 

Fig. 4. – Availability balance for 

operation with methane. 

During fuel injection and preparation, there is a gradual fuel availability 

influx. It can be noted that there is a significant loss of working medium 

availability during the fuel preparation. This is due to working medium cooling 

because of fuel vaporization. The production of irreversibility due to 

irreversible heat transfer during compression is negligible compared to the 

combustion irreversibility. In fact, significant production of irreversibility takes 

place only after ignition. This may be due to insufficient modelling of heat 

transfer within the framework of a single-zone. 

No Fig. experiments 5 – Availability were balance conducted for wi 

operation with methanol. 

Diesel Metan Metanol 

Fig. 6 – Injected fuel availability.


No experiments were conducted with either methanol (CH3OH) or 

methane (CH4) injection, since this would involve re-design of the employed 

hardware well beyond the scope of this work. The above analysis, though, 

shows that the total value of each of the terms of this balance at the end of the 

closed cycle is mainly a function of the total quantity of the injected fuel and 

not of the detailed preparation and reaction rates. 

If we substitute CH3OH and CH4 in the employed code as injected fuels 

and keep the equivalence ratio equal to 0.6 as in the diesel fuel case, we can 

compute “availability balances” like the ones shown in Figs. 4 and 5. The mass 

flow rate of air is also kept the same for all three fuels. This is one simple 

baseline that can be used for the comparison. However the methodology to be 

presented here can be used for more complicated reference criteria. 

The detailed time evolution of each of the terms of the presented balances 

of Figs. 4 and 5 is a function of the precise preparation and reaction rates of 

CH3OH and CH4 which can only be determined from experimental data. 

However, the exhaust gas availability (available for heat recovery devices) and 

the total combustion irreversibility can be approximated without the knowledge 

of this data, provided that reaction has been completed reasonably early relative 

to the exhaust valve opening timing. Figs. 6–8 show the comparison between 

injected fuel availability, exhaust gas availability and combustion irreversibility, 

for injection with the same equivalence ratio. 

Diesel Metan Metanol Diesel Metan Metanol 

Fig. 7 –Exhaust gas availability. Fig. 8 – Combustion irreversibility. 

It can be seen that the injected fuel availability is significantly higher for 

the methanol case. Because of the oxygen content of the fuel, higher molar 

quantity of fuel is necessary for combustion with the same equivalence ratio. 

Comparing the results for methane and diesel fuel, we observe a significant 

decrease of combustion irreversibility for methane combustion.


This is in agreement with the theoretical expectation that the 

decomposition of the larger hydrocarbon molecules from diesel fuel during 

chemical reaction should create higher entropy increase than the decomposition 

of a lighter methane molecule. The decrease of combustion irreversibility is 

even higher for the case of oxygenated fuel. 

An “second-law efficiency” 

can be defined as the ratio of 

extracted work over injected fuel 

availability. It should be stressed 

that this ratio is only calculated 

here for the closed part of the cycle 

and is not the second-law 

efficiency of the full cycle. 

Comparison of the efficiency for 

the three cases is presented in Fig. 

9. It seems that the use of 

alternative fuels leads to better 

second-law efficiency. However, 

Diesel Metan Metanol 

especially as far as the comparison 

Fig. 9 – Second-law efficiency. 

between methane and methanol is 

concerned, the difference is small so that conclusions should be drawn with 

caution given the assumptions of the current analysis. 

Methanol vapor, for instance, was assumed to be an ideal gas but this is 

only a coarse approximation. More accurate modelling of the thermodynamic 

properties of methanol may alter the injected fuel availability. 

Moreover, for the accurate calculation of the irreversibility created during 

combustion, an accurate computation of methanol combustion kinetics is 

necessary. It would also be interesting to investigate the effect of the fact that 

this is a single-zone analysis by incorporating the current arguments in multizone 

codes. Revisiting the above assumptions may yield more accurate 

information about the advantages Figs. 6–9 imply for oxygenated fuels. 

However, the main point remains that entropy increase during combustion is 

smaller for such fuels because of the lower mixing entropy of the mixture of 

reactants. 


1. A method for the analytic calculation of irreversibility and exhaust gas 

availability is generalized and used to evaluate alternative fuels in a directinjection, 

naturally-aspirated, four-stroke diesel engine. 

2. Combustion irreversibility is shown to be the main source of 

irreversibility during the engine operation and its differential variation is


calculated as an analytic function of the differentials of the quantities of the 

constituents of the working medium. 

3. Decrease in combustion irreversibility is achieved with the use of a 

lighter (methane) and, even more, of an oxygenated (methanol) fuel for the 

same equivalence ratio of operation. 

4. This is due to the combustion characteristics of these fuels which 

involve lower entropy of mixing in the combustion products. So far, the driving 

force for the study of alternative fuels has been to decrease pollutant emissions. 

5. Given the increasing interest in conservation, the use of such fuels may 

be seen in a different perspective since there are indications that it involves a 

more effective use of the chemical energy of the fuel. 

REFERENCES 

Abraham J., Magi V., MacInnes J., Bracco F.V., Gas vs. Spray Injection: Which Mixes 

Faster? SAE Paper No. 940895 (1994). 

Bedran E.C., Beretta G.P., General Thermodynamic Analysis for Engine Combustion 

Modeling, SAE Paper No. 850205 (1985). 

Bejan A., Advanced Engineering Thermodynamics. New York, John Wiley and Sons 

Inc., 1988. 

Bejan A., Tsatsaronis G., Moran M., Thermal Design and Optimization. New York, 

John Wiley and Sons Inc., 1996. 

Dunbar W.R., Lior N., Sources of Combustion Irreversibility. Combust Sci Technol., 

103, 41-61 (1994). 

Flynn P.F., Hoag K.L., Kamel M.M., Primus R.J., A New Perspective on Diesel Engine 

Evaluation Based on Second Law Analysis. SAE Paper No. 840032 (1984). 

Rakopoulos C.D., Andritsakis E.C., Kyritsis D.C., Availability Accumulation and 

Destruction in a DI Diesel Engine with Special Reference to the Limited Cooled 

Case. Heat Recov Syst CHP, 13, 61-75 (1993). 

Sabău A., Buzbuchi N., Şoloiu U.V.A., Combustion Numerical Modelling in Diesel 

Engines, “A XI-a Conferinţă Naţională de Termotehnică”, Galaţi, 2001. 

ANALIZA FUNCŢIONĂRII MOTORULUI DIESEL CU COMBUSTIBILII 

METAN, METANOL ŞI DIESEL 

(Rezumat) 

În această lucrare este prezentată o metodă pentru calculul irreversibilităţii şi a 

disponibilului de energie utilă în cazul unui ciclu diesel. Studiul efectuat se bazează pe 

utilizarea principiului al doilea al termodinamicii pentru evaluarea performanţelor 

termodinamice ale unui motor diesel. Analiza comparativă este realizată utilizând un 

program de calcul conceput şi realizat de autori, care permite aprecierea eficienţei 

energetice a motorului diesel la schimbarea motorinei cu metan sau metanol. Modelul 

de calcul utilizat este unul termodinamic, fiind completat acolo unde a fost cazul cu


formulări fenomenologice pentru a putea mai bine scoate în evidenţă, aspectele 

importante ale studiului. Calibrarea s-a facut utilizând date experimentale măsurate pe 

motorul T650 produs la Uzina Tractorul din Braşov la funcţionarea pe motorină. Studiul 

a relevat că procesul de combustie este principala sursă de ireversibilitate şi un rol 

hotărâtor îl joacă ireversibilitatea chimică dată de modul de descompunere al 

hidrocarburilor care compun combustibilul.


Publicat de 



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MIXING RULES FOR PREDICTING THE IGNITION 

PROPERTIES OF AUTOMOTIVE DIESEL ENGINE FUELS 

BY 

ANCA ELENA ELIZA STERPU* and ANCA IULIANA DUMITRU 

Received: March 30, 2012 

Accepted for publication: April 14, 2012 

”OVIDIUS” University, Constanţa, 

Department of Chemistry and Chemical Engineering 

Abstract. Biodiesel can be defined as fatty acid alkyl esters from 

vegetable oils or animal fats for using as fuel in diesel engines and heating 

systems due to its technical, environmental and strategic advantages. Because 

its physical properties and chemical composition are distinctly different from 

conventional diesel fuel, biodiesel can alter the fuel injection and ignition 

processes whether neat or in blends. Most research on computational 

combustion presents complex mathematical models where the fuel physical 

properties are important parameters. 

The aim of this study is to present the importance of the physical 

properties and their relation to the internal combustion engine and also 

proposing a method to determine the volumetric proportion of biodiesel 

which will have efficient combustion in compression engines. For this 

purpose, the basic properties (density, viscosity, flash point, heating value, 

calculated cetane index and three different points of the distillation curve: 

T10, T50 and T90) of several rapeseed (Canola) oil biodiesel–diesel fuel 

blends (B0, B5, B10, B15, B20, and B100) were measured according to the 

corresponding ASTM standards. In order to predict these properties, mixing 

rules are estimated as a function of the volume fraction of biodiesel in the 

blend. 

Key words: automotive diesel engine, ignition characteristics, 

biodiesel–diesel blends, calculated cetane index. 

* Corresponding author: e-mail: asterpu@univ-ovidius.ro

142 Anca Elena Eliza Sterpu and Anca Iuliana Dumitriu 


Diesel engines are the axis of world industry, dominating sectors like 

road transport, agricultural, military, construction, maritime propulsion and 

stationary electricity production. In recent years, the concern over the depleting 

world reserves of fossil fuels and more stringent emission regulations for 

harmful pollutants have led to the resolute efforts in searching for renewable 

alternative fuels and ultra-low emission combustion strategies. Biodiesel fuel, 

which is identified as a clean alternative fuel, has become commercially 

available in a number of countries (Zheng et al.., 2008),. In addition, this 

biofuel is completely miscible with petroleum diesel, allowing the blending of 

these two fuels in any proportion. Biodiesel can be used neat or blended in 

existing diesel engines without major modification in the engine hardware. 

However, differences in the chemical nature of biodiesel (mixture of monoalkyl 

ester of saturated and unsaturated long chain fatty acids) and conventional 

diesel fuel (mixture of paraffinic, naphthenic and aromatic hydrocarbons) result 

in differences in their basic properties, affecting engine performance and 

pollutant emissions. Biodiesel, produced from any vegetable oil or animal fat, 

generally has higher density, viscosity, and cetane number, and lower volatility 

and heating value compared to commercial grades of diesel fuel (Benjumea et 

al., 2008). Biodiesel most commonly are made from soybean oil in the United 

States and from rapeseed (Canola) oil in Europe using methanol. 

1.1. Biodiesel as Engine Fuel 

Since most modern diesel engines have direct injection fuel systems, and 

these engines are more sensitive to fuel spray quality than indirect injection 

engines, a fuel with properties that are closer to diesel fuel is needed (Canakci, 

2007). The best way to use vegetable oil as fuel is to convert it in biodiesel. 

Biodiesel is made from natural, renewable sources such as new or used 

vegetable oils and animal fats. The resulting biodiesel is almost similar to 

conventional diesel in its main characteristics. Biodiesel is an oxygenated fuel 

and leads to more complete combustion; hence CO emissions reduce in the 

exhaust. Biodiesel contains no petroleum products, but it is compatible with 

conventional diesel and can be blended in any proportion with mineral diesel to 

create a stable biodiesel blend. The level of blending with petroleum diesel is 

referred as Bxx, where xx indicates the amount of biodiesel in the blend (i.e. 

B20 blend is 20% biodiesel and 80% diesel). It can be used in compression 

ignition (CI) engine with no significant modifications to the engine (Agarwal, 

2007). 

1.2. Combustion Characteristics of Biodiesel 

It is important to know the combustion properties of biodiesel–diesel 

blends. Some of these properties are required as input data for predictive and


diagnostic engine combustion models. Additionally, it is necessary to know if 

the fuel resulting from the blending process meets the standard specifications 

for diesel fuels. Given the difficulty of obtaining the combustion properties of 

the blend by measurement, the ability to calculate these properties using 

blending or mixing rules is very useful. 

Some authors (Clements, 1996) proposed blending rules for estimating 

the density, heating value, viscosity, cetane number and cloud point of biodiesel 

as a function of its methyl esters profile. By comparing predicted values with 

experimental data, they found that the typical average errors were less than 2%, 

with the exception of viscosity, where the average error was 10%. Similar 

blending rules for calculating properties of methyl ester blends have also been 

used by other researchers for estimating properties of biodiesel–diesel blends. 

Tat and Van Gerpen (Tat & Van Gerpen, 1999, Tat & Van Gerpen, 2000) 

carried out a study on the density and kinematic viscosity of a commercial 

soybean oil biodiesel and its blends with two samples of diesel fuels at 75%, 

50% and 20% biodiesel (by weight). Results showed that both density and 

viscosity increased with the increase in the percentage of biodiesel. Blend 

kinematic viscosities were estimated by means of a blending rule. The 

maximum differences between predicted and measured viscosities were less 

than 3.74% of the measured values for the biodiesel blends with both diesel 

fuels. Experimental investigations have been carried out by Yuan, Hansen and 

Zhang (Yuan et al., 2004) to examine the density using three types of biodiesel 

(two produced from soybean oil and another biodiesel prepared from yellow 

grease). These were blended with diesel fuel at 75%, 50% and 25% biodiesel 

(by weight), and tested from close to the biodiesel crystallization onset 

temperature up to 100 o C. The measurements indicated that all biodiesel fuels 

and their blends with diesel fuel had a linear specific gravity–temperature 

relationship similar to pure diesel fuel. The densities of the blends estimated by 

a mixing rule showed an average absolute deviation (AAD) of less than 0.43% 

for all tested fuels in the temperature range studied. 

The present paper presents experimental density and viscosity data, 

together with the flash point and distillation curves for rapeseed oil biodiesel 

and its blends with diesel fuel. Additionally, the calculated cetane index and 

heating value of the tested neat fuels and their blends is obtained by ASTM 

D976 and ASTM D4737 for cetane index and ASTM D4868 for heating value 

recommended for hydrocarbon fuels. The aim of this work is to evaluate mixing 

rules for predicting the combustion properties of rapeseed oil biodiesel–diesel 

fuel blends as a function of biodiesel content (volume fraction). This approach 

is more likely to be applicable to other biodiesel fuels and blends than empirical 

correlations, which would be valid only for the specific blends tested.


2. Methods 

2.1. Blend Preparation 

The biodiesel used in this work was produced by basic methanolysis of 

rapeseed oil using a methanol/oil molar ratio of 12:1, with 0.5% NaOH by 

weight as the catalyst. The reaction temperature and time were 65 o C and 1 h, 

respectively. The biodiesel thus produced by this process is totally miscible 

with mineral diesel in any proportion. 

The experimental samples were prepared by mixing a commercial diesel 

fuel with rapeseed biodiesel in different proportions (B0, B5, B10, B15, B20, 

and B100). Blends were prepared on a volume basis at 25 o C. 

2.2. Experimental 

Both density and kinematic viscosity were simultaneously measured 

according to ASTM D445 using a Anton Paar device, SVM 3000 type. The 

Closed-cup Flash Point Analyzer, Pensky-Martens type, was used to measure 

the flash point of biodiesel samples with different compositions. This flash 

point analyzer is operated according to the standard test method, ASTM D93. 

The analyzer incorporates control devices to adjust the heating rate of a sample. 

The test interval is 1 o C and the heat rate is 2 o C/min. The low heating value was 

determined following ASTM D4868, as a function of the fuel density, sulfur, 

water, and ash content. All fuels tested were distilled following the procedure 

established by ASTM D86. In addition to the initial and final boiling points, 

temperatures for distilled percentage multiples of 10 were also registered. The 

calculated cetane index for the fuels tested was determined as a function of their 

densities and distillation curve data using the empirical correlations 

recommended by ASTM D976 and D4737. 

3. Results and Discussion 

3.1. Basic Properties of Pure Fuels 

Experimental data on the properties of rapeseed oil biodiesel and diesel 

fuel are presented in Table 1. Rapeseed oil biodiesel (ROB) is slightly denser 

and more viscous, and has a narrower boiling interval than diesel fuel, 

indicating that the fatty acid methyl esters in biodiesel have similar boiling 

points. Diesel fuel, meanwhile, is composed of a wide variety of hydrocarbons 

with different volatilities. ROB has a low heating value lower than diesel fuel 

due to the presence of oxygen in the methyl ester molecules. This difference in 

heating values, expressed as energy per unit mass, is slightly reduced when it is 

reported as energy per unit volume, given the greater density of ROB. As a 

result of these differences, the calculated cetane index for ROB is higher than 

for diesel fuel; this agrees with the tendency reported by several researchers for 

both cetane numbers and CCI.


Table 1 

The main properties of rapeseed oil biodiesel and diesel fuel 

Properties Units ASTM Diesel Rapeseed 

method fuel oil biodiesel 

Density at 25 o C g/cm 3 D445 0.8374 0.8695 

Kinematic viscosity at 40 o C cSt D445 4.6597 6.1415 

Flash Point 

o 

C D93 67 107 

Initial boiling point 

o 

C D86 170 296 

Temperature at 50% recovered 

o 

C D86 260 334.5 

Final boiling point 

o 

C D86 328 343 

Calculated cetane index - D976 46.99 52.23 

D4737 46.65 54.25 

Low heating value kJ/kg D4868 42650 40301 

Refractive index - D1218 1.467 1.45 

Sulphur content % D5453 0.07 - 

Water content % D6304 0.012 0.017 

Ash content % D482 0.001 0.003 

3.2. Effect of Biodiesel Content 

Mixing rules are used for estimating the properties of blends as a function 

of pure fuel properties and biodiesel content. The suitability of these rules is 

evaluated by means of the absolute average deviation (AAD), calculated as 

(Benjumea et al., 2008) 

100 

AAD 

NP i 

φ − φ 

NP 

EXP PR 

= ∑ , (1) 

= 1 φEXP 

where NP is the number of experimental points, φ is the property to be predicted 

and the subscripts are EXP for experimental and PR for predicted. 

For predicting viscosity of liquid mixture, a geometric volume average 

form equation (Benjumea et al., 2008) (Arrhenius-type equation) can be used 

ln η = V ln η + V ln η . 

B ROB ROB B0 B0 

In the above equation, V is the volume fraction and the subscripts are B 

for blend, ROB for biodiesel and B0 for diesel fuel. For the remaining 

properties, Kay’s mixing rule (Benjumea et al., 2008) is used 

n 

B i i 

i= 

1 

(2) 

φ = ∑ x φ , (3) 

where: φB is the property of the blend and φi is the respective property of the i th 

component. Using volume fraction instead of molar fraction, Eq. (3) for a 

binary mixture takes the form of an arithmetic volume average:


φ = V φ + V φ . (4) 

B ROB ROB B0B0 3.2.1. Viscosity. Fig. 1 shows the effect of biodiesel content on the 

viscosity of the blends. As can be seen in this figure, viscosity is directly 

proportional to biodiesel content. The AAD obtained using the selected mixing 

rule (Arrhenius-type equation) for estimating blend viscosity was 0.5%. 

Fig. 1 – Variation of blend kinematic viscosity with biodiesel content. 

3.2.2. Density. The variation in blend density with biodiesel content is 

shown in Fig. 2. As was expected, density slightly increases with biodiesel 

content. The AAD value obtained using Kay’s mixing rule to estimate blend 

densities was 0.05%. 

Fig.2 – Variation of blend density with biodiesel content. 

3.2.3. Heating Value. Low heating values (LHV) were determined, taking 

into account the ASTM D 4868 method. As can be seen in Fig. 3, the LHV of 

blends decreases in direct proportion to the biodiesel content. 

The AAD obtained using Kay’s mixing rule to estimate the LHV of 

blends was 0.12%.


Fig. 3 – Variation of blend low heating value with biodiesel conten t. 

3.2.4. Flash Point. The flash point of a flammable liquid is that 

temperature, 

as determined under experimental conditions, at which the vapor 

pressure of the substance is sufficient to form a combustible mixture with air. 

The specific flash point value is generally measured by a flash point analyzer 

with closed cup test methods. For the blends tested (B5, B10, B15 and B20) the 

effect of biodiesel content on the flash point values of the blend was not 

significant, and the results obtained using ASTM D93 method were slightly 

decreases. For B100, the value of the flash point obtained was 60% higher than 

diesel fuel flash point. The AAD obtained using Kay’s mixing rule to estimate 

the flash point values of blends was 30.77%. 

Fig. 4 – Variation of blend flash point with biodiesel content. 

As can be seen from Fig.4. and the AAD value, the Kay’s mixing rule is 

not suitable for predicting the effect of biodiesel content on the flash point 

values of the blend.


3.2.5. Distillation Curve Points. Fig. 5 shows the distillation curves for 

pure fuels (B0, B100) and for B5, B10, B15 and B20 blends. As can be seen in 

this figure, all curves 

tend to intercept at a point which defines two zones with 

different 

behaviour. Before the interception point (10% distilled percentage), 

distillation temperature increases with distilled percentage, while after this point 

the trend is reversed. 

3.2.6. Calculated Cetane Index. Cetane numbers rate the ignition properties 

of diesel fuel as a measure of the fuel’s ignition on compression as measured by 

ignition 

delay. The ignition 

delay in a diesel engine is defined as the time 

between 

the start of fuel injection and the start of combustion. The ignition 

quality of a fuel is usually characterized by its cetane number. Higher cetane 

number generally means shorter ignition delay (Canakci, 2007). Cetane number 

is measured in a single-cylinder engine compared with reference blends of ncetane 

and heptamethylnonane according to ASTM D613 or in a constant 

volume combustion apparatus following ASTM D6890. However, these tests 

are awkward and expensive. For this reason there have been many attempts to 

develop methods to estimate the cetane number of a fuel. In order to 

differentiate predicted from measured values, the former is called calculated 

cetane index (CCI), and the latter cetane number. 

Fig.5 – Distillation curves of rapeseed oil biodiesel – diesel fuel blends. 

The CCI can be determined using empirical correlations in accordance 

with the ASTM D976 and D4737. Using ASTM D976, this parameter 

is 

btaine d as a function of fuel density at 15 curve 

point. 

o o 

C and the T50 distillation 

ASTM D4737 additionally takes into account T10 and T90 values. Figs. 

6 and 7 show the CCI values using the two standards. Distillation curve points 

were taken according to the measured values shown in Fig. 5 and the values of 

density at 15 o C were obtained from Fig. 2. For comparison, the CCI values 

were then calculated from values of blend densities at 15 o C and distillation 

curve points obtained using Kay’s mixing rule.


Fig. 6 – Variation of blend calculated cetane index with biodiesel content. 

Fig. 7 – Variation of blend calculated cetane index with biodiesel content. 

The CCI values obtained using ASTM D4737 methods were almost 

similar to the CCI values resulted from mixing rule data and were in better 

agreement 

with the measured cetane number reported for ROB by other studies 

(Benjumea 

et al., 2008). Using ASTM D976, the values of the CCI obtained 

was higher than that obtained using the mixing rules. The better prediction 

obtained using ASTM D4737 is most likely due to the inclusion of the 

particular characteristics of the distillation curve. The AADs obtained using 

Kay’s mixing rule to estimate blend CCI were 0.26% and 4.39% for the data 

corresponding to ASTM D4737 and D976 respectively. 


The objective of this study was to characterize the 

effect of volumetric 

proportion on the combustion properties of biodiesel-diesel blends. The 

combustion characteristics of B0, B5, B10, B15, B20, and B100 were analyzed. 

Based on the experimental results, the following conclusions can be drawn:


a) Combustion properties of rapeseed oil biodiesel–diesel fuel blends 

were measured according to ASTM standards. According to the low values 

obtained for the absolute average deviations, it was found that simple mixing 

rules are suitable for predicting the combustion properties of rapeseed oil 

biodiesel–diesel 

blends as a function of biodiesel content (except the flash 

point). 

b) Density and viscosity of rapeseed oil biodiesel–diesel fuel blends 

slightly increases with biodiesel content. 

c) Kay’s mixing rule is not suitable for predicting the effect of biodiesel 

content on the flash point values of the blend, because the blends flash point 

values were slightly decrease with the biodiesel content and for B100 the flash 

point value was 60% higher than diesel fuel flah point. 

d) Biodiesel has low heating value, 5.5% lower than diesel and calculated 

cetane index, about 16% higher. 

e) All distillation curves tend to intercept at a point (10% distilled 

percentage) which defines two zones with different behaviour. Before the 

interception point, distillation temperature increases with distilled percentage, 

while after this point the trend is reversed 

f) The ASTM D4737 produces a calculated cetane index of rapeseed oil 

biodiesel in better agreement with the reported cetane number than the 

corresponding to the ASTM D976. This result is most likely due to the fact that 

the former standard takes into account the particular characteristics of the 

distillation curve. 

REFERENCES 

Agarwal A.K., Biofuels (Alcohols and Biodiesel) Applications as Fuels for Internal 

Combustion Engines. Progress in Energy and Combustion Science 33, 3, 233-271 

(2007). 

Benjumea P., Agudelo J., Agudelo A., Basic Properties 

of Palm Oil Biodiesel–diesel 

Blends. Fuel, 87, 12, 2069-2075 (2008). 

Canakci M., Combustion Characteristics of a Turbocharged DI Compression Ignition 

Engine Fueled with Petroleum Diesel Fuels and Biodiesel. Bioresource 

Technology, 

98, 6, 1167-1175 (2007). 

Clements D.L., Blending Rules for Formulating Biodiesel Fuel. Liquid Fuels and 

Industrial Products for Renewable Resources. 

Proceedings of the Third Liquid 

Fuels Conference American Society of Agricultural Engineers. Nashville TN., 

Sept. 15-17, 1996, pp. 44-53. 

Tat M .E., Van Gerpen J.H., The Kinematic Viscosity of Biodiesel and its Blends with 

Diesel Fuel. JAOCS, 76, 12, 1511-3 (1999). 

Tat M.E., 

Van Gerpen J.H., The Specific Gravity of Biodiesel and its Blends with Diesel 

Fuel. JAOCS, 77, 2, 115-9 (2000). 

Yuan W., Hansen A.C., Zhang Q., The Specific Gravity of Biodiesel Fuels and their 

Blends with Diesel Fuel. Agri. Eng. Int., CIGR J. Scientific Res. Div., Manuscript 

EE 4, 4, 1-11 (2004).


Zheng M., Mulenga M. C., Reader G. T., Wang M., Ting D. S-K., Tjong J., Biodiesel 

Engine Performance and Emissions 

in Low Temperature Combustion. Fuel, 87, 

12, 714-722 (2008). 

REGULI DE AMESTECARE PENTRU PREDICŢIA PROPRIETǍŢILOR DE 

ARDERE 

ALE COMBUSTIBILILOR PENTRU MOTOARELE DIESEL ALE 

AUTOVEHICULELOR 

(Rezumat) 

Datoritǎ proprietǎţilor fizico-chimice diferite faţǎ de cele ale combustibilului 

dies el convenţional (motorina), biodieselul poate afecta injecţia de combustibil şi 

procesul de ardere, indiferent cǎ este în stare purǎ sau 

sub formǎ de amestec cu 

motorinǎ. Multe cercetǎri în domeniul combustiei, 

prezintǎ modele matematice 

complexe ce conţin drept principali parametri, proprietǎţi 

fizice ale combustibilului. 

Scopul acestei lucrǎri este de a evidenţia importanţa proprietǎtilor de ardere şi 

influenţa lor asupra performanţelor motorului cu combustie internǎ şi de asemenea de a 

propune o metodǎ de a determina proporţia volumetricǎ de biodiesel care prezintǎ cea 

mai eficientǎ comportare în motor. In acest scop au fost determinate experimental unele 

proprietǎţi cum ar fi: densitatea, vâscozitatea cinematicǎ, punctul de inflamare, puterea 

calorificǎ 

inferioarǎ, indicele cetanic calculat şi curbele de distilare STAS ale diferitelor 

amestecuri de biodiesel din ulei de rapiţǎ (Canola) cu motorinǎ (B0, B5, B10, B15, B20 

şi B100), conform metodelor ASTM corespunzǎtoare. In acest scop au fost evaluate 

reguli simple de amestecare existente în literatura de specialitate ca funcţii ale fracţiei 

volumice de biodiesel în amestec. De asemenea au fost calculate valorile deviaţiei medii 

absolute (DMA) pentru verificarea aplicabilitǎţii regulilor de amestecare. 

În urma realizǎrii determinǎrilor experimentale şi calculelor s-a observat cǎ 

pentru predicţia proprietǎţilor de ardere ale combustibililor (B0, B5, B10, B15, B20 şi 

B100) se pot aplica reguli simple de amestecare ca funcţii ale compoziţiei şi 

proprietǎţilor combustibililor puri (B0 şi B100). Valorile DMA obţinute în acest caz 

sunt foarte scǎzute, fiind cuprinse în intervalul 0.05…4.39%. Excepţie de la aceastǎ 

situaţie face punctul de inflamare care nu poate fi determinat utilizând regulile de 

amestecare datoritǎ valorilor ridicate ale DMA obţinute în acest caz. Metoda ASTM 

D4737 pentru calculul indicelui cetanic este mai exactǎ decât metoda ASTM D976, ca 

urmare a utilizǎrii temperaturilor T10, T50 şi T90 din curba de distilare STAS a 

combustibilului în cazul metodei ASTM D4737.


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OPTIMIZATION OF THE AIR FLOW CONTROL FOR A CAR 

ENGINE COOLING SYSTEM 

BY 

MARIUS RECEANU ∗1 , DAN DĂSCĂLESCU 2 and ADRIAN SACHELARIE 2 

1 MITECO S.A. 

2 “Gheorghe Asachi” Technical University of Iaşi, Romania, 

Department of Automotive and Mechanical Engineering 



Abstract. The servomechanism for air flow control has as effect on cooling 

starting: fuel consumption and wear reduction. The shorting of the warming up 

engine period has been ascertained. As consequence the wear and the pollution 

reduce in time of cold starting. A relevant data in the prosess is the surrounding 

temperature. The variation of cooling air flow produces modifications in the 

dynamic behaviour of care-engine cooling system, especially when the engine 

and coolant liquid are cold by the time of approach to steady states. 

Key words: car engine cooling system, cooling air flow, car engine system, 

automatic control and servomechanisms. 


When fuel is burnt in an internal combustion engine roughly 30% energy 

in the fuel is available as us full work. The rest is rejected as heat in a 

accordance with the second law of thermodynamics. Roughly 30% of the 

energy is rejected in the exhaust flow, roughly 30% of fuel energy is rejected 

into the engine coolant and the rest is radiated off the engine structure into the 

atmosphere. 

The heat lost to circulating cooling is: 

Qr = qr cl ( T ri − T re) 

, (1) 

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

154 Marius Receanu et al. 

where: qr – the coolant flow rate in the radial loop, cl – the heat capacity of unit 

coolant, Tri – the coolant temperature at the inlet of the radiator, Tre – the 

coolant temperature at the outlet to the radiator. 

According to Eq. (1) it results that the heat Qr of car-engine cooling 

system can be modified through the coolant flow, qr or cooling air flow, 

ΔT=Tri-Tre. A method for the cooling air flow control is the utilization of blind, 

who obturates the radiator of car-engine for the ambient temperature < +5 ˚C. 

2. Experiments Results 

For two situation of the radiator the heat balances are: 

a) is not obturated 

b) is obturated 

h l e r l 

( ri re) 

exh rest 

Q C Q Q q c T T Q Q 

= − + − + + ; (2) 

( ri re) 

Q C Q Q q c T T Q Q 

' ' 

h l e r l 

= − + − + + (3) 

exh rest 

where: Ch is the fuel mass flow, Ql – specific energy content. 

The experiment was performed in the same conditions: the speed of 

motor vehicle has (approximately) constant value and the segment of direction 

is identically and it results: Qe=const. heat converted into mechanical energy, 

Qexh=const. heat lost in exhaust gases, Qrest=const. 

After the substraction of Eqs. (2) and (3) it results 

' ' 

( Ch Ch) cl( T re T re) 

Q q 

− = − . 

(4) 

l r 

Because the temperature T’re>Tre, when the radiator is obturated 

' 

Ch> Ch. 

(5) 

The fuel economy is evidently in the case b) when the radiator is partial 

or total obturated. The Eq. (4) has valability when the radiator is partial or total 

obturated and the process has practical validity when the ambient temperature is 

Tamb


Fig. 1 – Experimental installation mounted on the motor vehicle. 

Fig. 2 – Theoretical graphs Tri(t), Tre(t) and coolant flow qr(t) 

in the situations a) and b). 

In Fig. 2 we denote: tcl – the time after when the valve of thermostat 

becomes to open (1.FTP the first transient phase, the coolant flow becomes to 

increase on Laptop); top – the time after when the valve is fully open (approach 

to steady states). 

We define: ttr=top-tcl (2.STP the second transient phase, time between the 

extreme positions: valve is fully closed or fully open); Scl=ν tcl the distance 

covered with the speed ν in the time tcl. 

When the average fuel consumption is ck [l/100 km] for distance, it 

results the fuel consumption: ccl=ckscl. The average fuel consumption in the 

second transient phase is: ctr=cT - ccl and in hour is Ch=3600Ctr/ttr.


In Fig. 2 the functions of temperatures Tri and Tre tend to distinct values at 

approach to steady states. In situation b) - radiator obturated, the speed of 

increase of the temperatures at the approach to steady states (observed on 

Laptop) is greater as the situation a). 

The Eq. (4) can be written in the second transient phase for this situations 

a) and b) 

( Ch Ch) cl( Θ Θ) 

Q q 

' ' 

, 

l r 

− = − (6) 

where: Θ and Θ’ are the mean of functions ”product”: qr(t)[Tri(t)–Tre], 

respectively: q’r(t)[T’ri–T’re] in the second phases: [0, tr] or [0, t’r] 

t op 

= 1 

q () t ⎡ 

ri() t re d 

r T −T 

⎤ t 

top − ⎣ ⎦ 

tcltcl 

Θ ∫ , (7) 

t 

' 

op 

' 1 

' ' 

Θ = () () d 

' ∫ q t ⎡ 

r T ri t −T 

⎤ 

re t. 

(8) 

⎣ ⎦ 

top −tcl 

t 

cl 

The function Θ and Θ’ – (7), (8), can be calculated with software in 

LabVIEW, after experiments and the relations 

' 

Θ −Θ 

' 

C h C h cl Q l 

− = ; 

' 

Θ −Θ 

' 

C h C h cl Q l 

' 

h 

= − , 

and fuel economy, in percent : Ch 

−C 

ε = 100 , in MATLAB/Simulink. 

Ch 

For different position of the blind who obdurate the radiator: Fig. 3 (in 

positions: 0 –radiator is not obturated; 1, 2 ,3 – radiator is obturated) and some 

ambient temperatures


the graphs: Tri(t), Tre(t), qr(t) and T’ri(t), T’re(t), q’r(t) in the position “0” and 

“1” of blind. 

Table 1 

Poz ck v cT tcl top ccl ctr Ch Ch’ ε Ch-Ch’ 

. l/100 km Km/h l s s l l l/h l/h % l/h 

0 9.4 21.7 0.4 270 600 0.153 0.247 2.695 – – – 

1 9.5 22.1 0.4 270 500 0.157 0.243 2.646 2.46 7.414 0.196 

2 10.4 25 0.3 270 390 0.195 0.105 3.15 2.85 9.535 0.300 

3 9.1 20.2 0.2 270 370 0.138 0.062 2.237 1.85 17.31 0.387 

Tri(t) T’ri(t) 

t[min] 

Tre(t) T’re(t) 

t[min] 

qr(t) q’r(t) 

t[min] 

Fig. 4 – Graphs of Tri(t), Tre(t), qr(t) and T’ri(t), T’re(t), q’r(t) 

for the positions: “0” and “1” of blind. 

The experimental data demonstrate that the blind, who obturates the 

radiator, decreases time of the second transient phase and produces a fuel 

economy (Table 1). 

3. Servomechanism for the Air Flow Control 

The car-engine cooling (1) system can be modified through the coolant 

flow (qr) or cooling air flow (ΔT=Tri–Tre). A method for the cooling air flow 

control is the utilization of blind, who obturates the radiator of car-engine for


the ambient temperature


blind is an open position, temperature of coolant in engine has a value 

approach to steady states. In this moment the action of the blind is ended. 

2. The controller can be with the electronic, electromagnetic switches or 

the reliable programmable logic controller–PLC. 

3. The servomechanism for cooling air flow control, proposed in the 

paper, has a robust structure and limited mechanical angular displacement of the 

rockers. 

4. The design and adjustment of bi-positional controller is realized 

through utilizing of the experiment results. 

REFERENCES 

Dăscălescu S.C.D., Receanu M., Reduction Fuel Consumption Possibilities to Motor 

Vechicles by Use of Cooling Air Flow Devices. ESFA. The 7 th International 

Conference–Fuel Economy, Safety and Reability of Motor Vehicle, 2003. 

Dăscălescu S.C.D., Receanu M., Servomecanism de control al debitului de aer pentru 

răcirea radiatorului unui motor termic,RO–BOPI 4/2012, 30.04.2012, p. 36, 

Buletinul Oficial de Proprietate Industrială, Secţiunea de Brevete de invenţie 

(2012). 

Receanu Marius, Dăscălescu S. C. D., Fuel Economy of the Car-engine Using the 

Variation of Cooling Air flow. Bul. Inst. Polit. Iaşi, LVI(LX), 4B, s. Construcţii 

de Maşini, 67-75 (2010). 

OPTIMIZAREA CONTROLULUI DEBITULUI DE AER LA UN SISTEM 

DE RĂCIRE AL UNUI MOTOR TERMIC 

(Rezumat) 

Servomecanismul pentru controlul optim al debitului de aer de răcire al 

radiatorului, prezentat în lucrare, funcţionează când temperatura lichidului de răcire care 

îl traversează este sub o anumită valoare prescrisă, când autoturismul pleacă după o 

staţionare şi temperatura mediului ambiant este sub +5 ˚C. Când clapetele dispozitivului 

de obturare sunt deschise, în timpul deplasării autoturismului, temperatura Tm lichidului 

de răcire care traversează motorul termic a ajuns, pentru început la o temperatură 

apropiată de cea de regim. În acest moment sarcina servomecanismului s-a încheiat, iar 

temperatura Tm începe să crească până când este comandat, de exemplu, motorul electric 

al ventilatorului la o turaţie mai mare. Funcţionarea servomecanismului este 

independentă de cea a unităţii electronice de control a autoturismului (ECU) şi poate fi 

comandată de sisteme secvenţiale cu relee electronice, electromagnetice sau automate 

programabile (PLC). Construcţia servomecanismului, propusă în lucrare, este una 

robustă din punct de vedere mecanic şi cu mişcări limitate ale mecanismelor 

componente. Proiectarea şi reglarea sistemului de control automat bipoziţional s-a 

realizat pe baza încercărilor experimentale expuse în lucrare.


Publicat de 



Secţia 


MODELING A CAR POWERTRAIN SYSTEM WITH STABLE 

OPERATION 

BY 

FlORIN POPA, EDWARD RAKOSI * and GHEORGHE MANOLACHE 

“Gheorghe Asachi” Technical University of Iaşi, Romania, 

Department of Automotive and Mechanical Engineering 

Received: 5 March, 2012 

Accepted for publication: 15 March, 2012 

Abstract. From a mathematical model, previously developed by the 

authors, has done a simulation by inserting subroutines in the software package 

MATLAB 6.5 and a CATIA V5R16 environment modeling. By customizing the 

model obtained can define it completely stabilized operating modes of 

propulsion systems and their dynamic, economic and pollution parameters. This 

provides, even the design phase, definition of additional constructive and 

functional criteria against the current, to ensure stable and economic operation, 

in an area as extensive. 

Key words: automotive powertrain, stable and economic operation. 


Starting from the mathematical model developed has made a complex 

simulation of the propulsion system behavior in various constructive-functional 

situations, called MATCEL-PROP. The simulation was performed by inserting 

subroutines in the software package MATLAB 6.5 and Excel program. Adding a 

tutorial conducted to facilitate the development program. They could thus reveal 

a series of theoretical results, described briefly below. 

Determination of basic parameters required simulation was made for a 

typical car, VW Golf IV; significant values are highlighted in Fig. 1. 

Corresponding author: e-mail: edwardrakosi@yahoo.com

162 Florin Popa et al. 

Fig. 2 highlights characteristic of the propulsion system model is chosen, 

it is the actual power and timing variations in the actual engine propulsion 

system obtained by simulation, in the form of an executable. 

Fig. 1 – Basic parameters related coefficients picture. 

Fig. 2 – Characteristic of the propulsion system.


2. Simulations 

From these raw elements, further presented the most significant results of 

simulation MATCEL-PROP, looks especially the problem of propulsion system 

stability. Were considered, in particular, common situation of equipment for the 

real propulsion system model of car chosen, designed to lead to some 

conclusions with applicative features, useful in terms of its mechanical and 

functional optimization (Heywood, 1988). Results are summarized using the 

variation diagrams of engine torque, Mm, resistant torque reduced to motor 

shaft, Mrr and specific fuel consumption, ce, depending on the angular velocity 

ω of motor shaft, presented in figures below (Heisler, 2002),. For each case, 

algorithm simulation allows determining the position of the system operating 

point corresponding to maximum speed of the car. On the other hand, in order 

to emphasize the stability of propulsion, these diagrams contain the tangent 

lines to variation curves of torques in the operating point, Fct.tan.Mm, 

respectively, Fct.tan.Mrr. 

In this respect, a first case simulated and analyzed (ES1) with MATCEL- 

PROP considers basic equipment of the propulsion system, which includes an 

engine having the effective power value 64 kW, assimilated in the model with 

an engine power (Pe=Pm), and a 15" wheel diameter. By assigning the value 4 

for main gear ratio, i0, variations in Fig. 3 shows a stable operating point at 

maximum speed of 149 km/h, and further characterized by an acceptable 

specific effective fuel consumption, respectively 173 g/kWh. 

M (Nm) 

140 

120 

100 

80 

60 

40 

20 

stable operation i0=4 

Pe=64 [kW], d=15" 

1200 2200 3200 4200 5200 6200 n (rpm) 

160 

230 

220 

210 

200 

190 

180 

170 

160 

0 

150 

126 230 335 440 545 649 

ω (rad/s) 

ce (g/kWh) 

Mm 

Mr reduced 

Fct.tan.Mm 

Fct.tan.Mrr 

Fig. 3 – Conditions for stable operation of the propulsion 

system - highlighted by the simulation phase ES1. 

Still increasing main gear ratio by only 0.5, thus leading to the value 

i0=4.5, the propulsion system behavior becomes unstable, speed decreases 

slightly the value of 148 km/h, there the premises so that it can not be kept 

constant, while the value of specific effective fuel consumption rises to 173 

g/kWh to 186.4 g/kWh, situation highlighted by the diagrams in Fig. 4. 

On the other hand, from baseline, summarized in Fig. 3, reducing by 0.5 

the ratio main gear i0, it reaches i0=3.5, achieve a very stable operating point, 

leading to maximum speed 146.9 km/h, very stable value, made with a specific 

ce


effective fuel consumption of 162 g/kWh, which confirms the downward trend 

of this important parameter, with increasing stability of the system drive. This 

condition simulated propulsion system is presented through results obtained in 

Fig. 5. 

M (Nm) 

120 

100 

80 

60 

40 

20 

unstable operation i0=4.5 

Pe=64 [kW], d=15" 

1200 2200 3200 4200 5200 6200 n (rpm) 

140 

230 

220 

210 

200 

190 

180 

170 

160 

0 

150 

126 230 335 440 545 649 

ω (rad/s) 

ce (g/kWh) 

Mm 

Mr reduced 

Fct.tan.Mm 

Fct.tan.Mrr 

Fig. 4 – Conditions for unstable operation of the propulsion 


M (Nm) 

200 

150 

100 

50 

very stable operation i0=3.5 

Pe=64 [kW], d=15" 

1200 2200 3200 4200 5200 6200 n (rpm) 

250 

230 

220 

210 

200 

190 

180 

170 

160 

0 

150 

126 230 335 440 545 649 

ω (rad/s) 

ce (g/kWh) 

ce 

Mm 

Mr reduced 

Fct.tan.Mm 

Fct.tan.Mrr 

Fig. 5 Conditions for very stable operation of the propulsion 


Continuing the iteration of the MATCEL-PROP, the main gear ratio value 

i0=3 is reached extremely stable operation of the propulsion system, 

characterized by a maximum speed value of 136.3 km/h, in turn extremely 

constant value, in fact the smallest of speed values highlighted in this stage of 

the simulation and effective specific fuel consumption decreased to 155.6 

g/kWh, situation presented in Fig. 6. 

Simulate a situation leading to opposite results as shown in Fig. 7, the 

i0=5 value ratio of the main gear reach a very unstable operation, the operating 

point is characterized in this case by the speed increased to 144.9 km/h, but can 

not be maintained, obtained at engine speed of 6200 rpm and effective specific 

fuel consumption increased to a value of 200.4 g/kWh. 

ce


M (Nm) 

300 

250 

200 

150 

100 

50 

extremely stable operation i0=3 

Pe=64 [kW], d=15" 

1200 2200 3200 4200 5200 6200 n (rpm) 

350 

230 

220 

210 

200 

190 

180 

170 

160 

0 

150 

126 230 335 440 545 649 

ω (rad/s) 

ce (g/kWh) 

Mm 

Mr reduced 

Fct.tan.Mm 

Fct.tan.Mrr 

Fig. 6 – Conditions for extremely stable operation of the propulsion 


M (Nm) 

120 

100 

80 

60 

40 

20 

very unstable operation i0=5 

Pe=64 [kW], d=15" 

1200 2200 3200 4200 5200 6200 n (rpm) 

140 

230 

220 

210 

200 

190 

180 

170 

160 

0 

150 

126 230 335 440 545 649 

ω (rad/s) 

ce (g/kWh) 

ce 

Mm 

Mr reduced 

Fct.tan.Mm 

Fct.tan.Mrr 

Fig. 7 – Conditions for very unstable operation of the propulsion 


Simulation performed with MATCEL-PROP include a second phase of 

study (ES2), which takes into account the above stable conditions, characterized 

by the motorization of 64 kW and the main gear ratio value i0=4, but with twostage 

change diameter wheels. Thus, a diameter of 15"was increased to 16". 

The third phase of study (ES3) addressed through simulation MATCEL- 

PROP envisages an engine characterized by increased power to 74 kW, keeping 

the basic diameter of the wheel motors, thus is 15", but with the i0 main gear 

ratio change, with steps of 0.5 in the range of normal values. Stable operation is 

achieved with the main gear ratio i0=3.7. 

The simulation is MATCEL-PROP addresses equally the fourth stage of 

the study (ES4), leaving, primarily from reduced engine power, thus is 54 kW 

and secondly, as described above, driving wheel diameter 15". Also, in this case 

is considered the i0 main gear ratio change with steps of 0.5. Initializing the 

study from i0=4.2 is reached to a stable operating state. The results obtained in 

the four stages of the study are summarized in Fig. 8, Fig. 9, Fig. 10 and Fig. 11, 

actually describing simulated behavior of the propulsion system, assessed by the 

stability parameter variation Δ . 

ce


Fig. 8 – Variation of stability parameter with main gear ratio change obtained by 

simulating MATCEL-PROP for propulsion engine power of 64 kW. 

Fig. 9 – Variation of stability parameter obtained by simulating 

MATCEL-PROP for changing wheel diameter. 

Fig. 10 – Variation of stability parameter with main gear ratio change obtained 

by simulating MATCEL-PROP for propulsion engine power of 74 kW. 

Fig. 11 – Variation of stability parameter with main gear ratio change obtained 

by simulating MATCEL-PROP for propulsion engine power of 54 kW.


On the other hand, the evolution of specific fuel consumption values for 

different states of the propulsion system is highlighted in Fig. 12. It can thus be 

noted that increased stability during operation of the propulsion system, values 

of these inputs are reduced, which is very advantageous. 

effective specific fuel consumption, ce [g/kWh] 

190 

185 

180 

175 

170 

165 

160 

155 

unstable operation 

stable operation 

very stable 

operation 

Fig. 12 – Evolution of effective specific fuel consumption values 

for different states of the propulsion system. 

Using the Fig. 13 a,b,c shown the variation of these inputs for the three 

states of the system, and that each state consumption values are very close 

together, being practically constant regardless of equipment and characteristics 

of propulsion system that obtains the condition of its. 

173 

180 

160 

140 

120 

100 

80 

60 

40 

20 

171.16 0 

173 

a. 

166.2 

186.4 

200 

180 

160 

140 

120 

100 

80 

60 

40 

20 

183.9 0 

186.4 

161.1 

181.5 

180162 

160 

140 

120 

100 

80 

60 

40 

20 

0 

c. 

Fig. 13 – The range of values of effective specific fuel consumption 

corresponding of the three states of propulsion system. 

b. 

161



1. Based on MATCEL-PROP simulation, design using CATIA V5R16 

environment has created a virtual model of transmission (Sun, Kolmanovsky, 

Cook & Buckland, 2005), as a complex part of the propulsion system for the 

most stable operating situation highlighted. It was also considered the modeling 

of dynamic phenomena mainly in order to validate the equivalent mechanical 

model of the propulsion system. 

2. Since the equivalent model substitute the presence of trees, gears and 

other rotating masses, specific to propulsion system, also was considered useful 

an evaluation of a discrete model of the propulsion system, encompassing such 

elements assimilated in mechanical model. 

3. Thus, the Fig. 14 is presented all the rotating masses of real gearbox 

model consisting of a mechanical gearbox with fixed axes, with 5-speed 

synchronized. 

Fig. 14 – Virtual model for meshing the principal elements 

of propulsion system based on the simulation performed. 

REFERENCES 

Heisler H., Advanced Vehicle Technology, Elsevier Science. Reed Educational and 

Professional Publishing, 2nd edition, 2002. 

Heywood J. B., Internal Combustion Engine Fundamentals. McGraw-Hill Series in 

Mechanical Engineering, Library of Congress Cataloging-in-Publication, 1988. 

Sun J., Kolmanovsky I., Cook J.A., Buckland J.H., Modeling and Control of Automotive 

Powertrain Systems: A Tutorial. Proceedings of the American Control 

Conference, June 8-10, Portland, OR, USA, 5, 2005, pp. 3271-3283


MODELAREA UNUI SISTEM DE PROPULSIE AUTO CU FUNCŢIONARE 

STABILĂ 

(Rezumat) 

Lucrarea prezintă un model virtual al sistemului de propulsie generat de autori în 

mediul de proiectare CATIA V5R16 şi elaborarea programului de simulare MATCEL- 

PROP. 

Prin utilizarea modelului virtual s-a căutat zona de maximă stabilitate în 

funcţionare pentru sistemele de propulsie auto. 

La definirea acestui model s-a luat în considerare şi posibilitatea de modelare a 

fenomenelor dinamice, în principal cu scopul de a valida modelul mecanic echivalent al 

sistemului de propulsie. 

Deoarece modelul echivalent substituie prezenţa arborilor, roţilor dinţate şi a 

altor mase de rotaţie, specifice sistemului de propulsie, a fost de asemenea considerată 

utilă o evaluare a modelului discret al sistemului de propulsie care să cuprindă aceste 

elemente asimilate în modelul mecanic. 

Prin personalizarea modelului obţinut se pot defini regimurile de funcţionare 

stabilă a sistemele de propulsie auto precum şi parametrii lor dinamici şi economici încă 

din faza de concepţie.


Publicat de 



Secţia 


DETERMINATION OF PROPULSION SYSTEM COMPONENTS 

ON BOARD A TANKER SHIP 

BY 

MARIANA LUPCHIAN * 

“Dunărea de Jos” University of Galaţi, 

Department of Thermal Systems and Environmental Engineering 



Abstract. The paper presents the calculation of propulsion system 

components for different regimes of operation. Choosing the type of 

propulsion system must be the result of a technical-economic analysis that 

takes into consideration all factors that depend on the safety ship and 

economically operation. Propulsion performance are determined for linear 

movement of the ship. The ship must carry the parameters for which it 

was designed and built, thus satisfying all the technical aspects and 

economic. 

Key words: energetic plant, propeller, deadweight, Diesel. 


Power plants with internal combustion engines, which particularly 

become modern automatic systems, cannot be conceived without taking into 

consideration the efficiency degree, their framing in minim power, fuel, 

material or time consumption. 

The shape and size of the ship is of particular importance on the 

resistance to progress. 

Actual marine engines with compression ignition operating regimes of 

power and speed variables, implying change parameters indicate, effective, 

characterizing the engine operating regime. 

* e-mail: Mariana.Lupchian@ugal.ro

172 Mariana Lupchian 

Propulsion engines work in different operating conditions determined by 

the technical condition of the ship and propulsion plant and external factors that 

influence the operation. Conducting actual work processes of an engine is 

accompanied by the appearance of additional losses that adversely affect engine 

power and economy. 

For naval propulsion plant with internal combustion engines is 

considered as independent variables which give its operating regimes by the 

functional characteristics of internal combustion engine, power transmission 

characteristics, and consumer characteristics. 

2. Problem Formulation 

2.1. Calculation of Propulsion System Components 

Advance relative propeller is 

va 

J = , (1) 

n D 

el el 

Where: va – feed speed of the propeller; J – the relative advance of propeller in 

water; Del= 5.8 m – the propeller diameter; nel – the propeller rotation rot/s. 

v = v(1 − w) 

, (2) 

a 

where v m/s – the ship speed; w – shipwake coefficient; H/D m – the relative 

pitch of the propeller; zelice = 1 – number of propellers. 

4 

T = k ρn D [N], (3) 

2 

1 el el 

where k1 – the coeficient of pushing the propeller; ρ – density of water kg/m 3 . 

RT 

T = [N]; (4) 

( 1− 

t) 

where t – coefficient of suction; RT – ship resistance in advancement, kN; T – 

pushing propeller, N; Mel – torque of the propeller, Nm; k2 – the torque 

coefficient, 

2 5 

el 2 e el 

Propeller efficiency calculated with 

M = k ρn D . 

(5) 

J k1 

η0 

= . (6) 

2π 

k 

2


3. Case Study 

Characteristics of the ship: petroleum is equipped with a propeller, the 

propulsion of the vessel being provided by a diesel engine MAN B & W & 6cylinder. 

Engine power: 9480 kW, 127 rpm; deadweight in the sea water is 37000 

tdw, speed of service: 15 Nd. The ship is equipped with three Diesel generators 

each with many 6 cylinder in-line power of 960 kW, speed 900 rpm. 

Deadweight in the sea water is 37000 tdw. The manufacturer is running a 

trial race at full load, ship ballast water is high, the draft of 10.50 m, the draft 

load is 11 m. The ship is equipped with a fixed pitch propeller with four blades: 

DW – deadweight 37000 t; L,B,H – lenght, breadth, height of the construction; L 

– length overall: 179.96 m; B – beam of ship: 32.20 m; H – height 16.50 m; w – 

ship wake coefficient; zp – the number of the propeller’s blades; ηel – propeller’s 

efficiency in open water; D – the propeller diameter =5.8 m. 

Table 1 

Results obtained (ship loaded) 

Nr crt. v , [Nd] n , [rot/min] va , [m/s] t η0 J 

1. 9.448 80 3.3514 0.242 0.4641 0.4334 

2. 10.629 90 3.7773 0.242 0.4928 0.4342 

3. 11.810 100 4.2036 0.242 0.5161 0.4348 

4. 13.700 116 4.8868 0.242 0.5329 0.4348 

5. 14.999 127 5.3571 0.242 0.5249 0.4363 

6. 15.483 131 5.5325 0.242 0.5167 0.4369 

Table 2 

Results obtained (navigation in ballast) 

Nr. crt. v, [Nd] n ,[rot/min] va , [m/s] t η0 J 

1. 9.448 80 3.3049 0.243 0.574624 0.4273 

2. 10.629 90 3.7258 0.243 0.583859 0.4282 

3. 11.810 100 4.1473 0.243 0.586627 0.4290 

4. 13.700 116 4.8227 0.243 0.569931 0.4301 

5. 14.999 127 5.2878 0.243 0.540355 0.4307 

6. 15.483 131 5.4613 0.243 0.524893 0.4313

174 Mariana Lupchian 

Fig. 1 – The propeller efficiency of naval propulsion (ship loaded). 

Fig. 2 – The propeller efficiency of naval propulsion 

(navigation in ballast) 


1. The result is maximum efficiency of the propeller η0 = 0.586627, ship 

speed v = 11.810 Nd for navigation in ballast. The propeller efficiency η0 = 

=0.5329, ship speed v = 13.700 Nd to load navigation.


2. When designing the propulsion is looking to reduce the size of the 

compartment size of machines, to increase the size warehouses for goods 

transported and increase passenger space, but taking into account the records of 

the prescriptions governing class sizes for convenient and safe exploitation of 

the energy of the ship. 

3. The problem of achieving the energy plant with internal combustion 

engines with advanced technical, thermodynamic and high operating with their 

deployment of profitable and competitive economic activities. 

REFERENCES 

Bidoaie I., Sârbu N., Chirica I., Ionaş O. Indrumar de proiectare pentru teoria navei. 

Universitatea “Dunarea de Jos”, Galaţi, 1986. 

Caraghiulea M. (Lupchian), Exergetic Analysis of a Naval Propulsion Plant With 

Internal Combustion Engine Proceedings Of The Internationally Attended 

National Conference on Thermodynamics, Vol. 2C, Braşov, 2009, pp. 311-314. 

Caraghiulea M. (Lupchian), Thermoeconomic Optimization for a Naval Propulsion 

Plants with Internal Combustion Engines by Using Nonlinear Programming 

Methods. Bul. Inst. Polit. Iaşi, LIV(LVIII), 2, s. Secţia Construcţii de maşini, 

129-133 (2008). 

Ceangă V., Mocanu C. I., Teodorescu C, Dinamica sistemelor de propulsie. Ed. 

Didactică şi Pedagogică, Bucureşti, 2003. 

Dumitru Gh. Maşini şi instalaţii de propulsie navale. I, 1, I, 2, Universitatea din 

Galaţi, 1979. 

Simionov M., Instalaţii de propulsie navală. Ed. Universităţii din Galaţi, 2009. 

DETERMINAREA COMPONENTELOR SISTEMULUI DE PROPULSIE LA 

BORDUL UNUI PETROLIER 

Lucrarea prezintă calculul componentelor sistemului de propulsie pentru diferite 

regimuri de funcţionare ale instalaţiei energetice navale. Calculul este făcut pentru 

navigaţia petrolierului în balast şi pentru cursa cu încărcătură. 

Alegerea tipului de sistem de propulsie trebuie să fie rezultatul unei analize 

tehnico-economice care să ia în considerare toţi factorii de care depinde siguranţa navei.


Publicat de 



Secţia 


INTRODUCTION TO COMMON RAIL INJECTION SYSTEM 

BY 

VASILE GABRIEL NENERICA *1 , VASILE HUIAN 2 and DORU CĂLĂRAŞU 1 


Accepted for publication: 28 March 2012 


1 Department of Hydraulic, 

2 Department of Mechanical Engineering 

Abstract. The Diesel engine is known to generate great impellent/ 

propulsion and is becoming increasingly sought at the moment. The emergence 

of injection systems of gasoline engines and the requirements of emissions' 

reduction conducted to new solutions development so the Diesel engine can be 

used for trucks and cars, alike. These facts seem to have solved the exigencies in 

the development of so called Common Rail injection system. The present 

scientific approach is focused on Common Rail injection systems, in general and 

on mainly solenoid injectors, in relation with certain functional problems and 

correction factors 

Key words: Diesel engine, common rail, injector, cavitations. 


Injection equipment is designed to feed the engine combustion chamber 

with fuel, so that burning meets at any time the engine operating regime, 

determined to turn its external load. Diesel engine operation is based on selfignition 

of fuel injected and sprayed into the engine cylinders when the intake 

air above reaches, by compressing the piston cylinder, a temperature sufficient 

to produce self -ignition. For proper and economic operation engine at the same 

time, injection equipment must meet several requirements, among which the 

most important are: 

* Corresponding author: e-mail: nnrcvasile@yahoo.com

178 Vasile Gabriel Nenerica et al. 

i) To raise fuel pressure to a specified value and to spray into the 

combustion chamber so that air and fuel mixture is as good as possible and 

combustion is more complete. 

ii) Fuel injection-start at some point and finish within a well established 

time. 

iii) Fuel-injection to be made in accordance with the engine combustion 

process in terms of position and shape of the jet. 

iv) To inject an appropriate amount of fuel to the engine load at any time. 

Since fuel injection should be done with great precision, both in volume 

and time, and at a very high pressure, it is required that the clearances among 

moving parts that are pumping circuit be very small (1-3 microns). This leads to 

the need of a better execution of the pumping element , the relief valve and the 

spray especially with the highest precision encountered in engineering. As a 

result, size, shape and mutual position deviations of these parts work surfaces 

are prescribed and produced in very tight limits, and the same surface roughness 

is very small that the injection equipment is manufactured at high precision, 

which requires special attention from the operation and repair units. 

Fuel injection equipment is made up, in the order passed through by the 

fuel, of the following: fuel tank, fuel pump, low pressure pipes, filter battery 

(pre-cleaning filter and final filter), the injection pump with regulator, high 

pressure pipes, injection and spraying. 

The main functions of the high pressure system are provided by the high 

pressure injection pump. Thus, the pressure of injection, the dosage amount of 

fuel per cycle and cylinder, the injection advance, as well as the optimum 

injection characteristics during injection are achieved by the injection pump. 

Injection pumps are classified according to several criteria: 

І. Depending on the service cylinder engine there can be distinguished: 

a) individual pumps; 

b) injector pump; 

c) rotary distributor pumps; 

d) pump line, characteristic of this class is that each cylinder of the 

engine is served by one outlet item. 

II. After dose adjustment method of diesel: 

a) invariable aspiration and partial discharge (eg, piston pumps, gate); 

b) variable intake and total discharge (rotary distributor pumps); 

III. By mode of operation of the pumping element: 

a) mechanical (cam); 

b) electromagnetic actuator;


The main problem with injection pumps is to reach high injection 

pressures. Values up to 140 MPa maximum injection pressure can be ensured 

only by piston pumps. Size injection pressure implies high requirements to 

accuracy of execution of the piston and cylinder pumping element as well as to 

sealing the couple parts from the external environment. These requirements 

have led to reduction of functional clearances between piston and cylinder to 

values of 1.5 ... 3.0 μm and production of certain construction with increased 

plunger length compared to its diameter. 

This implementation involves fine grinding operations, with extremely 

limited deviations in form, surface quality and their mutual position, and 

running operations of the piston and cylinder mating. Piston-cylinder torque 

thus obtained has the component parts interchangeable. 

2. Common Rail System 

Common rail is a direct injection type, used in internal combustion 

engines, particularly with compression ignition engines. The most important 

aspect of a common rail engine is that, distribution of fuel to the injectors is a 

main pipeline (hence the common rail) with high pressure to each injector 

separately. 

The basic idea of the system is creating pressure for injection to occur 

independently of the engine speed, and fuel pressure at low speed is still 

maximum. 

The so called pressure, common rail, has meantime reached that the 

German Bosch company and Japanese DENSO to build power pumps that can 

sustain pressures up to 2000 bar, presented in 17/09/2007 at IAA (Internationale 

Automobil-Ausstellung) Frankfurt. 

This system, developed after more than 15 years, but was brought to 

commercial use only a few years ago. The system was manufactured by Fiat in 

the '80s, but those who have implemented for the first time in a vehicle for sale, 

was Mercedes. After them, others have followed, one of them being Peugeot 

that in 2000 had spectacular sales increases on vans equipped with such system. 

The visible difference from the old system is that fuel injection does not go 

directly to the injectors to pump but it builds a common platform and from there 

goes directly to the injectors. 

Development of common rail injection system was done trying to follow 

the principles of gas injection. Many ideas are taken from this system. 

The advantages that this system brings from previous versions of diesel 

injection systems are: 

1. Harm reduction due to better fuel spray into cylinder and due to better 

control over the amount of fuel entering the cylinder during injection, CO2 

emissions be reduced by up to 20% and carbon monoxide level is reduced by up 

to 40%. Unburnt hydrocarbons are decreased by 50%.


2. Better fuel spraying is due to high pressure injection that is achieved 

during injection (allow the use of same smaller nozzle holes). Such a method is 

shown in Fig. 1. 

Fig.1 – How to increase flowrate (www.delphi.com). 

In the old injection system injection pressure was limited by structural 

considerations as follows: the long distance of pipelines from the fuel pump to 

the injection pressure limited the increase because of the debt resonance 

phenomena which could occur in the pipeline. Even when the engine is stopped 

the fuel from common rail pressure is the same so that when it is needed to be 

sent to the injector in the shortest time. As a result a second ignition called post 

combustion is now possible by using a small amount of diesel fuel in the main 

explosive phase on loads of fuel injectors.The result is that it eliminates unburnt 

diesel particles after the first explosion reducing hazards through a catalytic 

heating installation faster. 

3. Reduction of fuel consumption, a doubling of torque at low revs and 

increases up to 25% of engine power. It also reduces noise and vibration level 

specific to Macs. 

4. Better control of injection timing leads to a quieter running engine, a 

better performance with a lower consumption; 

5. High reliability of engines equipped with such a system. 

The system consists of a fuel tank, electric fuel pump (which may exist or 

not) that serves to bring fuel from the tank through a filter to the high-pressure 

pump, the high pressure pump, a common rail, pipeline and injectors. On cars 

fitted with electric pump, when running out of fuel (it was a big problem of


diesel engines) the pump does not have to be primed because high-pressure 

pump is primed due the electric pump. To those not equipped with electric 

pump, there is a manual pump located by the filter cartridge and which is 

designed to prime the high-pressure pump. However, there are cars that have 

none of them, and then the ignition has to be achieved through old methods (eg 

Opel Astra 1.7 CDTI has not got such a pump, Renault Megane 1.5 DCI has 

got a manual pump). 

From filter cartridge, fuel reaches to high pressure pump. The role of 

high-pressure pump is to raise fuel pressure to around 2000 bar a wide range of 

speeds (Fig. 2). 

Fig. 2 – Pressure variation depending on the speed. 

Fig. 3 – Common rail system (www.delphi.com).


where 1 – the unit dose; 2 – the temperature sensor; 3 – the high pressure pump; 

4 – the pressure sensor; 5 – the injector; 6 – the Venturi; 7 – the valve; 8 – the 

rail; 9 – the fuel filter 

The values vary from one engine type to another, pressure reaching from 

1500 up to 2500 bar. The high-pressure pump is designed to transport fuel from 

the tank through a transfer pump that is built into the high pressure pump body. 

The fuel, leaves the pump through a unique pipe leading to the central rail. 

The Ramp pressure of fuel taken into account by the pressure sensor in 

the common rail pipe is then sent like signal to the engine control unit. 

Common rail pressure is regulated by the dosage unit. Dosage unit is 

commanded by the engine control unit. 

Injection period begins with component command injection via the 

engine control unit. 

Injection takes place as long as the injector is controlled. 

Current common rail systems allows the multiple division of injection 

for an improved combustion process. 

This common rail, feeds injectors through special pipes for high pressure, 

equal in length so that the pressure does not differ to injectors. Precision 

injectors or piezo-electric are controlled electronically commanded by the 

computer engine ECU (Electronic Control Unit) which can inject fuel up to six 

times per cycle combustion cylinder. Controlled injection can be performed at 

will, before, during, and after the burning, through which a lower noise is 

achieved, a reduction of NOx, CO2, CO, hydrocarbon according emissions 

standards (Fig. 4) and a better combustion for soot. 

Fig. 4 - European emissions standards [www.delphi.com] 

One of the most complex components of the injection sistem is Injector. 

Because it operates at very high pressures, 1600…2200 bar, the clearances are 

very small, micron, and require processing with the latest technology and 

assembly to be in rooms where humidity, temperature and cleanliness are 

controlled. Injector is very complex and requires a well developed series of 

tests.


A very important factor in direct injection diesel engines is the quality of 

the diesel spray. An improvement in the air fuel mixing leads to a better 

combustion process that results in higher performance, but also modifies 

pollutant emissions (Pierpont&Reitz, 1995). The spray characteristics are 

strongly influenced by the nature of the flow inside the injection nozzle hole 

(Sazhin et al., 2001, Pavri et al., 2004). This internal flow is largely dependent 

on the presence of the cavitation phenomenon (Soteriou et al., 1995, Chaves et 

al., 1995). On the other hand, momentum flux is a very important parameter for 

predicting the mixing potential of injection processes. Despite this, there are 

very few studies that use this kind of measurement (Kampmann et al., 1996; 

Siebers, 1999). Important factors such as spray penetration, spray cone angle 

and air entrainment depend largely on spray momentum (Rajaratnam, 1974). 

With the simultaneous determination of the momentum and mass flux it is 

possible to determine the velocity in the outlet section of the nozzle hole. In Fig. 

5, point i corresponds to the upstream point where velocity could be neglected. 

Point b corresponds to the outlet section of the hole. (Payri et al., 2004). It can 

be seen in Fig. 5 how the cavitation influence injection. 

Fig. 5 – Fluid separation inside a Diesel nozzle hole. 

The phase transition from liquid to vapour of a fluid due to low pressure 

is called cavitation. Cavitation in a diesel nozzle appears at the inlet of the 

nozzle hole. In this region, and due to the strong change in cross-section and 

flow direction, the boundary layer tends to separate from the hole wall and a socalled 

“vena contracta” is established. As a consequence, a recirculation zone 

appears between the vena contracta and the orifice wall. In this zone, there is a


pressure depression due to the acceleration of the fluid. If the static pressure 

falls below vapour pressure then the phenomenon of cavitation will take place 

(Bergwerk, 1959). 

This is one of the most dangerous and thus one of the most difficult 

problems of the injectors. This problem was solved by changing the geometry 

of the spray holes. In the case of the cylindrical nozzle, the minimum pressure 

suddenly falls to a minimum value of 0 MPa in an area that can be located at the 

upper corner of the orifice inlet. Considering that the vapour pressure is 0.08 

MPa, the critical pressure conditions of the onset of cavitation are achieved for 

the cylindrical nozzle. Nevertheless, for the convergent nozzle the minimum 

pressure is 6 MPa at the orifice exit, thus avoiding any cavitation phenomenon. 

Therefore, considering this definition, a critical cavitation number of Kcrit=1.25 

is found for the cylindrical nozzle which is the same as the value found 

experimentally for the pressure of 30 MPa in Fig. 6. Nevertheless, in the case of 

the convergent nozzle, even with an injection pressure of 130 MPa, critical 

cavitating conditions were not reached, as experiments have shown (Payri et al., 

2004). 

Fig. 6 – Results of the pressure field for the cylindrical (left) and conical (right) nozzle. 

Pi=30 MPa, Pb=6 MPa. 


1. Common rail injection system can operate at very high pressures and 

thus combustion is much better, so pollution is much lower 

2. With the advent of new processing technology and games could shrink 

the componentsand diameter spray holes leading to increased pressure injection. 

3. The cylindrical nozzle, cavitation was detected due to a collapse of 

mass flow. 

4. This resultmeans that the discharge coefficient will decrease when the 

cavitation coefficient decreases.


5. Conical nozzle cavitation effects have not even under severe pressure, 

which was explained by means of numerical simulations. 

REFERENCES 

Bergwerk., Flow Pattern in Diesel Nozzle Spray Holes. Proc. Inst. Mech. Eng., 173(25) 

(1959). 

Chaves H., Knapp M., Kubitzek A., Experimental Study of Cavitation in the Nozzle 

Hole of Diesel Injectors Using Transparent Nozzles. SAE Paper 950290 (1995). 

Kampmann S., Dittus B., Mattes P., Kirner M., The Influence of Hydro at VCO Nozzles 

on the mixture preparation in a DI diesel engine. SAE Paper 960867 (1996). 

Payri F., Bermudez V., Payri R., Salvador F.J., The Influence of Cavitation on the 

Internal Flow and the Spray Characteristics in Diesel Injection Nozzles. Fuel, 83, 

419-31 (2004). 

Payri R., Garcia J.M., Salvador F.J., Gimeno J., Using Spray Momentum Flux 

Measurements to Understand the Influence of Diesel Nozzle Geometry on Spray 

Characteristics. Science Direct Paper Fuel, 84, 551-561 (2005). 

Pierpont D.A., Reitz R.D., Effects on Injection Pressure and Nozzle Geometry on DI 

Diesel Emissions and Performance. SAE Paper 950604 (1995). 

Rajaratnam N., Turbulent Jets. Elsevier, Amsterdam, 1974. 

Sazhin S.S., Feng G., Heikal M.R., A Model for Fuel Spray Penetration. Fuel, 80,2171- 

2180 (2001). 

Siebers D., Scaling Liquid-phase Penetration in Diesel Sprays Based on Mixing-limited 

Vaporization. SAE Paper 1999-01-0528 (1999). 

Soteriou C., Smith M., Andrews R., Direct Injection Diesel Sprays and the Effects of 

Cavitation and Hydraulic Flip on Atomization. SAE Paper 950080 (1995). 

INTRODUCERE ÎN SITEMUL DE INJECŢIE COMMON RAIL 

(Rezumat) 

Se prezintă sistemul de injecţie Common Rail, adică rampa comună, componenta 

cât şi modul de funcţionare. De asemenea, am prezentat şi una din cele mai dificile 

probleme ale injectorului acestui sistem de injecţie şi anume cavitaţia. Acest fenomen 

este prezent în interiorul injectorului dar cel mai frecvent apare la orificiile de 

pulverizare. Fenomenul de cavitaţie poate fi evitat prin modificarea formelor geometrice 

ale orificiilor, şi anume din orificii cilindrice în orificii conice. Experimentele au arătat 

că astfel fenomenul de cavitaţie nu mai apare şi mai mult de atât, forma jetului este mult 

mai bună.


Publicat de 



Secţia 


STUDY OF A CHROMOPLASTIC MATERIAL SHEAR 

BREAKING 

BY 

COSTICĂ ATANASIU ∗ 1 , BOGDAN LEIŢOIU 2 and ŞTEFAN SOROHAN 3 

Received: May, 26, 2012 

Accepted for publication: June 30, 2012 

1 Technical Sciences Academy from Romania 

2 "Gheorghe Asachi" Technical University of Iaşi 

3 "Politehnica" University of Bucharest 

Abstract. The chromoplastic material is a compounded polymer prepared 

at the Chemical Research Institute from Bucharest. A tensile test characteristic 

type Prandtl was obtained for this material. Also, a color changing was observed 

at the occurrence of the plastic joint, whiten in color when a tensile flow, and 

darken in color when a compression load flow is producing. 

Authors present pure shear tests results researches on chromoplastic 

material specimens, processed according to ASTM requirements, which is based 

on Iosipescu shear test method. The elasticity theory relations are used and FEM 

numerical verifications are carried though to determine the shear broken stress, 

crack initiation and fracture section shape. 

This work shows benefits of using chromoplastic materials in elastic-plastic 

and plastic domain researches. 

Key words: chromoplastic material, pure shear, plastic domain, electrical 

resistive strain transducers, finite elements. 


Strength and rigidity, as well as the industrial or civil structures integrity 

calculations fulfilling, requires the knowledge of mechanical and elastic 

characteristics of materials and load applied. New materials issuing and 

numerical analysis methods development of the stress and strain state by 

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

188 Costică Atanasiu et al. 

computer programs, requires knowledge of these characteristics, which are 

experimentally determined. 

Iosipescu has established the pure shear test specimens shape (Iosipescu, 

1962; 1963), the way to accomplish the test, and also the shear fixture. Iosipescu 

revealed by photo-elasticity studies that the notches sides should be at 90º 

and the notch depth must represent a quarter specimen width, in order the shear 

stress be constant on the entire shear section. The pure shear method and the 

fixture, designed by Iosipescu, were patented in Romania, U. S. A., Switzerland, 

Germany and these were the basis of ASTM (ASTM D5379, 2005) on this 

test. 

2. Experimental Tests 

For determine the mechanical and elastic characteristics of the 

chromoplastic SDP2 material (Bălan et al., 1964) tensile tests were performed 

on a INSTRON 8801 testing machine, with 10mm diameter specimens load at a 

strain rate of 0.1mm/min. The material tensile test characteristic shown in Fig. 1 

allowed to obtain the module of elasticity E=3200MPa and the yielding limit 

σc=42MPa. This characteristic, resulted from the tensile test, is a slight 

hardening Prandtl type curve, having Ep=E/90 slope. 

Fig. 1 – The chromoplastic material tensile test characteristic curve. 

Shear behavior material research was performed on standard specimens 

(ASTM D5379, 2005), shape and size are shown in Figure 2. Specimens were 

taken from a 10mm thickness material plate. Loading to pure shear specimens 

was performed using a shear fixture (Mareş et al., 2006; Leiţoiu et al., 2007) 

mounted in a WDW50 testing machine, as is shown in Fig. 3. 

A special strain-gauge rosette N2A-06-C032A-500 (Micro- 

Measurements) was installed in the shear section between the specimen notches 

for determining the shear module of the chromoplastic material. This rosette 

contains two strain gauges, with side-by-side grids, each aligned to the principal


extensional strain directions, at 45º in respect with the shear section. The shear 

strain is obtained from the relation between the extensional strains (Atanasiu & 

Jiga, 2006) 

γ=± ε − ε 

(1) 

( ) 

1 2 . 

Fig. 2 – Pure shear specimen (ASTM D5379, 2005). 

Fig. 3 – The specimen in the pure shear fixture. 

Connecting the gauges in a Wheatstone half-bridge measurements circuit, 

is obtained at the instrument the shear strain 

ε = ε − ε 

(2) 

instr 1 2 , 

where, ε1 is supplied by the first gauge, and ε2 – by the second gauge of the 

rosette. 

From Eqs. (1) and (2) is concluding


γ= ε 

(3) 

instr . 

The measurements were performed with N2322/N2314 and P3 Vishay 

recording Wheatstone electronic bridges, and the shear characteristic data were 

used for building the curves shown in Fig. 4 and Fig. 5. The slopes of that 

curves are the shear module of the chromoplastic material. The shear module 

G=1241MPa (from the first curve) and G=1072MPa (from the second curve) 

was obtained after a statistical processing of data. 

Fig. 4 – The shear characteristic obtained from data supplied by N2322/N2314. 

Fig. 5 – The shear characteristic curve obtained from data supplied by P3-Vishay 

electronic bridge. 

Table 1 

Test machine crosshead movement near the limit values of the applied load 

Crosshead 

Crosshead The limit 

Specimen 

movement near 

the maximum 

Maximum 

force, N 

movement near 

the elasticity 

force of the 

elasticity 

force, mm 

limit, mm domain, N 

1 2.9475 4459 2.2762 3891 

2 2.6363 4296 1.8500 3501 

3 2.7812 4336 2.2150 3930 

4 2.6325 4416 1.9712 3820


The shear stress state in the symmetry notch section was studied on four 

specimens. In Fig. 6 is drawn the variation of the crosshead position, which load 

the fixture and the specimen, in respect with the force, for one of the specimens. 

The Table 1 contains the most important values of the force and the movement 

of the crosshead for all specimens. In Table 2 are listed the maximum shear 

stress for the tested chromoplastic material. These values are slightly affected 

by imperfections of specimens processing, especially in the notch angle, notch 

tip radius, notch symmetry in respect with the longitudinal axis processing. 

Load [kN] 

Table 2 

Maximum shear stress in the pure shear specimens 

Specimen Sectional area 

mm 2 

Maximum load, N Maximum shear 

stress, MPa 

1 126.25 4462 35.34 

2 126.50 4297 33.97 

3 126.05 4336 34.40 

4 125.44 4416 35.20 

5 

4.5 

4 

3.5 

3 

2.5 

2 

1.5 

1 

0.5 

0 

0 2 4 6 8 10 

Deformation [mm] 

Fig. 6 – The test machine crosshead movement in respect with the load force 

applied to the specimen. 

The same appearance fracture of chromoplastic material specimens also 

can be found to the graphite polycrystalline specimens (Manhani et al., 2007).


The same appearance fracture of chromoplastic material specimens also 

can be found to the graphite polycrystalline specimens (Manhani et al., 2007). 

Fig. 7 – Broken specimens. 

Fig. 8 – The material simplified characteristic curve. 

3. Numerical Determinations 

The numerical analysis by finite element method for a chromoplastic 

material specimen, at the maximum crosshead test machine displacement, was 

carried out in linear and nonlinear regime.


Fig. 9 – The gripping elements in the FEM of the Iosipescu specimen. 

Fig 10 – Crosshead movements in vertical direction of Iosipescu specimen halves. 

The elastic domain study shows that shear stress is constant in the 

specimen minimum area section and the maximum normal stress is at 45º in 

respect with specimen axis, being parallel to the notch side. The simplified 

material characteristic σ−ε, for the nonlinear behavior regime, was used a 

Prandtl type curve, having a hardening slope Ep=35.5MPa (Fig. 8). 

The gripping and the load assigned to the model elements was properly to 

the fixture load transfer from test machine to specimen (Fig. 9). 

In Fig. 10 is shown the vertical movement of the crosshead in relation 

with the applied load, experimentally obtained, and in Fig. 11 and Fig. 12 is 

shown the maximum normal stress distribution, and, respectively, the 

equivalent normal stress distribution in Iosipescu specimen, ascertaining the 

same broken zone and constant shear stress in the shear section, as in the 

experimental study.


Fig. 11 – The maximum normal stress distribution in Iosipescu specimen. 

Fig. 12 – Equivalent normal stress distribution in Iosipescu specimen. 


1. The performed researches allow to state that the chromoplastic 

materials provide advantages for resistance structures in elastoplastic and plastic 

regime study. Chromoplastic materials models structures study, by color


change, allows an easy detection of the plastic hinges and the determination of 

loads that have produced these hinges. 

2. Based on the findings, the finite element method application can lead 

to solving local problems that occur at the parts contact or at the stress 

concentrations. The researches on chromoplastic models must be made on parts 

obtained from the same batch that was used for the specimens to determine the 

mechanic and elastic characteristics, because these characteristics vary 

depending on the compounds components. 

REFERENCES 

Iosipescu N., Determination and Experimentation of a New Procedure for Testing 

Steels Subjected to Pure Shear. Etudes et Recherches de Mécanique Appliquée, 

XIV, 2, Bucureşti (1963). 

Iosipescu, N., Recherches photoélastiques sur un procédé correct d’essais au 

cisaillement pur des matériaux. Etudes et Recherches de Mécanique Appliquée, 

XIII, 2, Bucureşti (1962). 

*** Standard Test for Shear Properties of Composite Materials by the Notched Beam 

Method. ASTM Standard D 5379/D 5379M-05, 2005. 

Bălan S., Răutu S., Petcu, V., Cromoplasticitatea (Chromoplasticity). Ed. Academiei, 

Bucureşti, 1964. 

Atanasiu C., Pastramă Ş. D., Baciu F., Vlăsceanu D., Pure Shearing Tests of a Prandtl- 

Type Material. U.P.B. Sci. Bull., Series D, 72, 3 (2010). 

Mareş M., Leiţoiu B., In Plane Shear Modulus Determination for an Unidirectional 

Carbon/Epoxy Composite Using the Iosipescu Shear Test. Bul. Inst. Polit. Iaşi, LII 

(LVI), 6 B, s. Construcţii de Maşini, 223-226 (2006). 

Leiţoiu B., Mareş M., Leiţoiu E., Consideraţii cu privire la aplicarea testului Iosipescu 

pentru studiul proprietăţilor materialelor compozite. Conferinţa Naţională de 

Mecanica Solidelor, Ed. XXXI, Chişinău, 2007, pp. 267-270. 

Atanasiu C., Jiga G., Notions fondamentales de la théorie de l’élasticité et de la 

Plasticité. Ed. Printech, 2006, pag. 12-30. 

Manhani L., Pardini L., Neto F., Assessment of Tensile Strength of Graphite by the 

Iosipescu Coupon Test. Materials Research, 10, 3, 233-239 (2007). 

RUPEREA PRIN FORFECARE A UNUI MATERIAL CROMOPLASTIC 

(Rezumat) 

Materialele cromoplastice sunt polimeri compoundaţi obţinuţi la Institutul de 

Cercetări Chimice din Bucureşti. Aceste materiale prezintă la solicitări axiale o curbă 

caracteristică de tip Prandtl. Totodată materialele au proprietatea de a-şi schimba 

culoarea în momentul apariţiei articulaţiei plastice, albindu-se la culoare, când curgerea 

este produsă de solicitarea de întindere şi închizându-se la culoare, când curgerea este 

produsă de solicitarea de compresiune. 

Autorii prezintă rezultatele cercetărilor obţinute experimental prin încercări la 

forfecare pură pe epruvete din material cromoplastic realizate conform normelor 

ASTM, care au la bază procedeul Iosipescu. Se folosesc totodată relaţiile date de teoria


elasticităţii şi se fac verificări numerice prin MEF pentru a determina tensiunea de 

rupere la forfecare, iniţierea fisurii şi forma secţiunii de rupere. 

Cercetarea dovedeşte avantajele folosirii materialelor cromoplastice în 

cercetările din domeniul elasto-plastic şi plastic.


Publicat de 



Secţia 


DETERMINATION OF SHEETS METAL ROUGHNESS 

VARIATION DEPENDING ON TENSILE STRAIN 

BY 

OVIDIU NIŢĂ ∗ , VASILE BRAHA and ANDREI MIHALACHE 


Accepted for publication: October 15, 2012 

„Gheorghe Asachi” Technical University of Iasi, 

Department of Machine Manufacturing Technology 

Abstract. The aim of this paper is to determine, using an experimental 

method, a mathematical relationship of interdependence between thin 

sheets surface roughness and deformation reached in the uniaxial 

stretching. By determining this relationship is intended to verify the 

hypothesis that the roughness can be used as a criterion for the occurrence 

of deformation limit. Also, the authors wanted to highlight how the 

rolling direction of the material influences the peak strain. 

Key words: roughness, uniaxial stretching, deformation limit, rolling 

direction etc. 


The easiest method of estimate the deformability of sheet metal blanks 

used in stamping processes is the forming limit curve method (FLD). This curve 

provides an indication of admissible deformation in a certain area of piece 

during its deformation (Banabic et al., 1992). As a result we can say that the 

size of admissible deformations is dependent on the piece rejection criteria: 

necking or fracture. However, if we are talking about the quality of sheet metal 

parts, obtained by cold forming processes, we must view the whole range of 

defects that may occur in the process of drawing (Charca et al., 2010). Thus, 

along necking or material fracture, wrinkling or thinning of material, any major 

changes in the appearance or roughness of parts surfaces may be leading to 

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

198 Ovidiu Niţă et al. 

rejection of the piece. Therefore, it is particularly important that, in addition to 

plotting FLD, to establish a relationship of interdependence between maximum 

stamped parts deformation values and the parameters imposed by rejection 

criteria (Banabic et al., 2005; Altmeyer et al., 2012). 

In the literature are presented several experimental methods used to 

determine the deformation limits (Marciniak et al., 2002; Ramir et al., 2012). 

One of the methods that deserve a special attention is the method called after 

Kobayashi (Banabic et al., 1992). Together with his colleagues, Kobayashi 

found, after analyzing the influence of deformation of work piece on surface 

roughness, that, the appearance of necking coincides with a sharp increase in 

surface roughness. This sudden increase of surface roughness is used, by 

Kobayashi, as a criterion for the occurrence of localized necking. To define this 

moment, in the deformation processes, is necessary to measure surface 

roughness at different stages of deformation and develop a chart with roughness 

variation depending on major strain ε1. 

Based on the method developed by Kobayashi and his team, we wanted 

to establish, using the tensile method, a mathematical relationship between thin 

sheets surface roughness and the deformation of analyzed material. 

2. Preparing an Conducting the Experiments 

For this approach we have conducted a series a tensile test which were 

aimed to obtain a different numbers of deformation stages for each material 

analysed. For the tensile experiments we took into account three of the most 

common used metallic sheets produced in Romania. Thus, in the experimental 

research we used rectangular specimens cut from thin sheets of the following 

materials: steel (wide strip steel B2), brass (CuZn37) and aluminium (Al 

99.5%). 

In order to determine the number of experimental tests and, of course, to 

determine the number of samples necessary to meet the proposed approach, we 

resorted to a factorial plan. Given that the ultimate goal of the entire process is 

to make a chart that will reflect the changes of specimen surface roughness 

depending by deformation, the tensile force is considered the main parameter to 

be change during the test. Thus, we used five different levels of the force 

generated by the universal test machine (Table 1). 

Nr. 

Crt. 

1 

2 

Factor Cod 

Table 1 

Factors of influence 

Factor 

level 1 2 3 4 5 6 

Lamination 

direction 

A 6 0 0 

30 0 

45 0 

60 0 

75 0 

Tensile force 

(FkN) 

B 6 F1 F2 F3 F4 F5 

90 0 

F6


Another aspect that was taken into account for evaluating the behaviour 

of material is the lamination direction of blanks. Thus, six different directions to 

the direction of rolling were taking into account for the tests: 0 0 , 30 0 , 45 0 , 60 0 , 

75 0 and 90 0 (Table 1). In conclusion, for the specific experimental tensile tests 

of the study two factors were considered the main influential factors of the 

system. 

Should be noted that, for conducting the experimental research we used 

the universal test machine WDW-50E and the strain rate was considered 

constant at 10mm/min. Also, the measurements of the surface roughness were 

undertaken using TAYLOR HOBSON-Surtronic25. Both experimental 

planning and data processing was made in Minitab v.15 (trail version). 

In Fig. 1 are captured two specimens in various stages of deformation. 

From the images we can, easily, see the direction of localized necking or how 

the rupture is propagate. 

a b 

Fig. 1 – Specimens subjected to tensile test: 

a – aluminium specimen; b – brass specimen. 

Since the first phase of the research was found that, from all of the three 

types of material studied, steel and brass specimens showed the most 

pronounced elongation compare with tensile force applied to the sample. In 

aluminium specimens case, necking and material failure occurred very early and 

elongation value stood at less than a third of the elongation recorded by brass or 

steel. 

3. Experimental Results 

3.1. Thin Sheets Deformation Capability in Relation to Lamination Direction 

The aspect that stood out even during the first phase of experimental tests 

was that the lamination direction has a significant influence on the deformation 

capability of specimens. For example, in Fig. 2 are illustrated the characteristic 

diagrams, obtained from the tensile test, for specimens orientated at 0 0 and 90 0 

degrees orientation to main rolling direction. As can be seen from Fig. 2,


aluminium presents the highest sensibility to lamination direction. Basically, the 

breaking of aluminium specimen, cut at 90 0 to the lamination direction, 

occurred in the mid-range elongation obtained for the specimen orientated after 

rolling direction. 

a b 

Fig. 2 – Characteristic diagrams obtained by tensile test: 

a – aluminium; b – brass. 

It should be noted that, the steel specimens recorded the lower sensitivity 

regarding deformation capacity to the lamination direction. This phenomenon is 

illustrated in Fig. 3, where the characteristic diagram obtained for steel blanks is 

shown. 

Fig. 3 – Characteristic diagrams obtained for steel specimens. 

3.2. Mathematical Relations between Roughness and Tensile Strain 

After the accomplish of experimental tests, it was found that the 

transition from lower to higher levels, of the considered influencing factors, 

determine a noticeable variation in surface roughness. Both influencing factors 

consider have a direct influence on the roughness, but only increasing the 

tensile force leads to a significant increase in roughness. Varying the lamination 

direction orientation also produces a fluctuation of roughness values but this 

variation occurs randomly and not after a predictable pattern. 

The nest step after the acquisition of surface roughness values was to 

draw the roughness evolution diagrams depending of specimen elongation. For


exemplification in Fig. 4 are presented the graphics (drawn from point to point) 

of roughness evolution drawn for brass specimens oriented at 0 0 and 60 0 to the 

lamination direction. Thus, we can easily see that both graphs shows two zones 

of inflection, suggesting that at a certain level of specimen’s deformation 

roughness indicate a pronounced variation. 

A b 

Fig. 4 – Roughness evolution depending on elongation: 

a – brass specimens 0 0 ; b – brass specimens 60 0 . 

In the first phase of deformation the roughness remain relatively constant 

then, when the specimen elongation reaches around 35%...40% of maximum, 

Ra parameter register a rapid growth. After the first inflection point the surface 

roughness do not show a significant increase but it increases again sharply in 

the vicinity of the failure zone. According to the results analysis it seems that 

roughness value increases by 100%, near the area where break occurs, from the 

initial value read from specimen surface. 

In all studied cases the mathematical relationship between surface 

roughness, of the specimens subjected to tensile testing, and elongation can be 

described by an exponential equation 

ax 

y = be . (1) 

In Fig. 5 a it is presented, for exemplification, the regression function, 

that describes the phenomenon studied, for steel specimens orientated at 30 0 

from rolling direction. It is noted that the coefficient of determination (R 2 ) 

represents a measure of the quality of predictions about future dependent values 

and in almost each studied cases this coefficient was above 0.8 which confirms 

the validity of the model. 

Also, in Fig. 5 b are presented, for comparative purposes, the roughness 

evolution graphs for each of the six orientations to the lamination direction, 

determined for steel specimens. By analysing the graph it can be concluded that 

the lamination direction can influence the surface roughness but not in a 

significant way.


a b 

Fig. 5 – Roughness evolution depending on elongation: 

a – steel specimens 30 0 ; b – comparative graph. 

It is noted that the graphs and the regression functions were determined 

using Microsoft Office Excel software facilities. 


1. After analysing the results obtained here, we can say that the 

lamination direction has a strong influence regarding deformation capacity of 

the specimens consider in this study. Thus, it was noted that, for sheet 

specimens cut after a direction perpendicular to the rolling direction of material, 

the elongation value was approximately 20…25% lower than the situation when 

the blanks retain the same orientation to the lamination direction. 

2. If we were to make a classification of the three analysed materials we 

can certainly say that steel specimens have the highest elongation until the 

breaking point. Instead, when we used aluminium specimens the failure point 

occurred much sooner than expected. This indicates that the aluminium used in 

the experimental research does not lead itself to getting through the stamping 

process of stamping. 

3. Establishing a mathematical relationship of interdependence between 

surface roughness (Ra) and deformation of specimens strengthens the 

hypothesis that this phenomenon, a sudden increase in roughness, can be used 

as a criterion for developing localized necking. Thus, for defining the moment 

when a work-piece deformation reaches the critical region is sufficient to 

measure roughness at different stages of deformation and to represent it 

depending on the maximum deflection. 

4. In conclusion we can say that tensile testing is one of easiest ways of 

experimental testing because it allows the use of a conventional test machine, 

present in most research laboratories. Also, the specimens, used for this type of 

testing, are easy to achieve, especially ones with the rectangular calibrated 

zones. 

Acknowledgements. This paper was realized with the support of EURODOC 

“Doctoral Scholarships for research performance at European level” project, financed 

by the European Social Fund and Romanian Government.


REFERENCES 

Altmeyer G., Abed-Meraim F., Balan T., Formability Prediction of Thin Metal Sheets, 

Using, Various, Localization Criteria. International Journal of Material Forming, 

2, 423-426 (2005); www.springerlink.com, accessed at 16.04.2010. 

Banabic, D., Dörr, I.R., Deformabilitatea tablelor metalice subţiri. Metoda curbelor 

limită de deformare, Ed. OIDICM, Bucureşti, 1992. 

Banabic, D., H., Aretz, D.S., Comsa, L., Părăianu, An Improved Analytical Description 

of Orthotropy in Metallic Sheets. International Journal of Plasticity, 21, 3, 493- 

512 (2005); www.sciencedirect.com, accessed at 20.01.2010. 

Charca G., Ramos, et al., Study of a Drawing-quality Sheet Steel. (I) Stress/Strain 

Behaviors and Lankford Coefficients by Experiments and Micromechanical 

Simulation. International Journal of Solids and Structures, 47, 17, 2285-2293 

(2010); www.sciencedirect.com, accessed at: 08.09.2010. 

Marciniak Z.., Duncan J.L., Hu S.J., Mechanics of Sheet Metal Forming. Sec. Ed., 

Butterworth-Heinemann, Oxford, 2002. 

Ramin, Hashemi, et al., Application of the Hydroforming Strain-and Stress-limit 

Diagrams to Predict Necking in Metal Bellows Forming Process. 46, 5-8, 551- 

561, source: www.springerlink.com, accessed at: 12.02.2010. 

DETERMINAREA VARIAŢIEI RUGIZITĂŢII TABLELOR SUBŢIRI 

ÎN FUNCŢIE DE DEFORMAŢIA OBŢINUTĂ LA ÎNTINDERE 

(Rezumat) 

Pornind de la metoda elaborată de Kobayashi s-a dorit stabilirea, pe cale 

experimentală, unei relaţii de interdependenţă între rugozitatea suprafeţelor şi gradul de 

deformare a materialului. Pentru realizarea acestui demers s-au efectuat o serie de 

încercări la tracţiune ce au avut drept scop obținerea unui număr diferit de grade de 

deformare pentru fiecare material supus analizei. 

În urma analizării rezultatelor obținute s-a constatat ca direcția de laminare 

influențează puternic capacitatea de deformare a epruvetelor din tablă. Astfel, s-a 

observat că pentru epruvetele din tablă ce au fost debitate după o direcție perpendiculară 

față de direcția de laminare alungirea materialului a fost cu aproximativ 20...25% mai 

mică decât în cazul epruvetelor ce pastrează aceeași orientare față de direcția de 

laminare. Totodată s-a constat faptul ca aluminiul ales pentru studiu prezintă un 

comportament atipic față de previziunile făcute anterior începerii cercetărilor 

experimentale. Practic, ruperea materialului a survenit la grade de deformare mult 

inferioare celor preconizate. 

Cel mai important aspect ce trebuie menționat este faptul ca supunerea 

epruvetelor din tablă la întindere a permis determinarea unei relații de interdependență 

între rugozitatea Ra a suprafeţei şi gradul de deformare înregistrat de material. Aceasta 

relaţie, ce îmbracă forma unei ecuaţii de tip exponenţial, întăreşte ipoteza conform


căreia acest fenomen, de creştere bruscă a rugozităţii, poate fi utilizat ca şi criteriu de 

apariție a gâtuirii localizate sau a ruperii. Astfel, pentru definirea momentului în care 

deformația semifabricatului intră în zona critică este suficient să măsurăm rugozitatea în 

diferite stadii de deformare și să o reprezentăm în funcție de deformația maximă.


Publicat de 



Secţia 


DETERMINATION OF FORMING LIMIT DIAGRAMS USING 

HYDRAULIC BULGING TEST 


Accepted for publication: October 15, 2012 

BY 

OVIDIU NIŢĂ ∗ and VASILE BRAHA 

„Gheorghe Asachi” Technical University of Iaşi, 

Department of Machine Manufacturing Technology 

Abstract. Estimation of technological possibility of obtaining a piece 

by stamping can be done most effectively by using the forming limit 

diagrams (FLD). The aim of this paper was to determine the deformation 

limit diagrams, for three metallic materials commonly used in Romanian 

industry, using hydraulic bulging test. Also, the authors wanted to 

highlight the influence of blank thickness or the deformation path on 

FLD. 

Key words: forming limit diagrams, FLD, bulging test etc. 


Sheet metal deformability expresses their ability to be plastic deformed 

and to take a given form, without defects in the specimen (Banabic, 1992). One 

of the most popular methods used for assessing the capacity of sheets 

deformation is the forming limit diagrams (FLD) method. Basically, forming 

limit curves mark off (inside a rectangular system of axes) the acceptable 

deformations from the zone where deformation turns the specimen in to scrap 

(Fig. 1a) (Marciniak, 2002; Banabic, 2005). 

Forming limit diagram is a graph obtained by plotting the evolution of the 

main deformations (ε1, ε2) in plane sheet, when fracture and necking occurs 

(Fig. 1b). As a result of using two limit deformation criterions (necking and 

fracture) the FLD have two types of forming limit curves (FLC): one for 

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

206 Ovidiu Niţă and Vasile Braha 

necking and another one for breaking (Marciniak, 2002; Banabic, 2008). 

However, the moment when the necking of material occurs is considered a 

deformation limit and this is almost unanimously accepted to be the criterion 

that defines the forming limit curve (FLC) of a specific material (Bressan, 

2003). 

a b 

Fig 1 – Forming limit diagram (FLD) 

a – FLD (www.eqsgroup.com); b – forming limit curves (FLC) (Marciniak, 2002). 

The main objective in developing the proposed paper was to evaluate 

how the deformation limit curves characterize the deformation capacity of a 

specific material. While the forming limit diagram is the currently the most 

widely used tool for assessing the deformability of sheet metal plates, the 

approach was intended to identify and study the main factors that affect, or not, 

the quality of information transmitted by FLD. 

2. Experimental Testing Conditions 

2.1. Methods and Equipment Used 

The most common method used to determine, experimentally, the 

forming limit diagram of deformation is the “network” method. On the surface 

of the blank, which will be subjected to a deformation process, the operator will 

print a rectangular, circular or radial-circular network (Fig. 2). 

a b 

Fig. 2 – Circular network with interleaved patterns: 

a – before deformation; b – after deformation.


In the deformation process, the network geometry, previously mapped, 

will change shape, providing important information on the efforts direction or 

displacement. In this case the specimen surface was printed with a circular 

network with interleaved patterns. 

To achieve experimental tests we have designed and developed a 

hydraulic inflation device (Fig. 3). The equipment is distinguished both by 

manoeuvrability and ability to easily change the active plate type used in the 

deformation process. Thus, the facility includes a total of three interchangeable 

active boards with circular, square and elliptical shape. 

a b 

Fig. 3 – Hydraulic device 

a – assembly; b – section view (1 – lower body; 2 – port-mold plate (sliding); 3 – upper 

body; 4 – elastic membrane; 5 – fixing screw; 6 – active plate (or mold plate)). 

2.2. Condition Imposed for Experimental Test 

For conducting experimental research, aimed to determine FLD, three 

types of metallic materials were considered. Thus, we choose the most 

commonly materials used in Romanian industry: steel (wide strip steel B2), 

brass (CuZn37) and aluminium (Al 99.5%). 

The factors considered to be influencing factors of the system are the 

active plate shape and the thickness of specimen used for data determination. 

Thus, for the experimental test we used specimens with thickness of 0.5 mm, 1 

mm, 1.5 mm, depending on the studied material. It should be noted that the 

design of the hydraulic stand allowed varying active plate between the rounded, 

elliptical and square plate. 

After conducting the preliminary experiments it was found that, for 

aluminium specimen with 0.5 mm thickness, material fracture occurs at very 

low pressures and deformation. Thus, we take into account the introduction, for 

further experiments, of specimens with 1.5 mm thickness.



As was stated in the introduction Forming limit diagram (FLD) is 

determined experimentally through the points of coordinates (ε1, ε2), where ε1 

and ε2 define the appropriate limit deformations of a given mode of loading of 

the specimen. The mathematical relationship that characterizes the addiction 

between ε1 and ε2 is, in fact, the equation which outlines the forming limit 

curve. 

After conducting the experimental research it was found that, in all 

studied situations, the mathematical equation that fits the best on our case is the 

second-degree polynomial relation 

2 

y = ax + bx + c . (1) 

Fig. 4 shows, for exemplification, the forming limit diagrams determined 

for 0.5mm and 1.5mm, brass and aluminium specimens, using circular active 

plate. 

e1 

0.18 

0.16 

0.14 

0.12 

0.10 

0.08 

0.06 

0.04 

0.02 

0.00 

0.00 

0.02 

FLD brass 0.5mm 

0.04 

0.06 0.08 

e2 

0.10 

0.12 

0.14 

e1 

0.25 

0.20 

0.15 

0.10 

0.05 

0.00 

0.00 

FLD aluminum 1.5mm 

0.05 

0.10 

e2 

a b 

Fig. 4 – FLD obtained with circular active plate: 

a – brass specimen 0.5mm thickness; b – aluminium specimen 1mm thickness. 

The most important observation that lies in forming limit curves analysis 

is that, the method chosen to obtain strains does not give sufficient information 

regarding the negative deformation. Basically, the charts obtain by hydraulic 

bulging test shows the limit deformation curve only for the positive principal 

strains. 

If we analyze the variation of forming limit curves positions depending 

on active plate form we can easily see that the only valid curve is obtained only 

for the circular shape of the active mold. When using elliptical or square forms 

the specimens do not reach their full potential deformation and tearing 

prematurely install. It was also found that material fracture does not occur in the 

maximum deformation stage, as would have been natural, but along the edge 

radius that limits matrix form. This can only lead to the hypothesis that, when 

using elliptical and square plates, a series of additional tangential efforts 

appears, exceeding the maximum allowable material capacity. Thus, the 

0.15 

0.20 

0.25


material failure occurs not because the material has reached its deformation 

limits. In conclusion, the use of elliptical or square active shapes can affect the 

accuracy of results. 

In the following lines is presented, through a series of comparative charts, 

the variation of FLD depending on active plate shape and the thickness of 

specimens. 

e1 

e1 

FLD variation for brass 0.5mm 

0.18 

0.16 

0.14 

0.12 

0.10 

0.08 

0.06 

0.04 

0.02 

0.00 

0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 

e2 

0.25 

0.20 

0.15 

0.10 

0.05 

0.00 

0.00 

0.05 

Variable 

Circular 

Eliptic 

Square 

e1 

0.4 

0.3 

0.2 

0.1 

0.0 

0.0 

FLD variation for brass 1mm 

0.1 

0.2 

e2 

a b 

Fig. 5 – FLD variation depending on active shape (brass) 

a – 0.5mm thickness specimens; b – 1mm thickness specimens. 

FLD variation for aluminum 1mm 

0.10 0.15 

e2 

0.20 

0.25 

Variable 

Circular 

Eliptic 

Eliptic 

e1 

0.25 

0.20 

0.15 

0.10 

0.05 

0.00 

0.00 

0.05 

0.10 0.15 

e2 

0.3 

0.20 

0.4 

FLD variation for aluminum 1.5mm 

a b 

Fig. 6 – FLD variation depending on active shape (aluminium) 

a – 1mm thickness specimens; b – 1.5mm thickness specimens. 

0.25 

Variable 

Circular 

Eliptic 

Square 

Variable 

Circular 

Eliptic 

Square 

Looking at Fig. 6 it can be easily seen that, for aluminium, the use of 

elliptical and square active plates causes very small deformations of the material 

comparative to circular plate. This leads to the conclusion that the aluminium 

chosen for the experimental tests it is not suitable for stamping processes. 

The analysis of Fig. 7, which depicts the variation of FLC for steel 

specimens, leads to the idea that the use of elliptical active plate shape provides 

distorted information on material deformability under analysis. As stated earlier 

this phenomenon is due, on the one hand, to the big difference between the 

surfaces exposed to deforming pressure and blank flange, and on the other hand, 

to additional tangential efforts located on the radius edge what limits the active 

plate form.


e1 

e1 

0.20 

0.15 

0.10 

0.05 

0.00 

FLD variation for steel 0.5mm 

0.05 

0.10 

e2 

0.15 

0.20 

Variable 

Circular 

Eliptic 

Square 

e1 

0.35 

0.30 

0.25 

0.20 

0.15 

0.10 

0.05 

0.00 

FLD variation for steel 1mm 

0.00 0.05 0.10 0.15 

e2 

0.20 0.25 0.30 

a b 

Fig. 7 – FLD variation depending on active shape (steel): 

a – 0.5mm thickness specimens; b – 1mm thickness specimens. 

Variable 

Circular 

Eliptic 

Square 

After processing the experimental data was seen that the specimen’s 

thickness have a significantly influence, regarding the FLC position on graphs, 

only when using squares active plate. In other cases FLC position are very close 

and sometimes have tangent points (Fig. 8). 

0.12 

0.10 

0.08 

0.06 

0.04 

0.02 

FLD variation - square active plate 

0.000 

0.025 0.050 

e2 

0.075 

0.100 

Variable 

0.5 mm 

1 mm 

e1 

0.12 

0.10 

0.08 

0.06 

0.04 

0.02 

0.00 

FLD variation - elliptical active plate 

0.00 

0.01 

0.02 

0.03 

0.04 

0.05 

e2 

0.06 

0.07 

0.08 

0.09 

a b 

Fig. 8 – FLD variation depending on specimen thickness (brass): 

a – square active plate shape; b – elliptical active plate shape. 

Variable 

0.5 mm 

1 mm 

It should be noted that, both experimental planning and graphs were 

obtained using Minitab V15 (trail version) software. 


1. The research presented in this paper provided an opportunity to 

determine a mathematical relationship that characterizes the interdependence 

between the two principal strains ε1 and ε2. In all studied cases it was found that 

the best equation for describing the forming deformation curve is the second 

degree polynomial equation. 

2. One of most important conclusions that deriving from this study is that, 

the use of elliptical and square active plate shapes can affect the accuracy of 

results. This can only lead to the hypothesis that, when using elliptical or square


active plates, a series of additional tangential efforts exceed the maximum 

allowable material limit. Thus, failure of material occurs before its maximum 

limit of deformation. 

3. It should be noted that, the forming limit diagram obtained using 

hydraulic bulging test provides information only for the positive values of 

deformation. 

4. The constructive solution made, for the experimental, tests may be a 

suitable facility for a research laboratory because of the simplicity of 

construction and high mobility. 

Acknowledgements. This paper was realized with the support of EURODOC 

“Doctoral Scholarships for research performance at European level” project, financed 

by the European Social Fund and Romanian Government. 

REFERENCES 

Assempour A. et al., A Methodology for Prediction of Forming Limit Stress Diagrams 

Considering the Strain Path Effect. Computational Materials Science, 45, 2, 195- 

204 (2009); www.sciencedirect.com, accessed at 25.10.2009. 

Banabic D., Dörr, I.R., Deformabilitatea tablelor metalice subţiri. Metoda curbelor 

limită de deformare. Ed. OIDICM, Bucureşti, 1992. 

Banabic, D., Aretz H., Comşa D.S., Paraianu L., An Improved Analytical Description of 

Orthotropy in Metallic Sheets. International Journal of Plasticity, 21, 3, 493-512 

(2005), source: www.sciencedirect.com, accessed at 20.01.2010. 

Banabic D., Modelarea curbelor limită de deformare, un nou instrument al fabricaţiei 

virtuale în procesele de deformare a tablelor metalice, raport cercetare. 2008; 

http://www.mcld.utcluj.ro/Raport_2008.pdf, accessed at 13.07.2010. 

Braha, V., Nagit, Gh., Negoescu, F., (2003), Tehnologia presarii la rece. Ed. CERMI, 

Iaşi. 

Bressan J.D., Influence of Thickness Size in Sheet Metal Forming. International Journal 

of Material Forming, 1, sup. 1, pp.117-119, source: www.springerlink.com, 

accessed at 26.02.2010. 

Marciniak Z., Duncan J.L., Hu S.J., Mechanics of Sheet Metal Forming (sec. Ed.), 

Butterworth-Heinemann, Oxford, 2002. 

*** http://www.eqsgroup.com/sheet-metal-forming-services/steel-formability-anal- 

ysis-training.asp; accessed at 10.09.2012. 

DETERMINAREA CURBELOR LIMITĂ DE DEFORMARE UTILIZÂND METODA 

UMFLĂRII HIDRAULICE 

(Rezumat) 

Principalul obiectiv în elaborarea lucrării propuse l-a constituit evaluarea 

modului în care curbele limită de deformare pot caracteriza capacitatea de deformare a 

unui material. De asemenea, în cadrul demersului s-a dorit identificarea şi studierea 

unor factori ce pot altera, sau nu, calitatea informaţiei transmise prin intermediul CLD, 

în condiţiile în care la ora actuală curba limită de deformare constituie cel mai utilizat


instrument de apreciere a deformabilităţii tablelor metalice. Astfel, pentru efectuarea 

cercetărilor experimentale ce au vizat determinarea CLD s-au luat în calcul trei tipuri de 

materiale metalice des utilizate în mediul industrial: alamă, aluminiul si oţelul. Factorii 

consideraţi ca factori de influentă ai sistemului au fost grosimea semifabricatului şi 

forma plăcilor active folosite în cadrul dispozitivului experimental. Astfel, pentru 

efectuarea încercărilor experimentale s-au folosit semifabricate din table cu o grosime 

de 0,5mm; 1mm si 1,5mm în funcţie de materialul studiat. Trebuie specificat că pentru 

obţinerea deformaţiei s-a optat pentru procedeul de umflare hidraulică. De asemenea, 

soluţia constructivă a standului de lucru a permis varierea plăcii active între placă cu 

forma rotundă, eliptică si patrată.


Publicat de 



Secţia 


CONSIDERATIONS REGARDING THE BEHAVIOR OF STEEL 

XC45 AT DRY FRICTION AT A TEMPERATURE OF 80°C 

BY 

MARIAN TEODOR POPESCU ∗ , NICOLAE POPA, CONSTANTIN ONESCU 

and RADU NICOLAE DOBRESCU 



University of Piteşti 

Abstract. One of the technologies for modifying the mechanical properties 

of materials is the rough chroming. Because of the advantages obtained with this 

procedure, rough chroming of steel XC45 was used also for obtaining good 

tribological properties. 

We have used a pawn-disc tribometer in order to obtain data regarding the 

behavior at dry friction. The experimental results and the conclusions are 

presented in this paper. 

Key words: Tribometer pawn-disc, wear ratio, friction surface 


A great number of modern technologies for modifying the mechanical 

properties is used in the industry, to obtain advantages from the mechanical and 

tribological properties point of view. 

In paper (Popescu et al., 2011) are widely presented the advantages of 

rough chroming procedure. 

Since paper (Popescu et al., 2011) refers to the testing of steel XC45 at a 

temperature of 400°C on a pawn-disc tribometer, this paper presents the testing 

of the same material in the same conditions, so we will present only the results. 

∗ Corresponding author: e-mail: npopa49@yahoo.com

214 Marian Teodor Popescu et al. 

2. Experimental Device 

Atmospheric tribometer, equiped with a radiative oven (TFA). The 

tribometer used for tests was made by Adamou in the tribology laboratory of 

ENI Tarbes (Fig. 1). The contact configuration is pawn-disc type. 

Fig. 1– The test device and the samples and the pawn dimensions. 

During the test, the following parameters have been enlisted, with the 

contact pickoff help: the friction coefficient as a raport between the tangential 

force and the normal force, in ASCII or EXCEL format; the vertical movement 

of the samples, represented by the material wear; the friction coefficient 

evolution over time; the contact surface temperature (see Tables 1–3). 

The pawn having the plane friction surface is made from steal „Stub” 

X22CrNi17, hardness 247 HV30. 

Table 1 

Functional Parameters 

Friction Slipping Configuration, type Plane, open 

method 

of contact 

Type of Uniaxial Conformity of In accordance with the 

movement rotation 

contact 

surface 

Ø Punch 6 mm Ø disc 37 mm 

Surface of 

contact 

113 mm 2 Average pressure


Table 2 

Demand conditions 

Charge (N) 15 

Speed (m.s -1 ) 0.1…1.5 

Temperature (°C) 80±5 

Browsed distance (m) 19240 

Number of performed rotations 1631.16 

Theoretical duration (min-s) 35 … 2100 

Table 3 

Material parameters 

Punch (friction) Disc (sample) 

Materials Stub X22CrNi17 (AISI431) 

Stainless steel martensitic 

Elasticity 

Toughness H v=247 (H v(30)) 

Preparing the surfaces 

Grinding SiC 240, 400, 1000, 2500, 4000 

papers. First 4 papers with thin 

adhesive tape, the last paper with 

foamy adhesive tape 

SiC 600, 1200, 4000 papers. For 5 

minutes, each paper. 

Speed 150 tr/min 

Charge 13 N 

Solution Ethanol Water 

Device TFA (tribometer) Grinder, Charge 13 N 

Cleaning Ethanol Ultrason Ethanol -5min 

Oven (60°C) 

State of surfaces before friction 

Before 

(nm) 

After (nm) Before (nm) After (nm) 

Arithmetic average 

rugosity (Ra) 

--- 4523 63 4323 

Average square 

rugosity (Rq) 

--- 5566 82 5250 

Skewness (Sk) --- -0.34 --- --- 

Kurtosis (Ek) --- 2.72 29.58 --- 

Due to the fact that the friction device was not dismounted after the 

preparation, we do not have images, nor values for the roughness before the test. 

The values Ra, Rq, Ssk, Sku are computed according to the analysed surface. 

The values measured after the test for the sample are taken on the track. 

3. Results 

Starting from the values enlisted by the couple contact pickoff (Nm) and 

knowing the medium radius of the friction track on the disc (r = 0.0135 m) and 

the normal force applied on the pawn, the evolution of the friction coefficient in 

time can be followed. The friction quotient was deducted from d’Amontons law 

(Fig. 2 and Table 4), with the state of the surface presented in Fig. 3.


Speed, m/s 

μ 

max 

μ 

min 

Fig. 2 

Table 4 

Friction coefficient 

μ Evolution 

mediu 

1 0.1 0.58 0.23 0.48 Increasing 

2 0.25 0.98 0.52 0.69 High variations of the friction quotient 

value 

3 0.5 0.84 0.56 0.68 Variations decrease and appears the 

stability tendency 

4 0.75 0.77 0.55 0.63 Oscillations around 0.5 

5 1 0.73 0.55 0.61 Oscillations around 0.5 



Disc 

Before the test After the test 

Pawn after the test 

Fig. 3


Morfology. The profilometrical analysis of the disc and of the friction 

device is shown in the Fig. 4 

Disc 


Pawn after the test 


Fig. 4 

Wear Rtio. The discs and pawns observation after the tests was made 

using the binocular microscop NACHET Z45P mounted on a camera XCD 

Sony and operated with the Archimed programmes from Microvision 

Instruments.


These observations have permitted the surfaces wear proceeding 

knowledge, the confirmation of the detachable particules presence or absence 

and the wear surface dimension. 

In order to compute the wear brought during the test, the topographic 

characterization of the surface was used. This characterization was made by 

interferometry with white light VEECO NT 1100 on a disc portion. 

The device permitts the surfaces visualization in 2 or 3 D and gives the 

posibility of assesing the material volumes set on one side and the other from a 

reference plain (see fig. 5 and Table 5). 

Nr. 

Wear volume (net missing 

volume) 

μm 3 

Fig, 5 

Table 5 

Wear values 

Damage volume (total 

displaced volume) 

μm 3 

Width of the surface 

μm 

1 -809652 112703056 2,40E+03 

2 -896144 109160720 2,40E+03 

3 -957504 141512048 2,40E+03 

4 -687784 110011512 2,40E+03 

5 -819720 110514248 2,40E+03 

Total -4170804 583901584 ------- 

Average -834161 116780316,8 2.0E+03 

Total wear volume µM3 -2.95E+07 

Wear ratio (µm 3 .N -1 .m -1 ) -1.22E+03 

Total damage volume µm 3 

4.13E+09 

Damage ratio (µm 3 .N -1 .m -1 ) 171437.8707



1. At the beginning of the test, with the low speed, the friction quotient 

tends to increase from a low value to a maximum value of 0,58. 

2. The speed increasing (0.25; 0.5m/s), the friction quotient increases 

with a variation between 0.9 and 0.56; at the speed of 0.75; 1; 1.25 and 1.5 m/s 

the friction quotient stabilises at an average value of 0.62, with oscillations 

around 0,65. 

3. In the same test conditions but with a temperature of 400°C, the 

average value of the friction quotient was 0.56. 

REFERENCES 

Adamou A.S., Thèse de l’Institut National Polytechnique de Toulouse, France, 2005. 

Dalverny D., Thèse de l’Universitè de Bordeaux, France, 1998 

Denape J., Introduction a la Triboligie. Frottement et usure des Materiaux. ENI Tarbes 

(2008-2009). 

Popa N., et al., Elemente de Tribologie. Ed. Univ. Piteşti, 2007. 

Popescu M.T., Alexis J., Recherches concernant le comportement des aciers revêtus au 

frottement sec et à haute temperature. Ingineria Automobilului, 14 (2010). 

Popescu M.T., Popa N., Dobrescu R.N., Denape J., Onescu C., The Rough Chromates 

Steel XC45 Wear Behaviour at Dry Friction and High Temperature. Proceedings 

of the 7th BALKANTRIB’11 International Conference on Tribology, 

Thessaloniky, Grecce, 3-5 oct 2011, pp. 81-87. 

CONSIDERAŢII PRIVIND COMPORTAREA OŢELULUI XC45 LA FRECARE 

USCATĂ LA TEMPERATURA DE 80 0 C 

(Rezumat) 

Una din tehnologiile de modificare a proprietăţilor mecanice ale materialelor este 

şi cromarea dură. Deoarece se obţin avantaje cu acest procedeu s-a utilizat cromarea 

dură a oţelului XC45 pentru a se obţine şi proprietăţi tribologice bune. 

Pentru obţinerea de date privind comportarea la frecare uscată s-a utilizat un 

tribometru pion-disc. Rezultatele experimentale şi concluziile sunt prezentate în această 

lucrare.


Publicat de 



Secţia 


NUMERICAL RESULTS ACHIEVED THROUGH THE 

DYNAMIC SHAPING OF THE ELBOW JOINT 

BY 

PETRONELA PARASCHIV ∗1 and CIPRIAN PARASCHIV 2 

1 “Gheorghe Asachi” Technical University of Iaşi, 

2 Universitaty of Medicine and Pharmacy “Gr.T. Popa”, Iaşi 

Received: May 10, 2012 

Accepted for publication: June 10,.2012 

Abstract. The general analytical shaping of the biomechanical systems can 

be realized according to the estate of the analyzed system. The stability of the 

mathematical model of a system, i.e. the determination of the equations that 

govern its dynamic processes, can be realised from more points of view, out of 

which at least two are fundamental. It has been noticed that the dynamic analysis 

can be done, likewise the cinematic analysis, adopting either the structural model 

with a degree of freedom, or the one with two degrees of freedom, taking into 

consideration only the flexure-extension move, or for an ampler analysis, two 

independent moves, flexure –extension and pronation - supination; the dynamic 

model adopted was the one with a single degree of freedom. Irrespective of the 

structural model to be adopted, in the case in which the dynamic analysis is 

supposed to determine the torso of articulation reactions, the vectorial method of 

bodies separation earlier presented can be applied; in this situation the forces that 

cannot be analytically or experimentally determined are reduced to a point given 

by a unique torso (force and moment). 

Key words: elbow articulation, dynamic analysis. 


The analytical shaping of biomechanical systems can be realised 

according to the analysed system state, therefore: 

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

222 Petronela Paraschiv and Ciprian Paraschiv 

i) static shaping, when the system is in equilibrium, stable or unstable, 

without move; 

ii) cinematic shaping, when only the system movement is regarded, 

without taking into consideration the mechanical load (force and force 

moments) that produce the movement; 

iii) dynamic shaping, when the system movement is analysed taking into 

consideration all forces and moments that determine the movement. 

The analytical shaping generally behaves the following stages: physical 

or structural shaping, mathematical shaping, pysical or structural shaping 

supposes the formulation of a “physical model”, which behaviour is supposed to 

approximate as well as possible the one of the real system. The physical model 

resembles the real system concerning the basic characteristics, but it is simpler 

and therefore more approachable to the analysis. Therefore, the component 

elements of a biomechanical system can be shaped through solid bodies, rigid or 

elastic, bows, buffers etc., while the mutual action of two bodies can be drawn 

through concentrated forces, concentrated couples, distributed loads etc. 

(Budescu & Iacob, 2005). 

In many cases, the dynamic response of the biomechanical structures can 

be represented through a model with `concentrated parameters`, composed from 

masses, bows and buffers. 

The approximations done to the formulation of the physical models refer 

to: neglecting the secondary effects; neglecting some interactions with the 

environment; replacing the characteristics `distributed` through similar 

“concentrated” parameters; linearization of cause – effect relations between 

physical variables; neglecting the time variation of some parameters. 

Along the improvement of the model and of defining more precisely the 

problem analysed, one will give up part of these approximations. 

Mathematical shaping supposes the elaboration of a `mathematical 

model` that will represent the physical model, respectively the writing of the 

estate equations (cinematic, static, dynamic) of the physical system 

(Memislogu, 2003). 

Passing from the physical model to the mathematical model is realised 

through successive stages: 

i) choosing variables that describe the estate of the system at a given 

moment; 

ii) establishing the estate equations (for instance, equilibrium, static or 

dynamic equations) for the analysed system; 

iii) establishing the compatibility equations, which express the link 

between the moves of interconnected subsystems; 

iv) writing the physical laws, meaning the constitutive relationships for 

each component element of the system (Burnstein & Wright, 1994).


2. Numerical Results 

The stability of the mathematical model of a system, meaning the 

determination of the equation that govern its dynamic processes, can be done 

from more points of view, out of which, at least two are fundamental. 

The first one supposes that the system parameters as well as the signals 

(the stimulus) that act over it represent determinist sizes, i.e. sizes which 

evolution in time can be established if it is known their anterior evolution. A 

system governed by these kinds of laws is described through a determinist 

mathematical model and it is obvious that it should be studied easier than a 

system that would listen to the laws of hazard. Generally, those kinds of 

systems are analyzed for the `normal` estate of health of the shaped 

biomechanical system. 

The second point of view supposes that an indeterminist system cannot be 

described unless by static or probabilistic sizes. The equations that express in 

this case the system behaviour represent the probabilistic mathematical model. 

Such a biomechanical system is closer to traumatic or pathological states that 

can intervene at a given moment over the analysed subject. 

Articular rehabilitation at the level of the superior or inferior member of 

the human body, using either a mobile orthosis either a specially made 

mechanism for generating and controlling the movement of a certain body 

segment, supposes the knowledge of some values of the cinematic 

characteristics from the respective articulation (especially, the angular 

amplitude for active and passive moves), as well as date of the dynamic 

characteristics regarding the force and moment of reaction from the articulation 

(Deitz, 2003). 

Normal articular amplitudes are the same, with small differences, 

irrelevant, for all individuals in the same age category, irrespective of sex or 

anthropometric data, while the reactions` torso from an articulation depends on 

the anthropometric sizes of the analysed subject and the state of mechanical 

load at which the considered body segment is subjected. This is why the static 

or dynamic analysis, after case, of a customized body biomechanical system is 

very important for the conception and realization of any technical system of 

articular rehabilitation for a passive move mainly (Elias et al., 2004; Hamilton 

& Luttgens, 2002). 

For solving the dynamic equations of the forearm, there are necessary 

more anthropometric sizes (segmentary lengths, masses, inertness moments 

etc), presented in Table 1. 

The numeric analysis was accomplished for a human subject with the 

height H = 1.71 m and weight M = 76 kg, for whom were determined the 

following anthropometric sizes, calculated according to the relations in Table 1: 

Lb = 0.31806 m; Lab = 0.24966 m; Lab-m = 0.43092 m; Lms = 0.74898 m.


The position of the weight centre: proximal: arm – 0.13867 m; forearm – 

0.10735 m; forearm and hand – 0.29388 m; superior member – 0.39695 m; 

distal: arm – 0.17938 m; forearm – 0.14230 m; forearm and hand – 0.13703 m; 

superior member – 0.35202 m; weight: arm – 2.128 kg; forearm – 1.216 kg; 

forearm and hand – 1.672 kg; superior member – 3.8 kg; mass moment, in ratio 

with: weight centre: arm – 2.50630 kg · m 2 ; forearm – 1.74199 kg · m 2 ; forearm 

and hand – 7.17301 kg · m 2 ; superior member – 7.70866 kg · m 2 ; proximal: arm – 

7.10102 kg · m 2 ; forearm– 5.24969 kg · m 2 ; forearm and hand – 22.39862 kg · m 2 ; 

superior member – 23.68117 kg · m 2 ; distal: arm – 10.05638 kg · m 2 ; forearm – 

7.94275 kg · m 2 ; forearm and hand – 10.45459 kg · m 2 ; superior member – 

20.21977 kg · m 2 . 

Table 1 

Anthropometric sizes of calculation 

Segment 

Forearm Whole 

Arm Forearm and superior 

Anthropometric 

size 

Hand member 

Lenght 0.186H 0.146H 0.252H 0.438H 

Position of 

The weight 

proximal 0.436Lb 0.430Lab 0.682Lab-m 0.530Lms 

centre distal 0.564Lb 0.570Lab 0.318Lab-m 0.470Lms 

Weight 0.028M 0.016M 0.022M 0.050M 

Weight (0.322) 

centre 

2 

(0.303) 

MLb 

2 

(0.468) 

MLab 

2 

(0.368) 

MLab-m 

2 

MLms 

proximal (0.542) 2 

MLb 

(0.526) 2 

MLab 

(0.827) 2 

MLab-m 

(0.645) 2 

Mass moment, 

in ratio with 

distal (0.645) 

MLms 

2 

(0.647) 

MLb 

2 

(0.565) 

MLab 

2 

(0.596) 

MLab-m 

2 

MLms 

Observations: H [m] – subject height; 

M [kg] – subject weight; 

Lb , Lab [m] – length of arm, respectively forearm; 

Lab-m [m] – length of arm-hand segment; 

Lms [m] – length of the superior member. 

Using the anterior numerical data, as well as the analytical relations 

(Kwak et al., 2000; Nordin, 2005), there were determined using the Matlab 

programme, the chart of the variation of the components of the reaction force 

from the elbow articulation, Rcx and Rcy and the moment of reaction from the 

elbow articulation, in flexure move for 1 s, represented in Figs. 1 and 2. 

The method of reducing the forces to a given point is useful in the 

converse dynamic analysis, given the fact that the muscular forces, 

tendons forces, ligament forces and the forces from the bony surfaces in


contact are unknown, they reducing to a unique torso of the articular 

reactions. 

Fig. 1 – Variations of the components of the reaction force. 

Fig. 2 – Variation of the moment of reaction from the elbow articulation.



As a consequence of the dynamic analysis of the elbow articulation for 

the flexure – extent move of the forearm there can be formulated the following 

conclusions: 

1) The dynamic analysis can be done likewise in the case of cinematic 

analysis, adopting either the structural model with a degree of freedom, or the 

one with two degrees of freedom, taking into consideration only the flexureextension 

move, or for an ampler analysis, two independent moves flexure – 

extension and pronation - supination; the dynamic model adopted was the one 

with a single degree of freedom. 

2) Irrespective of the structural model to be adopted, in the case in which 

the dynamic analysis is supposed to determine the torso of articulation 

reactions, the vectorial method of bodies separation earlier presented can be 

applied; in this situation the forces that cannot be analytically or experimentally 

determined are reduced to a point given by a unique torso (force and moment). 

3) In the case in which it is desired to obtain the movement equation, it is 

used one of the analytical dynamic methods. 

4) The torso of the articular reactions, depending on the state of the 

mechanical load of the body segment analyzed, represents a functional 

parameter of any technical system used in articular rehabilitation of the superior 

of inferior member, particularly the elbow articulation. 

REFERENCES 

Azar F. M., Calandruccio J. H., Arthroplasty of the Shoulder and Elbow. Canale S. T., 

Beaty J. H. (Eds.), Campbell’s Operative Orthopaedics, 11th Ed. Mosby, 

Philadelphia, 2008. 

Budescu E., Iacob I., Bazele biomecanicii în sport. Ed. Universităţii “Al. I. Cuza” Iaşi, 

2005. 

Burnstein A.H., Wright T.M., Fundamentals of Othopaedic Biomechanics. Williams & 

Wilkins Ed., New York, 1994. 

Deitz D., Optimizing Orthotic Design with FEA. Journal of Biomechanical Engineering, 

125, 6, 913 (2003). 

Elias J.J., Wilson D.R., Adamson R., Cosgarea A.J., Evaluation of a Computational 

Model Used to Predict the Patellofemoral Contact Pressure Distribution. Journal 

of Biomechanics, 37(3), 295-302 (2004). 

Fisher D. M., Borschel G. H., Curtis C. G., Clarke H.M.,.Evaluation of Elbow Flexion 

as a Pedictor of Otcome in Obstetrical Brachial Plexus Palsy. Plast. Reconstr. 

Surg., 120, 1585-1590 (2007). 

Hamilton N., Luttgens K., Kinesiology. Scientific Basis of Human Motion. McGraw – 

Hill Comp. Inc., New York, 2002.


Hatze H., The Fundamental Problem of Myoskeletal Inverse Dynamics and its 

Implications. Journal of Biomechanics, 35, 109-115 (2002). 

Kwak S.D., Blankevoort L., Ateshian G.A., A Mathematical Formulation for 3D Quasistatic 

Multibody Models of Diarthrodial Joints. Computer Methods in 

Biomechanics and Biomedical Engineering, 3, 41-64 (2000). 

Memişoglu A., Human Motion Control Using Inverse Kinematics. Master of Science 

Thesis, Bilkent University, 2003. 

Nordin F., Biomechanics of Bone. Lea & Febiger, Philadelphia, 2005. 

Pop C., Biomechanical Model of Human Body Using Bondgraphs, Inverse Dynamics, 

Simulation and Ccontrol. Masters Thesis, University of Waterloo, Canada, 2000. 

Pop C., Khajepour A., Huissoon J.P., Patla A.E., Bondgraph Modeling and Model 

Evaluation of Human Locomotion Using Experimental Data. Gait and Posture, 

35, 4 (2001). 

REZULTATE NUMERICE OBŢINUTE LA MODELAREA DINAMICĂ A 

ARTICULAŢIEI COTULUI 

(Rezumat) 

Modelarea analitică a sistemelor biomecanice trebuie elaborată în conformitate 

cu starea sistemului analizat. Stabilitatea modelului dinamic al unui sistem, adică 

determinarea ecuaţiilor dinamice, poate fi realizată din mai multe puncte de vedere 

dintre care două cel puţin sunt fundamentale. S-a remarcat că analiza dinamică poate fi 

realizată în acelaşi mod cu analiza cinematică adoptând modelul structural cu un grad de 

libertate sau cel cu două grade de libertate, luând în consideraţie numai mişcarea de 

flexie – extensie sau pentru o mai amplă analiză – două mişcări independente flexie – 

extensie şi pronaţie – supinaţie. Modelul dinamic adoptat este cel cu un grad de 

libertate. Independent de modelul structural adoptat, în cazul în care analiza dinamică 

presupune determinarea torsorului de reacţiune din articulaţii, metoda vectorială a 

separării corpurilor poate fi aplicată. În această situaţie forţele care nu pot fi determinate 

analitic sau experimental sunt reduse într-un punct dat de un torsor unic (forţă şi 

moment).


Publicat de 



Secţia 


DYNAMIC ANALYSIS OF ANKLE 

JOINT – 3D BIOMECHANICS MODEL 

BY 

VICTOR COTOROS and EMIL BUDESCU ∗ 

Received July 20, 2012 



Department of Mechanisms Theory and Robotics 

Abstract. The paper presents three-dimensional dynamic model of the 

ankle joint, considering the independent movements the dorsal-plantar flexion 

and the inversion-eversion. Internal forces which acting on the foot were reduced 

in the ankle joint center to a single force and a moment of force. Using inverse 

dynamic analysis were written equations of motion of the foot, for four different 

positions of contact between foot and ground, during walking. Equations can be 

used to determine the resultant force and moment of force of ankle joint, 

requiring values of anthropometric characteristics, force and moment of inertia 

and the contact force between the foot and the ground. 

Key words: biomechanics, ankle joint, 3D structural model, dynamics. 


The goal of the paper is to present a 3D dynamic model of the ankle joint 

for determining the resultant force and moment acting on this articulation, in 

different stages of the contact between the foot and ground during the gait. 

The ankle joint in made by two big articulations, namely: an upper joint 

formed by the lower extremities of the shank bones (tibia and fibula) and the 

upper face of the talus (talocrural joint) and a lower joint formed by talus and 

calcaneus (talocalcanean joint) (Papilian, 1982). 

∗ Corresponding author: e-mail: ebudescu2006@yahoo.com

230 Victor Cotoros and Emil Budescu 

The 3D structural model of the assembly shank-foot, used in the 

followings as required model for biomechanical analysis, is represented in Fig. 

1, where the notations signify: 

1, 2, 3 – represent the shank, the talus and respectively, the calcaneus 

together with the foot; 

A, B, C – represent the joint of the knee, the talocrural joint and 

respectively, the talocalcanean joint; 

L, M – represent the main anterior, posterior and lateral ligaments and 

respectively, the main muscular groups which perform the motions of dorsalplantar 

flexion and eversion – inversion of the foot with respect to shank. 

The same structural model but in which the shank is detailed represented 

with tibia and fibula, the foot is in contact with the ground and the muscular 

groups are not shown, can be observed in Fig. 2. In this case, the following 

notations were used: 

1 – thigh; 2 – tibia; 3 – fibula; 4 – talus; 5 – calcaneus-foot assembly; 6 – 

ground; 7, 7’, 8, 8’ – medial and external colateral ligaments; A – knee joint; B 

– talocrural joint; C – talocalcanean joint; D – contact between foot and ground 

(external link); 

( x1O1y1z 1) 

– fixed orthogonal reference system with the origin in knee 

joint, with the medio-lateral axis O1y 1, 

vertical axis Oz 1 1 and the axis Ox 1 1 

completing the right orthogonal system (antero–posterior); 

( x2O2y2z 2) 

– mobile orthogonal system solidar with the shank (elements 

2 and 3), with the origin in the talocrural joint and with the possibility to 

perform the flexion-extension motion of the shank with the angle ϕ 1 ; the 

rotation of shank is done around the medio–lateral axis O1y 1, 

the axes O1y 1and 

O2y 2 staying parallel; 

( x4O4y4z 4) 

– mobile orthogonal system solidar with the talus (element 

4), with the origin in the talocalcanean joint (C), where the axis y 4 is 

perpendicular to y 2 and around which the foot can perform the eversion– 

inversion motion; the angle of rotation of eversion–inversion was denoted with 

ϕ 4 ; 

( 5 5 5 5 

x O y z ) – mobile orthogonal system solidar with the assembly talusfoot, 

with the origin in the contact point of the foot with the ground (D), 

respectively the external link between foot and ground (element 6). 

The successive coordinate systems solidar with the elements of the 

kinematic chain, represented in Fig. 2, are useful for the kinematic analysis of 

the biomechanical system shank-ankle-foot, being possible to define the 

position matrix operators, function of the two geometric angular coordinates φ1 

and φ4.


In the case of dynamic analysis, during the gait, the force and the moment 

in the ankle joint can be determined, through inverse analysis, using the method 

of bodies isolation, by writing the motion equations either for the shank, or for 

the foot, thus being removed one of the segments of the biomechanical system 

shank-ankle-foot. For inverse dynamic analysis, the linear and angular 

accelerations of the foot must be known, these being usually experimentally 

determined. At the same time, the contact force between foot and ground must 

be known, in order to solve the system of dynamics equations. 

A 

M 

3 

L 

M 

L L 

B 

M 

2 L 

C 

M 

Fig. 1 – Structural biomechanical model shank-foot. 

2. Forces in the Ankle Joint 

Determining the reaction forces and the rotation moment in the ankle 

joint, requires to know all internal forces acting on muscles, tendons and 

ligaments and between bones surfaces in contact, the insertion points of the 

muscles, as well as the distances between the insertion points and the rotation 

center of the ankle. Taken together, all these conditions are, practically, 

impossible to be realized, because the static or dynamic biomechanical model 

would be very complex. To get rid of this drawback, the method of force 

reducing can be used. Through this method, the resultant force and moment in 

the articulation is determined. The method of force reducing, applied to the 

ankle, implies to cover the following steps (Fig. 3): 

i) there are highlighted, on the structural model, all internal and external 

forces acting on the system ankle-foot; 

ii) a force denoted with F, is considered to be the resultant of all forces 

exerted by muscles, tendons, ligaments and bones in contact; 

1


iii) a new force denoted with F * , is considered in ankle joint, as being the 

resultant of all the forces exerted by muscles, tendons, ligaments and bones in 

contact, like previous F, but translated in the articulation; 

iv) for system balance, that is, to cancel the new introduced force F * , the 

force -F * is considered, which has the same application point, the same direction 

but inverse sense with respect to force F * ; 

v) the couple of forces (F, - F * ) is representing a muscular moment Mg 

acting on the ankle joint. The muscular moment Mg takes into account both 

agonist and antagonist muscles; 

vi) the resultant force F * , with the application point at the rotation center 

of the ankle, is decomposed on the directions Ox, Oy and Oz , into R , R 

and respectively R , for the 3D structural model. 

g z 

6 

x5 

x2 

5 

z4 

x1 

4 

x2 

ϕ 

1 

x1 

z5 

ϕ 4 

8' 

D 

8 

ϕ 4 

A 

3 

x4 

O 1 

z1 z2 

ϕ 

1 

Fig. 2 – Definition of geometrical variable parameters (ϕ1 and ϕ4). 

C 

z2 

O 2 

B 

y5 

2 

7 

7' 

1 

y2 

y1 

y2 

g x g y 

Using this method of force reducing, all the forces are reduced to a 

unique resultant force and resultant moment into a certain point. This is the 

preliminary stage for writing and solving the equations of static or dynamic 

balance, for the biomechanical system shank-ankle-foot. In Fig. 3 there is 

presented, in successive images, the method of force reducing for the ankle. In 

this figure, the notations represent: F are the forces in the muscles acting 

m1,2,3..n 

on the foot; F are the forces acting in tendons; Fo is the force acting in 

t 1,2,3..n 

bone structures (tibia, fibula and talus), including the joint ligaments; RSxyz 

are 

the reaction forces of the ground on the foot along the directions Ox, Oy,Oz; GP is the weight force of the foot; F is the resultant of all forces previously


mentioned (force F is reduced in rotation center C of the ankle to a torsor 

formed by the force F * = F and a moment denoted Mg, equal to the moment of 

the couple of forces (-F * , F)); Rgx,y,z are the reaction forces in ankle joint on the 

directions Ox, Oy and Oz; O is the mass center of the foot; A, B, C, D are the 

tips of the tetrahedron representing the anatomic system ankle-foot, respectively 

heel, ankle and the tips of the fingers, for a spatial model 3D. At the foot level, 

there are taken in discussion the weight force of the foot, Gp, the inertia force 

with its components along the three axes, mpax, mpay, mpaz, the reaction force 

between foot and ground with the components RSx, RSy and RSz, the resultant 

force in the ankle decomposed along the three axes, Rgx, Rgy and Rgz, as well as 

the resultant moment, Mg. 

The successive foot positions, represented in Figs. 4....7, are 

corresponding to the initial contact of the foot with the ground, on the heel, the 

middle of total support, the begin of support on the tips of fingers and 

respectively the end of support on the tips of fingers. In these figures the 

following notations were used: 

Rgx,y,z is the resultant reaction force in the ankle, acting on directions Ox, 

Oy and Oz; Rsx,y,z is the reaction force of the ground on the foot, acting on 

directions Ox, Oy and Oz, with the application point at the heel A (for the first 

stage of support), at 70.5% from the length of the foot from the heel (for the 

second stage), on the tips of fingers B (for stages III and IV); Gp, the weight 

force of the foot, which has the origin in the mass center of the trihedron 

ABCD, as can be seen in Fig. 3; Mg, the moment acting in the ankle joint; ax,y,z 

is the acceleration of the foot at the mass center of that, on the directions Ox, Oy 

and Oz; α is the angle between the ground plane and plantar surface of the foot; 

β is the angle between the horizontal and the symmetry axis of the shank; C, is 

the ankle; A, is the heel; B and D, are the tips of fingers, considered to be the 

tips of the first and the fifth metatarsian; O, the mass center of the foot; d1,2, the 

distances, on the horizontal and vertical, between the rotation center of the ankle 

and the point of support of the foot on ground; d3,4, the distances, on the 

horizontal and vertical, between the mass center of the foot and the rotation 

center of the ankle; d5, the horizontal distance, in frontal plane, between the 

mass center, O, and the support point of the foot on ground. 

3. The Equations of Inverse Dynamic Analysis 

In order the foot or any other body segment to be in dynamic 

equilibrium, it is necessary that the sum of all mechanical loads acting on the 

body, namely: mechanical loads (forces and moments) external (active and 

passive), internal (active and passive), constraints and inertia forces, to be equal 

to zero (Budescu & Iacob, 2005; Poteraşu & Popescu, 1995). The dynamic 

balance equations for each of the four positions of support of the foot on the 

ground, represented in Figs. 4 ... 7, are the following:


where 

for the dynamic balance of the initial contact on the heel (Fig. 4), 

⎧RSx 

− Rgx = mpax, ⎪ 

⎪RSy 

− Rgy = mpay, ⎪ 

⎨ 

⎪RSz 

+ Rgz − Gp= mpaz, ⎪ 

i 

⎪⎩ + ( R ) + ( G ) + ( F ) = J , 

M M M M ε p 

g C s C p C p 

i j k 

M ( R ) = r × R = x 0 − z = i( zR ) + j( −xR− zR ) + k xR , 

C S A S A A A Sy A Sz A Sx A Sy 

R R R 

Sx Sy Sz 

i j k 

M ( G ) = r × G = x y − z = i( − z G ) + k( −x 

G ) , 

C p O p O O O O p O p 

0 −G 

0 

i j k 

p 

i i 

M ( F ) = r × F = x y − z = i( 

m a y + m a z ) + 

C O O O O p z O p y O 

ma ma ma 

p x p y p z 

+ j( −max − maz ) + k( 

max −may 

). 

p z O p x O p y O p x O 

mp, Jp – the mass and respectively the mass inertia moment of the foot, εp – 

angular acceleration of the foot, F i – inertia force of the foot, yielding the 

expressions to calculate the unknowns: 

⎧ Rgx = RSx −mpa, 

⎪ 

⎪ xRgy = RSy −mpay, 

⎪ 

⎪ Rgz =− RSz + Gp+ mpa, ⎨ 

y 

⎪Mgx 

=− zARSy + zOGp+ Jpxε px −may p z O −maz 

p y O, 

⎪ 

Mgy = RSx yA+ RSz xA+ Jpyε py + max p z O + maz p x O, 

⎪ 

⎪ 

⎪⎩ Mgy =− RSyxA+ GpxO+ Jpzε pz − max p y O + may p x O. 

For the dynamic balance of total support of the foot on the ground (Fig. 5), 

(1) 

(2)


Fig. 3 – Force reducing for biomechanical system ankle-foot.


Rgx 

RSx RSy 

RSy 

A 

M 

RSz 

RSy 

Rgz 

mpaz 

mpay 

RSz 

M 

Rgy 

RSx 

mpax 

Gp 

α 

d2 

Fig. 4 – Mechanical load for the initial contact on heel. 

RSz 

d5 

O 

RSx 

Gp 

d2 

β 

Z 

Fig. 5 – Mechanical load at the middle of total support of the foot on ground. 

β 

Z 

A 

d1 

C 

d1 

Y 

C 

d3 

Y 

X 

d3 

O 

X 

B 

α 

A O B 

d4 

d4

Rgz 

A 

Rg 

m pa y 

M 

C 

Gp 

R Sx 

R Sy 

R Sz 

Bul. Inst. Polit. Iaşi, t. LVIII (LXII), f. 4, 2012 

Rg 

m pa z 

O m pa x 

d5 

R Sz 

O 

Gp 

B 

R Sy 

R Sx 

d2 

α 

β 

A 

C 

O 

Fig. 6 – Mechanical load at the begin of the support on the fingers tips. 

M 

A 

R gz 

m pa y 

Gp 

C 

R gy 

O 

B 

m pa x 

R Sz 

R gx 

Z 

mpa z 

R Sy 

R Sx 

R Sx 

R Sy 

Y 

RS 

d 5 

Fig. 7 – Mechanical load at the end of the support on the fingers tips. 

Z 

X 

d 2 

O 

Gp 

α 

Y 

β 

X 

d3 

A 

d1 

d 3 

C 

B 

O 

d 1 

B 

d 4 

d 4 

α 

α 

237


were 

⎧RSx 


⎪RSy 

− Rgy = mpay, ⎨ 

⎪RSz 


⎩ g + C( Rs) + C( Gp) = Jp 

, 

M M M p ε 

i j k 

MC( RS ) = rE× RS= xE0 zE= i( − zER Sy ) + j( − xERSz + zERSx ) + kxER Sy , 

R R R 

Sx Sy Sz 

i j k 

MC( Gp) = rO× Gp = xO yO − zO = i( − zOGp) + k ( −x 

OG 

p ) , 

0 −G 

0 

p x p y p z 

p 

i j k 

i 

i 

MC(F ) = rO × F = xO yO − zO = i( 

mpazyO + mpayzO) + 

ma ma ma 

+ j( −max − maz ) + k( 

max −may), 


yielding the expressions to calculate the unknowns 

⎧Rgx 

= RSx −mpax, 

⎪ 

⎪Rgy 

= RSy −mpay, 

⎪ 

⎪ Rgz = Gp− RSy + mpaz, ⎪ 

⎨ 

⎪ 

Mgx = zERSy + zOGp+ Jpxε px −may p z O −maz 

p y O , 

⎪ 

⎪Mgy 

= xERSz − zERSx + Jpyεpy + max p z O + maz p x O, 

⎪ 

⎪ 

⎩Mgz 

=− RSy xE+ GpxO+ Jpzε pz − max p y O + may p x O. 

For the dynamic balance of the start of support on the tips of fingers (Fig. 6), 

the equations are 

(3) 

(4)

where 


⎧− 

RSx + Rgx = mpax, ⎪ 

⎪RSy 

− Rgy = mpay, ⎪ 

⎨ 

⎪RSz 

+ Rgz− Gp= mpaz, ⎪ 

⎪⎩ 

+ ( R ) + ( G ) = J , 

M M M ε p 

g C s C p p 

i j k 

M (R ) = x y z = i( 

y R −zR 

) + 

C S E E E E Sz E Sy 

−R 

R R 

Sx Sy Sz 

+ j( −x R − z R ) + k(x 

R + y R ), 

E Sz E Sx E Sy E Sx 

i j k 

M ( G ) = r × G = x y − z = i( − z G ) + k ( −x 

G ) , 

C p O p O O O O p 

O p 

0 −G 

0 

i j k 

p 

i i 

M ( F ) = r × F = x y − z = i( 

m a y + m a z ) + 

C O O O O p z O p y O 

ma ma ma 

p x p y p z 

+ j( −max − maz ) + k( 

max −may 

), 



⎧Rgx 

= RSx + mpax, ⎪ 

⎪ Rgy = RSy −mpay, 

⎪ 

⎪ Rgz = Gp− RSy + mpaz, ⎪ 

⎨ 

Mgx = zERSy − yERSz + zOGp+ Jpxε px −may p z O −maz 

p y O, 

⎪ 

⎪ 

⎪ Mgy = xERSz+ zERSx + Jpyε py + max p z O + maz p x O , 

⎪ 

⎪ 

⎪⎩ 

Mgz =−RSx yE− xERSy + GpxO+ Jpzε pz − max p y O + may p x O. 

For the static balance of the end of support on the tips of fingers (Fig. 7), there 

are available the equations: 

, 

(5) 

(6)


where 

⎧RSx 


RSy − Rgy = mpay, ⎪ 

⎨ 

⎪RSz 


⎪⎩ g + C( Rs) + C( Gp) = Jp 

, 

M M M ε p 

i j k 

MC( RS ) = rE× RS= xE0 zE= i( − zER Sy ) + j( − xERSz + zERSx ) + kxER Sy , 

R R R 

Sx Sy Sz 

i j k 

MC( Gp) = rO× Gp = xO yO − zO = i( − zOGp) + k ( −x 

OG 

p ) , 

0 −G 

0 

p 

i j kk 

i i 

MC( F ) = rO× F = xO yO − zO = i( 

mpazyO + mpay O) 

+ 

ma ma ma 

z 

r 

p x p y p z 

+ j( −max− maz ) + k( 

max−may ), 



⎧Rgx 

= RSx −mpax, 

⎪ 

⎪ 

Rgy = RSy −mpay, 

⎪ Rgz = Gp− RSy+ mpaz, ⎪ 

⎨ Mgx = zERSy + zOGp+ Jpxε px −may p z O −maz 

p y O , 

⎪ 

⎪ Mgy = xERSz − zERSx + Jpyε py + max p z O + maz p x O, 

⎪ 

⎪ 

⎩ 

Mgz =− RSy xE+ GpxO+ Jpzεpz − max p y O + may p x O , 

For the dynamic balance of the start of support on the tips of fingers (Fig. 6), 

the equations are given by Eqs. (5) and (6) 

(7) 

(4)


From Eqs. (2), (4) and (6), (7) there can be seen that the unknowns Rgx, 

Rgy, Rgz and Mgx, Mgy, Mgz are linearly depending on the components of the 

reaction with the ground Sx R , Sy R and Sz R the weight force of the foot Gp, the 

linear and angular accelerations of the foot but, at the same time, they are 

depending on the anthropometrical data of the foot, geometrical position of the 

contact point between the foot and ground, as well as the angles of dorsal – 

plantar and inversion – eversion flexions. 


1. From Eqs. (2), (4) and (6), (7) there can be determined the reaction 

forces and moments in ankle joint, which are necessary to evaluate the 

mechanical stresses loading tibia, fibula or astragal. Function of the value of 

these stresses, there can be analyzed various osteosynthesis techniques and can 

be made recommendations regarding the gait during bone healing 

(osteosynthesis) as simultaneous technique of articular rehabilitation. 

2. In case of tibial malleolus fracture, the mechanical load of the 

osteosynthesis assembly during the gait should not exceed the critical values for 

fracture fragments displacement. Thus, on the basis of reaction forces and 

moments in ankle, there can be analyzed the distribution of mechanical stresses 

in tibia and certain features of the gait can be recommended to patient, for a 

correct healing and articular rehabilitation. 

REFERENCES 

Budescu E., Iacob I., Bazele biomecanicii în sport. Ed. Univ. “Al. I. Cuza” Iaşi, 2005. 

Papilian V., Anatomia omului. Vol. 1, Aparatul locomotor. Ed. Didactică şi Pedagogică, 


Poteraşu V.F., Popescu D., Curs de mecanică teoretică. Vol. 1 şi 2, Ed. Universităţii 

Tehnice “Gheorghe Asachi”, Iaşi, 1995. 

ASPECTE PRIVIND BIOMECANICA ARTICULARĂ: 

ARTICULAŢIA GLEZNEI 

(Rezumat) 

Lucrarea prezintă modelul dinamic tridimensional al articulaţiei gleznei, 

considerând ca mişcări independente flexia dorsală-plantară şi inversia-eversia. Forţele 

interne care acţionează asupra piciorului au fost reduse în centrul articulaţiei la o forţă şi 

un moment unice. Folosind analiza dinamică inversă, au fost scrise ecuaţiile de mişcare 

ale piciorului, pentru patru poziţii distincte ale contactului dintre picior şi sol în timpul 

mersului. Ecuaţiile pot fi folosite pentru determinarea forţei şi momentului articular din 

gleznă, fiind necesare valori ale caracteristicilor antropometrice, forţei şi momentului de 

inerţie şi forţei de contact dintre picior şi sol.


Publicat de 



Secţia 


TIBIAL MALLEOLUS FRACTURE: VIRTUAL MODEL AND 

ANALYSIS WITH THE FINITE ELEMENT METHOD 

BY 

VICTOR COTOROS and EUGEN MERTICARU ∗ 

Received July 20, 2012 



Department of Mechanisms Theory and Robotics 

Abstract. The paper presents virtual models of the tibial malleolus fracture 

and the analysis of the state of stress for three orthopedical surgical procedures. 

Three osteosynthesis assemblies are CAD modeled and the models are subjected 

to finite element analysis (FEA). The states of stress for each osteosynthesis 

assembly are compared one to each other, resulting the most suitable surgical 

procedure. 

Key words: biomechanics, tibia, virtual model, strength, stress, finite 

element. 


Virtual simulation of the mechanical behavior of an osteosynthesis 

assembly provides to the surgeon useful information regarding the most suitable 

surgical techniques, being possible multiple analysis by the multitude of 

simulation variants, both from technical and surgical point of view. 

One of the most used methods of analysis of the state of stress and 

deformation into a virtual assembly of bodies is the finite element method 

(FEA). The idea laying to the base of the finite element method is that the 

irregular geometrical structure is partitioned in many parts with regular 

geometry, the so called finite elements, whose properties can be expressed by 

∗ Corresponding author: e-mail: emertica@mail.tuiasi.ro

244 Victor Cotoros and Eugen Merticaru 

the computer through equations that can be solved. The partitioning of the entire 

structure in many small parts is named discretization. Through this 

discretization, a difficult problem (practicly impossible to solve) is replaced 

with a simplier one. Mechanical stresses obtained with finite element method 

analysis are known as "von Mises stresses", these being equivalent mechanical 

stresses, that are icluding both normal and shear stresses. These stresses are 

giving an overall image regading the stress distribution into the analyzed 

assembly, they being able to be used for dimensioning some technical elements. 

For the proposed purpose, there was started from the virtual model of the 

osteosynthesis assembly bone-implant, designed in CAD software SolidWorks 

and ProEngineering. The model was then imported in the software ANSYS for 

finite element analysis. 

ANSYS is a software for numerical simulation with finite element. It is 

used to analyse the complex problems of mechanical structures, thermal 

proceses, fluid mechanics, magnetism, electric field etc. ANSYS has many 

graphical abilities that may be used to assess and present the anaysis results. 

There were analyzed 3 surgical orthopedical procedures: osteosynthesis 

with malleolar screw; osteosynthesis with Kirschner broaches and link with 

metalic wire; osteosynthesis with Kirschner broach, link with metalic bracing 

wire and cortical screw. 

2. Finite Element Analysis of the Osteosynthesis with Malleolar Screw 

There was started from the virtual model of the osteosynthesis assembly 

bone-implant screw, designed in the CAD software SolidWorks and 

ProEngineering. The model was then imported in the software ANSYS for 

finite element analysis. 

The geometrical complexity of the implanted functional unit leads to a 

complex static structural behavior which is impossible to be assessed by classic 

strength calculus. For this reason, there was used the finite element analysis 

with the software ANSYS. The assessment of the state of stress, strain and 

contact of the designed elements was done by structural static analysis of the 

assembly, in the same conditions of loading and support. Thus, there were 

assessed the effects of loading at the level of interface bone-screw. 

In order to assess the state of stress and strain that occurs in bone, due to 

malleolar screw, the osteosynthesis assembly was analyzed with the finite 

element method for traction loading, as can be seen in figure 1. This mechanical 

loading was considered to be the most disadvantageous for fracture 

osteosynthesis, that is, the type of loading which shows if such a type of 

assembly resists or not to maintain the contact between the fracture fragments. 

At the same time, the analytical data can be compared with the experimental 

data, because the experimental testing to traction is an usual loading for special 

testing machines.


The mechanical properties of the elements of the analyzed assembly, 

required for finite element analysis in ANSYS, are presented in Table 1. 

separati 

on 

F 

F 

Fig. 1 – Traction loading of the assembly bone-implant. 

Table 1 

Mechanical properties required for FEA 

Materials 

Elascicity modulus 

E [MPa] 

Poisson 

coefficient, ν 

Austenitic stainless steel 1.93⋅10 5 0.31 7.86⋅10 -6 

Porous bone 100 0.27 1⋅10 -6 

Periosteum 15.3⋅10 3 0.28 1.8⋅10 -6 

Compact (cortical) bone 17.5⋅10 

3 

0.28 

-6 

2⋅10 

Density 3 

ρ [kg/mm ] 

The first stage of the numerical analysis with finite element was to 

import the assembly CAD model and to declare all the materials involved in 

analysis. Thus, the uased materials were: austenitic stainless steel for the screw, 

porous bone for the internal layer of the lower epiphysis of tibia, periosteum for 

surface layer of tibia and compact (cortical) bone for lower layer of tibia 

diaphysis. The next stage was to declare the contact between elements. This 

operation was done automatically, the software ANSYS identifiyng the surfaces 

in contact and declaring links. 

Numerical analysis with finite element continued by dicretizing the 

assembly into finite elements. Discretization was done automatically, using 

tetrahedrical elements, respectively “10-Node Tetrahedral Structural Solid”. 

Initially, was tried manual discretization of the assembbly, but certain regions of 

the threaded zone could not be discretized due to assigning too small 

dimensions for discretized elements. Thus, the automatic discretization was 

chosen. Another reason for chosing automatic discretization is that the calculus 

requirements are increasing exponentially at the same time with decreasing the


dimension of the finite element. In Fig. 2 there is presented the osteosynthesis 

assembly bone-screw, after discretization. 

Fig. 2 – Discretization of the assembly bone-screw. 

The next stage was to establish the constraints and the loads acting on the 

assembly. Thus, the assembly is subjected to a traction force, and at the upper 

zone it is constrained through embedding (zero degrees of freedom). The value 

of traction force was adopted to be equal to maximum resultant reaction force, 

this value being Rg ≈ 750 N. The direction of the force was considered to be 

vertical. 

Fig. 3 – Distribution of stress in the model tibia-malleolar screw.


Graphical representation of the equivalent stress von Mises for the 

oesteosynthesis tibia-malleolar screw, obtained with the software ANSYS, is 

represented in Fig. 3. Analyzing the stresses, graphically represented with 

different colours, there can be observed that they are varying between 156 MPa 

and 941 MPa. The unit measure Pa, called Pascal, is equivalent with N/m 2 , and 

MPa represents 10 6 · Pa = 1 N/mm 2 . 

From the analysis of stresses in longitudinal section through 3D model, 

there can be seen the fact that the biggest values of stress occur at the limit 

between the screw thread and porous bone tissue, that corresponding to 

observed reality. Comparing the value of stress with the admissible one, there 

can be observed that if the fractured ankle is not immobilized, but subjected to 

normal gait, the maximum stresses could exceed the admissible values, the bone 

tissue could break and the osteosynthesis assembly could be destroyed. 

3. CAD Model of the System Bone-Kirschner Broaches–link with Metalic 

Wire and Finite Element Analysis of the Osteosynthesis Assembly 

The CAD model of the two Kirschner broaches was obtained very easy 

with the software ProEngineering, the broach being modeled as a steel rod, as 

can be seen in Fig. 4. 

Fig. 4 – Kirschner broach. 

The CAD model of the assembly tibia-Kirschner broaches–metalic wire 

(wire of thickness 3 [mm]) is presented in Fig. 5, in overall view, and in Fig. 6, 

in section view. 

The metalic wire was taken as a closed loop, because the spliced 

representation could not be possible. 

Analysis of the state of stress and strain done with the software ANSYS, 

emphasyzed the following aspects for this technique of osteosynthesis of the 

tibia malleolus fracture: 

i) when the separation plane of the fracture is traction loaded, the 

mechanical stress varies between 121 MPa and 364 MPa, as can be seen in 

Fig.7; 

ii) the stresses in the lower malleolar fragment have big values at the 

interface between broaches and bone, so that, for the given load, the bone tissue 

could be destroyed; 

iii) compared with the osteosynthesis technique with malleolar screw, the 

mechanical stresses are smaller in this case, the maximum value being much 

smaller, this meaning that this techinque is more suitable than the first one;


iv) displacement of the lower fractured fragment is smaller than for the 

first analyzed technique of osteosynthesis, so that this technique is better also, 

from this point of view, than the previous technique. 

Fig. 5 – CAD model of the assembly tibia – Kirschner broaches – metalic wire. 

Fig. 6 – Longitudinal section through osteosynthesis assembly.


Fig. 7 – Stress distribution for the osteosynthesis assembly bone-broaches-wire. 

In Fig. 7 the fracture fragments were born away, after simulation, in order 

to envision the stresses in broaches. 

Simulation performed for the virtual model tibia-two broahes-bracing 

wire was done for ideal condition of mechanical loading, namely for normal 

gate, with a healthy articulation, without taking into account that ankle 

immobilization reduces the force acting on tibia at only 30 … 40 N, in regard to 

value of 750 N considered for simulation. 

4. CAD Model of Kirschner Broaches, Link with Bracing Wire and 

Cortical Screw and Finite Element Analysis of Virtual Assembly 

In the case of the third osteosynthesis technique, there were modeled, 

additionally to the previous technique, the cortical screw and the bracing wire 

for tieing the cortical screw with the two Kirschner broaches. 

The connecting screw is intended for joints with cortical bone structures 

and is characterized by saw shape profile with small deep and filleted bottom. 

The cortical screw is standardized element with the symbol M 3.5 and has the 

following features (dimensions): central axis diamater: 2.4 mm; spire diameter: 

3.5 mm; thread length: 12 mm; total length: 20 mm; thread deep: 0.45 mm; 

thread pitch: 0.5 mm; head diameter: 6 mm; internal hole diameter: 3 mm.


The virtual model of the screw was done with software SolidWorks, as 

can be seen in Fig. 8. 

Fig. 8 – Cortical screw. 

With the software ProEngineering there was obtained the CAD model of 

the whole osteosynthesis assembly, formed by the two Kirschner broaches, 

cortical screw and bracing wire, as can be seen in Fig. 9, in overall view and in 

Fig. 10, in longitudinal section through tibia. 

The bracing wire was difficult to represent because there was not possible 

to represent the knot bonding; consequently, the bracing wire was represented 

as a continuous loop in order to obtain the virtual CAD model of the link with 

the metalic wire between the two broaches and the cortical screw. Another 

modeling difficulty was to represent the contact between the wire, bone, 

broaches and the cortical screw, due to varying geometry of the external surface 

of tibia. At the same time, as can be seen in Fig. 9, at the loop with „8” shape of 

the bracing wire, there was avoided direct contact of the wire segments in order 

to not be taken as being welded at that point (otherwise, the type of link is 

altered). The CAD model was transferred in software ANSYS were the 

assembly was discretized, as can be seen in Fig. 11. 

Then, there were established the constraints (embeding in the upper zone 

of tibia) and mechanical loads (vertical traction force equal to maximum 

resultant force in ankle joint, in dynamic regime), identical like in the two 

previous cases. For a correct simulation of the displacement of the fractured 

fragment of the malleolus with respect to tibia, there was considered a 

separation plane between the fragments, at which level the traction force acted, 

situation being similar with the previous cases. 

Finite element analysis has emphasized the following aspects of the 

osteosynthesis technique with two Kirschner broaches, cortical screw and link 

with bracing wire: 

i) the traction force was equal to 750 [N], mechanical stress varies from 

81 [MPa] up to 246 [MPa]; after simulation, a broach was removed in order to 

see the stress distribution in the hole in bone (Fig. 12);


ii) the most loaded zone is the interface bone-Kirschner broaches, being 

possible destruction of bone tissue at the considered value of traction force; 

iii) the values between which the mechanical strees varies are the smallest 

among the three analyzed osteosynthesis techniques; thus, this last surgical 

technique is recommended from biomechanical point of view; 

iv) deformations produced in the model are smaller with respect to those 

from the technique with only broaches and wire (previous case); 

v) theoretically, as resulting from simulation, there are not relative 

displacements between fracture fragments, these being very smal, by the order 

of microns, being insignificant; 

vi) theoretically, if destruction of the porous bone tissue arround the 

broaches would not occur, the patient could stress the ankle in normal gait 

condition, without relative displacement of the fracture fragments. 

Fig. 9 – General view for tibia-broaches-cortical screw-bracing wire. 

The osteosynthesis technique with Kirschner broaches, cortical screw and 

bracing wire ensures a very good stability of fracture fragments, even when the 

ankle is loaded with forces equal to the entire body weight. In reality, due to 

patient immobilisation until the bone is completely recovered, as well as due to 

the fact that mechanical load acting on the ankle is smaller, of approximately 30 

… 50 N, fracture fragments have not any relative displacement one to each 

other.


Fig. 10 – Longitudinal section through tibia with fractured malleolus. 

Fig. 11 – Discretization of the osteosynthesis assembly.


Fig. 12 – Distribution of mechanical stresses in osteosynthesis model. 

The finite elemnt analysis allows the orthopedic surgeon to use the 

simulation of various techniques, personalized for a patient, so that, before 

operation, to chose the most suitable technique. 

Like in engineering, in biomechanics there are always used the „safety 

coefficients”, with value bigger than unit, as big as the uncertainty level of the 

model is big and the risk of accident is pregnant. 


Virtual modeling of the osteosynthesis assemblies for fracture of tibia 

malleolus and the analysis with finite elements have emphasized the following 

aspects: 

a) virtual 3D modeling of the anatomic elements is a proces of imagistic 

technique still troublesome, due to complexity of represented surfaces; 

b) there is the possibility to derive some virtual models, that is, to alter 

the scale ratios or the anthropometrical dimensions of a model already done; 

c) with the help of finite element analysis there can be simulated various 

osteosynthesis techniques, trying both various techniques and various 

dimensions of the elements of a certain technique, aiming to identify the 

suitable solution for a given case; 

d) among the three types of analyzed osteosynthesis techniques, the most 

suitable for the given case, is the technique with two Kirschner broaches, 

cortical screw and bracing wire, that ensuring the smallest state of stress as well 

as the smallest strains in bone tissue;


e) the malleolar screw produces, during traction loading, a very big stress 

in bone tissue which is exceeding the breaking stress of the tissue, this situation 

leading to fractured fragments displacement and disparagement of 

osteosynthesis; this technique may be used when the ankle is immobilized or 

the load is very small; 

f) simulation and finite element analysis provide the possibility to 

determine the limit values of the forces acting on the osteosynthesis assembly, 

so that the surgeon to make the recommendation regarding the gait of the 

patient after the operation (for example, the gait with or without support). 

REFERENCES 

Bozday T., Burjan T., Bagdi C., Floris I., Vendegh Z., Varadi K., Evaluation of 

Stabilization Methods of Pelvic Ring Injuries by Finite Element Modeling. Eklem 

Hastalik Cerrahisi, 19, 127-132 (2008). 

Budescu E., Iacob I., Bazele biomecanicii în sport. Ed. Universităţii “Al. I. Cuza” Iaşi, 

2005. 

Papilian V., Anatomia omului. Vol. 1, Aparatul locomotor. Ed. Didactică şi Pedagogică, 


Poteraşu V.F., Popescu D., Curs de mecanică teoretică. Vol. 1 şi 2, Univ. Tehnică 

“Gheorghe Asachi” Iaşi, 1995. 

FRACTURA MALEOLEI TIBIALE: MODEL VIRTUAL ŞI 

ANALIZA CU METODA ELEMENTELOR FINITE 

(Rezumat) 

Lucrarea prezintă modele virtuale ale fracturii maleolei tibiale şi analiza stării 

de tensiuni pentru trei proceduri ortopedice chirurgicale. Trei ansamble de osteosinteză 

sunt modelate CAD şi modelele sunt supuse analizei cu elemente finite (FEA). Starea de 

tensiuni pentru fiecare ansamblu de osteosinteză este comparată cu fiecare din celelalte, 

rezultând cea mai potrivită procedură chirurgicală.


Publicat de 



Secţia 


LABORATORY STAND FOR TRAFFIC IMPACT STUDY OF 

AGRICULTURE AND TECHNOLOGY WORKS ON PHYSICAL 

PROPERTIES OF SOIL 

BY 

IOAN ŢENU ∗ , RADU ROŞCA, PETRU CÂRLESCU and VIRGIL VLAHIDIS 

University of Agricultural Sciences and Veterinary Medicine of Iasi, 

Department of Pedotechnica 


Accepted for publication: December 10, 2012 

Abstract. The physical degradation of soil due to the interaction with the 

active parts of the agricultural units and with the wheels of the agricultural units 

consists mainly in the deterioration of its structure and also in its compaction. In 

order to quantify these aspects, experimental studies should be developed to 

establish the values of the active parts and wheels’ working parameters leading 

to soil degradation. A laboratory test rig with soil bin was designed and built in 

order to perform studies regarding the interaction of agricultural machineries 

active parts and wheels with the ground. The studies carried out on the rig allow 

the determination of the working parameters for the active parts and the wheels 

(forces of resistance, slip etc), and their impact on the physical and mechanical 

properties of soil. 

Key words: active parts, wheel, soil channel. 


The interaction of the active parts and wheels of the agricultural 

equipments with the ground determines its physical degradation, in particular by 

compaction, but also through the deterioration of its structure (Căproiu et al., 

1982; Jităreanu et al., 2007; Ţenu et al., 2010). Experimental studies are 

necessary to quantify these aspects, in order to establish the values of the active 

∗ Corresponding author: e-mail: itenu@uaiasi.ro

256 Iona Ţenu et al. 

parts and wheels’ working parameters leading to soil degradation and 

compaction. In this context, correlations must be established between the 

wheels’ working parameters, soil compaction and the degradation level 

(Căproiu et al.,1973; Drăgan,1969; Neculăiasa,1971; Şandru et al.,1983), as 

well as the evaluation of the impact of the active parts on soil physical 

degradation indices. 

In order to solve the above-mentioned problems, a laboratory rig with soil 

bin was built in order to study the interaction of the active parts and wheels with 

the ground. 

2. Material and Method 

The test rig (Fig. 1) consists of the frame of the soil bin (1), the soil bin 

(2) and the carriage (3), on which the tillage active part (6), the tyre wheel (7) 

and the compacting roller (8) are mounted; and at the end of the soil bin frame 

an electric cable drum is attached, in order to tow the carriage by the means of 

the cable (5). 

The electric cable drum consists of an electric motor (9), a cylindrical 

gear drive and a drum (10); the latter one operates the cable (5), which is towing 

the carriage (3). The towing cable ends (5) are connected to the carriage frame 

by the means of two strain gauge load cells (4), allowing the measurement of 

the traction force needed to displace the carriage or to drive the tyre wheel. 

An electric motor (12) is mounted on the carriage and also a cylindrical 

gear for driving the tyre wheel (7). The lifting screw mechanisms (13) are used 

in order to adjust the working depth of the active part (6). The screw 

mechanism (14) allows the adjustment of the wheel’s vertical position; the 

vertical position of the compacting roller is adjusted by means of the screw 

mechanism (15). 

The carriage levelling-compacting roller (8) is used in order to achieve a 

certain level of soil compaction before the operation of the tillage active part or 

before testing the tyre wheel. 

When the carriage is towed by the means of the cable and the electric 

motor (9), the tyre wheel (7) has a free rotation movement due to its interaction 

with the soil and allows the modelling of the wheel-ground interaction process 

for a driven tractor wheel. When the carriage is not towed by the means of the 

cable, movement is achieved due to the tyre wheel (7), which is driving the 

carraige: the wheel is driven by electric motor (12) through a cylindrical gear 

and a trapezoidal belt drive. This latter situation allows modelling the wheelground 

interaction for tractor driving wheels. 

The force measuring system positioned in the front of the carriage is used 

to measure the resistance force developed by the active part at the interaction 

interface of the agricultural equipment active part-ground. The force 

measuring system located at the rear of the carriage allows the measurement of 

the driving force developed by the driving tyre wheel.


The electrical equipment of the laboratory test rig is used to supply 

electrical power to the two electric motors. Variable-speed and braking is 

achieved by means of a frequency converter. 

Fig.1 – Laboratory test rig with soil channel for the active parts-soil and wheel-soil 

interaction study: a – longitudinal vertical plane view; b – horizontal plane view; 

c – vertical cross section view. 

The traction force is measured by the force transducer. This device 

measures and displays the tension in the towing cable that pulls the carriage 

back and forth, by means of two 1000 daN load cells (strain gauge load cells), 

provided with ball joints at both ends, which operate in tandem with a 

programmable state-of-the-art weighing controller. 

The electric motor for cable carriage towing has a power of 5.5 kW and a 

speed of 1000 rpm and the electric motor used to drive the carriage powered 

wheel has 3 kW, with a speed of 1000 rpm.The carriage travelling speed when it 

is towed by the means of the 5.5 kW electric motor is 0.5… 1.55 m/s (1.8… 

5.58 km/h). The bin of the test rig was filled with a cambic chernozem soil, with 

a loam-clay texture, the size of the aggregates between 0.02… 50 mm and 

humidity of 17 to 19 %. 

3. Results and Discussions 

The laboratory test rig for the study of the interaction between the tyre 

wheel and ground and between the active part and ground was tested in two 

stages . 

In the first stage, the laboratory test rig was tested in order to check the 

constructive-functional parameters imposed by design. It has been found that 

the designed and built laboratory rig has the following constructive-functional


parameters: working depth of the tillage equipment active parts: 0...300 mm; the 

setting angle of the tillage tool: -25 0 ...+25 0 (it is possible to adjust the active 

part soil penetration angle with 25 degrees more or less than average angle); the 

carriage speed (when towed by the cable by the 5.5 kw electric motor): 0.5… 

1.55 m/s; the tyre wheel and compaction roller maximum vertical load: 500 

daN; the maximum towing force (of the carriage) at the carriage speed of 0.55 

m/s: 800 daN; the maximum towing force (of the carriage) at a carriage speed 

of 1.55 m/s: 280 daN; the minimum breaking strength of the towing cable: 

40.83 kN. It was concluded that there were no significant differences between 

the design parameters and the achieved ones. 

In the second stage, the laboratory rig was tested in order to evaluate the 

moldboard plough body-ground interaction and tyre wheelground interaction 

(Fig.2). 

3.1. Tests under Laboratory Conditions Regvarding the Active Part-soil 

Interaction 

A plough body with the working width of 200 mm, mounted on the rig 

carriage, was tested. In this case, the effects of the working depth, soil 

penetration resistance and working speed of the plough body over the towing 

force and specific power consumption were evaluated. The results of the tests 

are 

presented in Table 1. 

The results show that the increase of plough body travelling speed leads 

to a increase of the towing resistance force. At the same time, the increasing 

speed of the plough body has the effect of a pronounced increase of the specific 

power consumption. It may also be noted that an increased soil penetration 

resistance leads to a significant increase of the towing resistance force. In the 

meantime, an increased soil penetration resistance leads to a significant increase 

of the specific power consumption. 

It was also found that as the body plough working depth increases, there 

is a notable increase of the towing resistance force. 

In the meantime, the increase of the plough body working depth causes a 

slight and uneven variation of the specific power consumption: increasing the 

working depth from 100 mm to 150 mm leads to a slight decrease of the 

specific power consumption; when further increasing the working depth from 

150 mm to 200 mm, an increase of the specific power consumption was 

recorded. These variations of the specific consumption of power may be thus 

explained: for low working depths (bellow 15 cm), the slice of soil displaced 

does not form a furrow (there is no slice roll-over), so the specific power 

consumption is low; for working depths above 15 cm the conditions for the 

furrow rollover are becoming better (in this situation, the torsion and inversion 

of the furrow are achieved and the specific power consumption increases), so 

that the specific energy consumption is getting higher.


Table 1 

The working parameters of the test rig for the plough body-ground modelling 

interaction 

Plough body 

working 

depth 

mm 

100 

150 

200 

Soil penetration 

resistance 

MPa 

0.2 

0.4 

0.2 

0.4 

0.2 

0.4 

Plough body Towing Specific power 

speed resistance 

force 

consumption 

m/s 

N 

W/cm 2 

0.75 705 2.65 

1.00 720 3.60 

1.25 735 4.59 

0.75 925 3.47 

1.00 940 4.70 

1.25 960 6.00 

0.75 1055 2.64 

1.00 1070 3.57 

1.25 1080 4.50 

0.75 1380 3.45 

1.00 1400 4.67 

1.25 1420 5.92 

0.75 1450 2.72 

1.00 1470 3.67 

1.25 1485 4.64 

0.75 1859 3.47 

1.00 1890 4.72 

1.5 1930 6,03 

a b 

Fig. 2 – The test rig experimentation under laboratory conditions of the tyre wheel and 

active part interactions with the ground: a – the wheel-ground interaction study; 

b – the plough body-ground interaction study.


3.2. Tests under Laboratory Conditions Regarding the Wheel-soil Interaction 

Respecting the laws of similarity, a 5.00-12/4PR TA60 tread traction tyre 

wheel, mounted on the the laboratory rig carriage, was tested in laboratory 

conditions. In these tests, the tyre wheel performed as a driving wheel, being 

driven by the 3 kW electrical motor mounted on the carriage, so the tyre wheel 

moved the carriage along the rig soil channel. 

Driving 

wheel 

speed 

rot/min 

20 

30 

40 

Table 2 

Main working parameters of the test rig with driving wheel 

Soil Wheel Carriage Driving Driving 

penetration load speed wheel slip wheel 

resistance 

traction 

force 

MPa N m/s % 

N 

500 0.50 17 230 

0,2 750 0.51 15 360 

1000 0.53 14 580 

500 0.51 15 290 

0,4 750 0.52 15 442 

1000 0.54 13 610 

500 0.75 16 220 

0,2 750 0.76 15 350 

1000 0.79 12 570 

500 0.76 15 285 

0,4 750 0.77 14 438 

1000 0.79 12 600 

500 0.99 15 220 

0,2 750 1.01 14 340 

1000 1.02 11 560 

500 1.01 14 280 

0,4 750 1.01 13 410 

1000 1.04 10 580 

The purpose of these tests was to evaluate the main laboratory rig 

working parameters for the driving wheel condition of service. Research has 

been conducted in order to evaluate the effect of the driving wheel rotational 

speed, penetration resistance of the soil and the wheel load over the speed of the 

carriage, driving wheel slip and traction force. The results obtained in these 

experiments are presented in Table 2. 

Based on the experimental results it was established that the increase of 

the driving wheel’s rotational speed leads to a significant increase of its 

travelling speed and also to the decrease of wheel slip.


By analysing the experimental results it was found that an increased soil 

penetration resistance is leading to a slightly increased driving wheel travelling 

speed; this is explained by the fact that for more compact soil the driving 

wheel-soil adhesion increases, causing the decrease of wheel slip and therefore 

increasing the carriage travelling speed. It was also concluded that an increased 

soil resistance to penetration increases the traction force developed by the 

driving wheel, as a result of the wheel–soil adhesion increase. 

Regarding the influence of the driving wheel vertical load over the soil, it 

was established that the rise of this force leads to an increased travelling speed 

of the driving wheel, and, therefore, of the carriage, due to the increase of 

wheel- soil adhesion, that makes its slippage to diminish. 

In all cases, driving wheel slip had values that fell within the limits laid 

down by agro-technical requirements; these requirements provide that on 

compacted soil driving wheels slip should not exceed a maximum of 20 %. 


1. The laboratory test rig with soil channel, aiming to study the wheelground 

interaction, is a very important achievement, absolutely necessary in the 

experimental researches carried out in laboratory conditions regarding this 

interaction. This laboratory test rig was constructed as a result of the PNCDI II- 

52107 research grant. 

2. Experimental research carried out with plough body shows that the 

increase of plough body travelling speed leads to a increase of the towing 

resistance force and of the specific power consumption (a significant increase 

of specific power consumption). It may also be noted that an increased soil 

penetration resistance leads to a notable increase of towing resistance force and 

of the specific power consumption. 

3. It must be emphasized that as the body plough working depth 

increases, there is a increase of the towing resistance force.Regarding the 

specific power consumption, it was concluded that: increasing the working 

depth from 100 mm to 150 mm leads to a slight decrease of the specific power 

consumption and by increasing the working depth from 150 mm to 200 mm it is 

found a significant increase of the specific power consumption. These 

variations may be thus explained: the working depth increase leads to a increase 

of specific power consumption for rolling over and breaking down the 

displaced slice of soil. 

4. Based on the experimental results, it was established that the increase 

of the driving wheel rotational speed leads to a increase of its travelling speed 

and traction force and a a decrease of the wheel slip. It was also established that 

an increased soil penetration resistance resulted in increased driving wheel 

speed and traction force, while its slip decreased.


REFERENCES 

Căproiu Şt. et al., Maşini agricole de lucrat solul, semănat şi întreţinere a culturilor. 

Ed. Didactică şi Pedagogică, Bucureşti, 1982. 

Jităreanu G. şi colab., Tehnologii şi maşini pentru mecanizarea lucrărilor solului în 

vederea practicării conceptului de agricultură durabilă. Ed. "Ion Ionescu de la 

Brad", Iaşi, 2007. 

Popescu V., Cum lucrăm pământul. Ed. Tehnica Agricolă, Bucureşti, 1993. 

Roş V., Maşini agricole pentru lucrările solului. Institutul Politehnic Cluj, 1984. 

Scripnic V., Babiciu P., Maşini agricole. Ed. Ceres, Bucureşti, 1979. 

Şandru A. et al., Exploatarea utilajelor agricole. Ed. Didactică şi Pedagogică, 


Ţenu I. et al., Interacţiunea solului cu organele de lucru ale agregatelor agricole. Ed. 

"Ion Ionescu de la Brad", Iaşi, 2010. 

STAND DE LABORATOR PENTRU STUDIUL IMPACTULUI TRAFICULUI 

AGRICOL ŞI A LUCRĂRILOR TEHNOLOGICE ASUPRA PROPRIETĂŢILOR 

FIZICE ALE SOLULUI 

(Rezumat) 

Degradarea fizică a solului, determinată de interacţiunea organelor de lucru şi a 

roţilor de sprijin ale agregatelor agricole, se referă în special la deteriorarea structurii şi 

la compactarea acestuia. 

În vederea cuantificării acestor aspecte sunt necesare studii pentru a se stabili la 

ce valori ale parametrilor de funcţionare a organelor de lucru şi a roţilor de sprijin 

începe procesul de degradare fizică a solului. 

Pentru efectuarea acestor studii s-a proiectat şi realizat un stand de laborator cu 

canal de sol pentru studiul interacţiunii organelor de lucru şi a roţilor de sprijin ale 

utilajelor agricole cu solul. Cercetările ce se efectuează, prin intermediul standului, 

permit determinarea atât a parametrilor de lucru pentru organele de lucru şi roţi (forţe de 

rezistenţă, patinare etc), cât şi a impactului acestora asupra însuşirilor fizico-mecanice 

ale solului.


Publicat de 



Secţia 


THE STUDY TO OBTAINING THE TRAJECTORIES OF SOLID 

PARTICLES ON AN OSCILLATORY FLAT SURFACE 

BY 

EMILIAN MOŞNEGUŢU ∗ , VALENTIN NEDEFF, OVIDIU BONTAŞ, 

NARCIS BÂRSAN and DANA CHIŢIMUŞ 



„Vasile Alecsandri” University of Bacău 

Abstract. In this article presents a method to determine the trajectory of a 

real particle on an oscillating surface. A special attention was given to recording 

the mode of trajectory and for this purpose we used two cameras of the same 

type. Tridimensional trajectory of the particle on the oscillating surface was 

achieved through SyntEyes program. And been it was also determine the real 

speed of the particle. 

Key words: flat surface oscillatory motion, the particle trajectory. 


Vegetable products from agriculture and some of the products resulting 

from various industrial processes (grinding, granulating, briquetting etc.) is a 

heterogeneous mixture. To separate such a mixture may use different methods, 

which are chosen taking into account the nature of the system and its properties 

(phase discontinuous nature, size, temperature etc.) (Moşneguţu et al., 2007). 

Separation of a mixture of solid particles on their size difference is the 

most popular method, used to both cleaning and sorting mixtures of particles 

(Nedeff et al., 2001). 

The separations of solid particles by size represent the most old 

separation technology and we find many studies about it. But the most 

∗ Corresponding author: e-mail: emos@ub.ro

264 Emilian Moşneguţu et al. 

significant studies have been done by Brereton and Dymott in 1973, Rose and 

English in 1973, De Pretis in 1977 and Ferrara in 1988, studies which aim, in 

especially, to increase the efficiency of separation (Tsakalakis, 2001), 

(Soldinger, 1999). 

However the separation of solid particles by size is not well known, in the 

specialty literature, have appeared three trends to study this process: 

i) experimental studies (Trumic, 2011); 

ii) theoretical studies (Soldinger, 2002), (Li et al., 2003), (Stoicovici et 

al., 2008), (Szymański et al., 2003), (Xie, 2007); 

iii) studying the process through the simulation (Dong, 2009), (Zhao, 

2010), (Jianzhang & Xin, 2012). 

From specialty literature study is established that the process separation 

of solid particles after their size depends on several factors (Tsakalakis, 2001): 

i) size and shape aperture sieve; 

ii) particle size and form; moisture particles undergo the process of 

separation; intensity of vibration, respectively frequency and amplitude; 

iii) quantity of material that is fed sieve; sieve length. 

These factors also influence the behavior of solid particle on to surface 

oscillating respectively its trajectory. 

2. Materials and Equipment 

In the specialty literature there are very few experimental determinations 

which involving the use of real particles to determine the trajectory and speed of 

movement thereof on an oscillating surface. Therefore, the experimental 

determinations have performed using real particles, respectively particles of 

beans and using the worksheet - Triplot was determined form of these types of 

solid particle (Fig. 1). 

To generate oscillatory movement was used laboratory stand equipped 

with a crank rod mechanism ( 

Fig. 2). Because the laboratory stand has tyrants horizontal and vertical 

the sieve 

block executes oscillations on three directions. 

It is very difficult to follow the solid particle on the surface oscillating, so 

it was used two video cameras, Sony DCR-SR 36, which has record speed is 25 

frames/second. Cameras have been placed in two perpendicular planes (Fig. 3), 

in order to track particle movement 

on plans: 

a) XOY – camera no. 1; 

b) XOZ – camera no. 2.


During these measurements we used a blind screen inclined to the 

horizontal at an angle of 7. Inclined flat surface was subjected to oscillatory 

motion of amplitudes: 

a) on axis OX on 0.723 mm; 

1 

2 

3 

4 

b) on axis OY on 0.38 mm; 

c) on axis OZ on 1.01 mm. 

Fig. 1 – Determination of solid particle shape: a – length, b – width, c – thickness. 

5 6 7 8 

F ig. 2 – Laboratory stand components: 1 – flow control system, 2 – block the sieve, 

3 – device to control of block angle of sieve, 4 – frameworks, 

5 – counterweight, 

13 

9 

10 

11 

12


6 – eccentric the shaft, 7 – horizontal longitudinal thrusts, 8 – belt, 9 – vertical rod, 

10 – transverse horizontal thrusts, 11 – electric motor, 12 – collection cassette of 

fractions, 13 – evacuation chute. 

Values have been determined using a Vibrotest 60. 

To follow as possible the path exactly of particle solid, on the working 

surface 

was limited 

the distance on which has followed the travel of particle and 

was draw a marker line in order to be view the particle movements in sideways 

( Fig. 4). 

video camera no. 1 

X 

Fig. 3 – The video cameras location. 

Fig. 4 – Marks used on oscillating surface. 

Z 

O 

Y 


For obtaining the coordinates of the point followed, regardless of video 

camera position, was used software SynthEyes, such:


i) values obtained from the analysis of motion in the plane XOY (Fig. 5) 

will be determine: the travel time of solid particle on monitoring distance; the 

solid particle trajectory on oscillating surface; velocity of solid particle; 

ii) the second video camera (shooting in the plane XOZ) aims to 

determine the solid particle jumps on which made them at the time they travel 

on blind sieve (Fig. 6). 

Fi g. 5. Manner of determining 

the Fig. 6. Manner 

of determining the 

coordinates 

of point in the plane XOY. coordinates of point in the plane XOZ. 


In the processing of experimental data was taken into account: focus 

distance, aiming 

at the values obtained from the calculation correspond to real 

coordinates; 

the angle of the plane surface. 

So after processing the data obtained with the help of SynthEyes, it could 

determine: 

A) Solid particle trajectory on the two planes tracked by video cameras 

(Fig. 7). 

B) Tridimensional trajectory of the solid particle moving on the 

oscillating flat surface, obtained by combining the two trajectories previously 

presented (Fig. 8); 

C) Knowing the time to follow the solid particle on oscillating surface 

and the distance traveled of particle after each frame it’s possible to determine 

the speed of movement (Fig. 9);


Fig. 7 – Moving the particle solid on oscillating surface. 

11.0 

Axis OY movement (mm) 

8.8 

6.6 

4.4 

2.2 

0.0 

0 

100 

200 

6.5 

0.0 

Axis OX movement (mm) 

F ig. 8 – 3D trajectory of a particle on the oscillating surface. 

Fig. 9 – Velocity of solid particle 

on surface oscillating. 

13.0 

300 

32.5 

26.0 

19.5 

Axis OZ movement (mm)



1. The separation of solid particles after size is a complex process 

influ enced by several factors, which depend of the particles and characteristics 

of equipment used. 

2. In the experimental measurements were used real particles and from 

analysis of the Triplot chart 

shows that particle is used as lamellar particle. 

3. Because tyrants, stand used in the experimental determinations 

executed 

oscillations along three axes. 

4. By using two video cameras with purpose to tracking the solid particle 

trajectory was be achieved, with help of SynthEyes program, obtaining a threedimensional 

trajectory. 

5. Also from experimental data could determine the variation of solid 

particle 

velocity on the oscillating surface. 

REFERENCES 

Dong K.J., Yu A.B., Brake I., DEM Simulation of Particle 

Flow on a Multi-deck 

Banana Screen. Minerals Engineering, 22, 910-920 (2009). 

Li J. , Webb C., Pandiella S.S., Campbell G.M., Discrete Particle Motion on Sieves — A 

Numerical Study Using the DEM Simulation. Powder Technology, 133, 1-3, 

30, 

190-202 (July 2003). 

Moşneguţu E., Panainte M., Savin C., Măcărescu B., Nedeff V., Separarea 

amestecurilor de particule solide în curenţi de aer verticali. Ed. Alma Mater, 

Bacău, 2007. 

Nedeff V., Moşneguţu E., Băisan I., Separarea mecanică a produselor granulometrice 

şi pulverulente din industria alimentară. Ed. Tehnica-Info, Chişinău, 2001. 

Shilin Xie, Siu Wing Or, Helen Lai Wa Chan, Ping 

Kong Choy, Peter Chou Kee Liu, 

Analysis of Vibration Power Flow from a Vibrating Machinery to a Floating 

Elastic Panel. Mechanical Systems and Signal Processing, 21, 1, 389-404 (January 

2007). 

Soldinger 

M., Transport Velocity of a Crushed Rock Material Bed on a Screen. 

Minerals Engineering, 15, 1-2, 7-17 (January 2002). 

Stoic ovici D.I., Ungureanu M., Ungureanu N., Bănică, M., A Computer Model for 

Sieves’ Vibrations Analysis, Using an Algorithm Based on the “False-position” 

Method. American Journal of Applied Sciences, 

5, 12, 48-56 (2008). 

Szymański T., Wodziński P., Scre ening on a Screen with a Vibrating Sieve. 

Physicochemical 

Problems of Mineral Processing, 37, 27-36 (2003). 

Trumic M., Magdalinovic N., New Model of Screening Kinetics, Minerals Engineering, 

24, 42-49 (2011). 

Tsakalakis K., Some Basic Factors Affecting Scree Performance in Horizontal 

Vibrating Screens. The European Journal of Mineral Processing and 

Environmental Protection, I, 42-54 (2001). 

Xiao Jianzhang, Tong Xin, Particle Stratification and Penetration of a Linear Vibrating 

Screen by the Discrete Element Method. International Journal of Mining Science 

and Technology (2012).


STUDIUL OBŢINERII TRAIECTORIEI UNEI PARTICULE SOLIDE PE O 

SUPRAFAŢĂ PLANĂ OSCILANTĂ 

(Rezumat) 

O parte din produsele din agricultură sau rezultate din diferite procese industriale 

sunt supuse separării. Cea mai utilizată metodă de separare a unor astfel de produse o 

reprezintă separarea după dimensiunile particulelor care alcătuiesc amestecul eterogen. 

Cu toate că procesul de separare după dimensiuni este unul dintre cele mai răspândite, 

în literatura de specialitate sunt puţine studii care să prezinte modul de determinare a 

traiectoriei unei particule pe suprafaţa oscilantă a blocului de site. Articolul de faţă 

prezintă o metodă de determinare a acestei traiectorii, pentru o particulă reală, utilizând 

o sită oarbă care execută o mişcare oscilantă plană. Pentru determinarea traiectoriei 

particulei reale s-au utilizat două camere video care au avut drept scop filmarea 

deplasării particulei solide după planurile XOY şi XOZ. Cu ajutorul soft-ului SynthEyes 

s-au extras coordonatele punctului urmărit. Datele obţinute au fost prelucrate ţinându-se 

cont de distanţa de focalizare şi de unghiul de înclinare al suprafeţei plane. Prelucrarea 

datelor a avut drept scop determinarea traiectoriei reale a particulei şi viteza de 

deplasare a acesteia.


Publicat de 



Secţia 


LABORATORY STAND FOR THE STUDY OF CUT STRAINS 

VEGETABLE WITH KNIFE-FINGERS TYPE CUTTING 

APPARATUS 

BY 

GELU NUŢU *1 and IOAN BĂISAN 2 

1 APIA Bacău 

2 „Gheorghe Asachi” Technical University of Iaşi 



Abstract. This paper presents the construction of a stand to study plant 

stems cutting force, cutting apparatus knife-finger type. Stand using a force 

transducer and a displacement with high sensitivity, which allows to obtain 

accurate values in real working conditions. Are presented the experimental 

results for three variants of the knife and three species of cereal grains, the cut 

stems between nodes and on the node. 

Key words: laboratory stand, cutting blade, cutting force. 


Plant stems cutting operation is particularly important in the mechanical 

harvesting of plants, in terms of work quality indices and the energy 

(Neculăiasa 2002). Knife cutting machines, reciprocating finger found in 

construction combine harvester, a forage harvester and some mowers. 

Theoretical study of cutting operation with these devices is subject to a number 

of factors that can not be put into mathematical equations, and therefore the 

design and execution of their extensive use of experimental data. 

The force required depends on the shape of plant stems cutting blades, 

by way of bearing the strain during cutting, the physical and mechanical 

_______________________________ 

∗ Corresponding author: e-mail: ngsax63@yahoo.com

272 Gelu Nuţu and Ioan Băisan 

characteristics and which are variable over time, depending on the state of 

vegetation (Filipescu 1998). 

2. Material and Working Method 

To study the operation of plant stems cutting machines with cutting knife- 

finger was designed and built a laboratory stand whose construction is shown in 

Fig. 1. 

Fig. 1 – Overview of the laboratory stand. 

Fig. 2 – Schematic diagram of the operation stand. 

Fig. 2 shows the schematic diagram of the operation of the stand. In bar 4 

port blades are mounted blades 5 and is engaged with a screw mechanism 

provided at one end to the handle 11 for actuation. Between the screw 

mechanism is mounted sensor bar port power strip 3. 2 through connector that 

allows rotating screw mechanism, horizontal movement is made, the entire 

assembly for displacement transducer 7. Plant stems are cut under 6 placed 

between fixed fingers. Signals from the two transducers are taken converters 8


and 9, and sent purchasing card 10. After each experimental determination will 

get a feature cutting force-displacement type knife. 

Fig. 3 – Scheme for measuring the cutting force and blade movement. 

Fig. 3 present a schematic diagram of measuring force and displacement 

cutting blade during cutting plant stems. 

For cutting force measurement using a force transducer CTS63200KC25 

type and displacement measuring transducer type TLDT 50, whose sizes were 

taken out two type converters TA4D/2. Processed signals are transmitted to a 

data acquisition board model NI USB-6008 and from there to the central unit 

where information processing takes place using specialized software. 

Measurement resulting from a diagram that includes variation of force 

and displacement during cutting blade, such as in Fig. 4 

Table 1 

Geometric dimensions of the blades 

Type blade A B C D E F G 

Vindrover 21.6 58.4 80 12 50 76 2.27 

SEMA 17 58 75 7 50 76 2.3 

FENDT 30 50 80 15 51 76 2.61 

For experimental trials were used three variants of the cutting blades 

fitted vindrovers, combine SEMA and FENDT, whose geometric features (in 

mm) are presented in Fig. 5 and Table 1. 

The study was performed on the cut stalks of wheat, barley and oats at 

harvest, upright and inclined at 45 0 , between nodes and the node cutting 

vertically.


Fig. 4 – Diagram obtained by computer. 

Fig. 5 – The structural characteristics of the cutting blade. 

Strains used in the tests had medium size, whose size was determ ined 

with a micrometer. 


Strains, from measurements had values between 4.51 to 4.74 mm for 

wheat, between 2.94 to 3.02 mm for barley and between 

2.82 to 3.17 mm from 

oats


From experimental data presented in Table 1 can be seen that the shear 

force has the highest values in the case of strains of wheat, and here vindrover 

knives 

are cut with the smallest force, the vertical strains and strains inclined. 

As expected, the shear force values register lower in all cultures whose 

stem is inclined, in this case cutting is more advantageous energetically. 

Cutting force size changes significantly when cutting is performed on a 

node of 

the stem. The biggest differences are the cut stems of wheat is 2.5 times 

higher in type FENDT knife blades, knife type blades 2.7 times at SEMA and 6 

times higher in type vindrover knife blades. 

Table 2 

Variation of cutting force and displacement of the cutting blade plant strains 

Vindrover SEMA 

FENDT 

force (N) displacement force (N) displacement force (N) displacement 

(mm) (mm) (mm) 

vertical strain-cutting between nodes 

wheat 14,95-22,09 10,1-11,2 27,16-27,36 10,1-11,1 28,18-31,24 

17,1-18,7 

barley 9,85-12,1 9 ,1-10,5 1 4,95-19,03 16,2-17, 6 1 4,95-15,97 14 ,8-16,3 

oats 11,89-13,93 9,4-10,7 15,97-19,03 7,7-8,9 10,81-12,91 19, 2-20,2 

inclined stra in-cutting between nodes 

wheat 1 3,93-22,09 10,1-11,2 25,16-29,2 11,8-13,2 27,45-32,21 15,1-16,7 

barley 9,25-11,89 12,8-13,8 11,89-16,98 1 1,4-13,1 14,73-16,57 15,4-16,9 

oats 7,81-11,89 14,9-16,5 13,25-17,02 8,9-10,1 7,77-9,79 17,6-19,2 

vertical strain-cutting p er node 

wheat 117,8-132,1 8,2-10,1 62,84-73,1 12,2-13,3 62,84-76,9 19,2-21,1 

barley 15,79-20,05 8,9-10,1 32,28-44,48 10,0-11,3 2 2,09-27,16 14,8-16,3 

oats 26,14-34,3 6,7-7,8 57,74-68,5 7,9-9,3 25, 12-26,14 19,1-20,2 

The 

cutting force cutting strains of barley is a smaller increase, from 1.66 

times at 

vindrover and 2.31 ti mes at SEMA, while the oats betwee n 2.02 

times 

at FENDT and 3.57 times at SEMA. 


1. The stand realized allows the 

determination with sufficient accuracy 

cutting force and displacement stems vegetable knife 

while cutting the node or 

between nodes. 

2. The construction is equipped with load and displacement with high 

sensitivity and presentation software makes chart as measured values. 

3. The stand 

can be used to measure cutting force for any crop at all 

stages of vegetation and any constructive variant of type knife-finger cutting 

apparatus. 

REFERENCES 

Filipescu I. , Contribuţii 

la tăierea mecanizată a viţei de vie în uscat. Teză de 

doctorat, Universitatea Tehnică ,,Gheorghe Asachi” Iaşi, 1998.


Neculăiasa V., Operaţii şi procese de lucru ale maşinilor 

agricole de recoltat. Ed. 

,,Gheorghe Asachi”, Iaşi, 2002. 

STAND DE LABORATOR PENTRU STUDIUL 

TĂIERII TULPINILOR VEGETALE 

CU APARATE DE TĂIERE 

DE TIP CUŢIT-DEGET 

(Rezumat) 

Lucrarea faţă de prezintă construcţia şi funcţionarea unui stand de laborator 

destinat studiului tăierii tulpinilor vegetale şi care permite măsurarea forţei 

de tăiere şi a 

deplasării cuţitului în timpul tăierii, parametri ce se pot determina pentru orice fel de 

cultură agricolă şi pentru diferite tipuri constructive şi unghiuri de ascuţire a lamelor. 

Standul utilizează un traductor de forţă şi unul de deplasare cu sensibilitate 

ridicată, fapt ce permite obţinerea unor valori precise, în condiţii reale de lucru. 

Încercările experimentale au fost efectuate pentru trei culturi agricole, grâu, orz 

şi ovăz, toate la momentul recoltării efective din câmp, pe stand fiind folosite lamele 

tăietoare ce echipează trei game de maşini de recoltat: vindrovere pentru plante furajere 

şi combinele SEMA, respectiv FENDT. S-a urmărit determinarea forţei de tăiere pentru 

plantele aflate în poziţie verticală (în cea mai mare parte ) şi înclinate la 45 0 (în cazul 

culturilor culcate de vânt), tăierea fiind executată atât pe intervalul dintre nodurile 

tulpinii, cât şi pe nod.


Publicat de 



Secţia 


RESEARCHES CONCERNING THE CUTTING OF STRAINS 

PLANT WITH CUTTING DEVICES FROM COMBINE 

HARVESTERS 

Received: 

Accepted for publication: 

BY 

GELU NUŢU *1 , IOAN BĂISAN 2 

1 APIA Bacău 

2 Technical University „G. Asachi” from Iaşi 

Abstract. This paper presents experimental results obtained from strains 

cutting of crops, with the help of cutting apparatus, knife-finger type, out of 

grain harvesters in use in the Romanian agriculture. Cutting devices have been 

studied from combine DROPIA, SEMA GLORIA, CLASS, JOHN DEERE, 

FENDT and NEW HOLLAND, on a stand that allowed the cutting force 

measurement of strains on the interval between nodes and on the node. 

As study material, we used strains of main agricultural crops which are 

harvested by these machines, during special maturity stage at harvest. 

Key words: cutting force, strain vegetable, harvester. 


Machine harvesters are generally designed for harvesting cereal grains, 

which have adapted working equipment for harvesting and other crops like 

corn, sunflower, bean peas, soybean, etc.. 

Cutting machines used in construction machinery plant for harvesters are, 

in most, like knife finger, the finger pads mounted cutting board. They work 

properly in cultures with densities ranging from 200...800 plants/m 2 , between 

1.5 to 4.5 mm diameter stems and cut resistance between 0.2 to 0.8 N/cm 2 

(Bochat, 2009). 

______________________________ 

∗ Corresponding author: e-mail: ngsax63@yahoo.com


To calculate the power necessary for operating the cutting machine knifefinger, 

experimental data is being used which recommend values between 

0.58...1.18 kW/m working width linear harvesting grain, and between 

1.10...1.84 kW/m to forage harvesting. The drive power needed for the cutting 

machine increases to values of 1.77 to 2.2 kW/m when fitted with two knives 

(Neculăiasa, 2002). 

If we take into account the above, knowledge of cutting force of strains, 

depending on the design of the cutting device, is required for optimal choice in 

terms of energy alternatives. 

2. Material and Working Method 

For experimental studies there were used active elements of the cutting 

apparatus of harvesters SEMA, DROPIA 1110, GLORIA 1120 and GLORIA 

1422 from romanian production, and the harvesters of CLASS, NEW 

HOLLAND, JOHN DEERE and FENDT, from import. The structural 

characteristics of the cutting blades (A-H in mm, and the angles i and α in 

degrees) are shown in Fig. 1 and Fig. 2, respectively Table 1. Sharpening angle 

of the cutting blade is 20O for romanian combine and 19O for the import and 

the cutting was done with upright stems as found in most of their field. 

With a force transducer, the force required to cut stems of 11 cultures was 

measured, the most widely used in agricultural practice in the country (wheat, 

barley, oats, triticale, corn, sunflower, bean, peas, soybeans, sorghum and 

rapeseed) both when cutting between strain nodes and on the node. Strains 

humidity was specific during the optimal timing of harvesting crops studied. 

Fig. 1 – Construction of the cutting blade from combines SEMA, DROPIA, GLORIA, 

NEW HOLLAND and CLASS.


Experimental attempts aimed cutting force variation for the four types of 

cutting devices, in order to find advice on choosing the type of cutting device 

according to culture, such as cutting energy consumption to be minimized. 

Fig. 2 – Construction of the cutting blade from combines JOHN DEERE and FENDT. 

SEMA 

DROPIA 

GLORIA 

NEW 

HOLLAND 

Table 1 

Geometric characteristics of the blades from combine 

A B C D E F G H i α 

7 58 17 50 75 2,3 5 1.45 20 55 

13 53 30.4 51 83.4 2.82 7 1.31 19 27.3 

CLASS 16 51 33 51 84 2.44 9 1.83 19 31 

JOHN 

DEERE 

13 50 31 51 81 2.67 5.5 1.4 19 46 

FENDT 15 50 30 51 80 2.61 7.5 1.75 19 38 

Experimental attempts aimed cutting force variation for the four types of 

cutting devices, in order to find advice on choosing the type of cutting device 

according to culture, such as cutting energy consumption is minimized. 


Experimental data obtained from measurements are presented in Table 2, 

for cutting stems between nodes and Table 3 for cutting stems per node.


Table 2 

Cutting force for cutting plant stems between nodes (in N) 

DROPIA CLASS FENDT JOHN NEW 

SEMA 

GLORIA 

DEERE HOLLAND 

Wheat 21.39-29.55 21.41-24.43 23.41-26.49 16.32-19.35 20.37-25.47 

Soybean 64.18-83.54 38.71-67.24 46.05-58.09 45.85-59.11 41.77-51.97 

Rapeseed 103-114 93.7-107 80.5-108 90.68-104 86.60-95.78 

Sunflowers 178-191 153-170 142-154 126-138 153-168 

Sorghum 137-167 106-136 133-160 149-166 132-173 

Barley 9.18-11.22 5.04-8.16 6.8-10.20 7.14-8.89 5.10-9.18 

Oats 6.16-10.22 3.06-6.08 4.04-8.12 8.20-11.24 5.16-8.55 

Triticale 77-102 72.3-92.7 80.5-93.74 78.46-89.66 81.50-98.72 

Peas 6.12-11.41 4.62-8.06 5.08-11.10 7.14-9.16 5.87-10.2 

Beans 20.37-28.53 18.29-25.47 17.34-23.53 19.35-26.31 12.27-22.29 

Corn 484-653 552-577 426-482 396-471 570-633 

Table 3 

Cutting force for cutting plant stems on the nodes (in N) 

DROPIA CLASS FENDT JOHN NEW 

SEMA 

GLORIA 

DEERE HOLLAND 

Wheat 65.20 68.26 62.15 61.13 61.45 

Soybean 130.4 114.1 97.82 59.11 83.54 

Rapeseed 131.4 125.3 138.5 137.5 129.3 

Sunflowers 200.9 206.1 183.0 190.5 208.5 

Sorghum 243.4 210.3 232.9 221.0 223.1 

Barley 26.67 18.25 22.41 21.43 25.47 

Oats 29.23 25.27 21.39 28.73 22.14 

Triticale 131.81 112.74 95.78 109.22 101.88 

Peas 23.43 14.20 17.26 18.30 21.39 

Beans 52.99 55.10 45.85 49.93 62.30 

Corn 752 814 620 674 727 

From the analysis of experimental data can be seen that there is no a 

constructive type cutting blades or cutting device that records the minimum 

values of cutting force at all the cultures. So, at wheat, sunflower, triticale and 

corn the lowest values were recorded with JOHN DEERE blades type, at 

soybean, rapeseed and bean with NEW HOLLAND blades type , at sorghum, 

barley, oats, triticale and peas with CLASS blades type, while only on bean 

with FENDT blades type. 

It can be seen that the cutting blades model from romanian combines are 

not recorded cutting force values closer to the minimum in any one culture, 

which should lead for more attention of finding the corresponding solutions.


There have been the measurements of cutting force strains on the node, 

the values in Table 3 were obtained for the following node diameters: 4.74 mm 

for wheat, 3.02 mm at barley, 3.17 mm at oats, 6.93 mm at triticale, 4.92 mm at 

soybean, 3.45 mm at pea, 4.72 mm at beans, 11.56 mm at rapeseed, 22.68 mm 

for sunflower, 8.13 mm at sorghum and 31.02 mm for maize. So, for wheat the 

lowest values of cutting force is recorded at the JOHN DEERE and NEW 

HOLLAND blades type, at soybean, rapeseed, sorghum, oats and peas CLASS 

model, at sunflower oats, triticale and corn bean FENDT model, followed by 

JOHN DEERE model at soy, oats and corn. 


1. Having regard to specific geometric characteristics of each type of 

cutting blades, in terms of width blade, sharpening angle and opening angle of 

the blade grained, one can say that there is no constructive alternative to record 

the minimum cutting force values for all the cultures studied. 

2. On the basis of experimental data can be choosing a type of harvester 

for each crop, or a group of crops for which the values of the cutting force are 

minimum. 

3. As regards of the process of cutting the stems, cutting blades from the 

romanian combines registered values not closer to the minimum in any of the 

cultures studied, and this must be investigated and find a solution to reduce 

cutting force levels comparable with the imported machinery. 

REFERENCES 

Bochat A., Identification of Quasi-static Cutting Force of Triticale-straws for 

Designing Use of Scissor and Finger Cutting Sets. Journal of Polish CIMAC, 4-3, 

17-22 (2009). 

Neculaiasa V., Operaţii şi procese de lucru ale maşinilor agricole de recoltat. Ed. 

,,Gheorghe Asachi”, Iaşi, 2002. 

CERCETĂRI PRIVIND TĂIEREA TULPINILOR VEGETALE CU APARATE DE 

TĂIERE DE LA COMBINELE DE RECOLTAT CEREALE 

(Rezumat) 

Se prezintă rezultatele experimentale obţinute la tăierea tulpinilor la 11 culturi 

agricole, cele mai importante în agricultura românească, folosind aparatelor de tăiere de 

tip cuţit deget din dotarea unor combine de recoltat cereale aflate în exploatare. Au fost 

studiate aparatele de tăiere de la combinele româneşti DROPIA, SEMA şi GLORIA, 

precum şi de la combinele CLASS, FENDT, JOHN DEERE şi NEW HOLLAND, pe un 

stand care a permis măsurarea forţei de tăiere a tulpinilor pe intervalul dintre noduri şi 

pe nod. 

Cu ajutorul unui traductor de forţă s-a măsurat forţa necesară tăierii tulpinilor la 

grâu, orz, ovăz, triticale, porumb, floarea-soarelui, mazăre, soia, fasole, sorg şi rapiţă, 

atât la tăierea între nodurile tulpinilor, cât şi la tăierea pe nod.


Rezultatele experimentale au arătat că nu există în acest moment un aparat de 

tăiere care să realizeze forţe de tăiere minime pentru toate culturile luate în studiu. Pe 

baza datelor experimentale se poate face alegerea unui tip de combină de recoltat pentru 

fiecare cultură agricolă, sau pentru un grup de culturi pentru care se înregistrează valori 

minime ale acesteia.


Publicat de 



Secţia 


STUDY THE BEHAVIOR OF FOOD MATERIALS WITH SOFT 

TEXTURED SUBJECT TO SHRED BY CUTTING THROUGH 

THE TEXTURE ANALYSIS METHOD 

BY 

MIRELA PANAINTE ∗ , VALENTIN NEDEFF, CIPRIAN OLARU, 

CLAUDIA TOMOZEI and OANA IRIMIA 

1 ”Vasile Alecsandri” University of Bacău, 

Faculty of Engineering 

Department of Environmental and Mechanical Engineering 


Accepted for publication: December 10, 2012 

Abstract. Studies and experimental research in the shredding by cutting 

field at food materials with soft texture are particularly important; their results 

can provide more clarification in terms of the shred analysis at food materials 

with soft texture regarding aspect on the cut results, quantitative and qualitative 

analysis of material losses, to determine cutting forces. 

The paper presents results of a study that followed an analysis in the 

laboratory conditions by shredding by cutting process: how to behave material to 

shred, the forces necessary to shredding, and the shredding aspects. 

Key words. shredding process, soft texture, process, force. 


Many manufacturing processes require shredding of food in advance of 

food materials with soft texture. In general notion of particle is used to illustrate 

a volume of material. These particles can be the result of a natural process 

(fruits and vegetables) or obtained as a result of technological processes 

(cheese, crap food, butter etc.). 

∗ Corresponding author: e-mail: mirelap@ub.ro

284 Mirela Panaite et al. 

Shredding of food materials with soft texture is the process to creation of 

new surfaces. 

Shredding of food products must ensure the properties of materials, in the 

shredding process must not appear excessive heating or possibly tearing due to 

improper choice of the working body. To improve performance can be use different 

types of automation systems to ensure proper shredding and increased 

efficiency of shredding (Brown, 2005), (Brown et al., 2000), (Fellows, 1999), 

(Reilly, 2004). 

At shredding by cutting of soft textured materials, their condition changes 

in multiple aspects. They are required during shredding both mechanical and 

thermal. As they develop conditions of deformation and structural changes 

especially at newly created surfaces. 

The research conducted so far shows that are not known in detail 

shredding forces on products subject to shred and cut body influence the process 

of shredding and shred productivity (Ciumac, 1995), (Raven, 1998), 

(Shmulevich, 2003). 

Few studies in the literature have investigated in detail the behavior of 

food materials with soft texture during the shredding by cutting operation. In 

general it was considered food materials bending test, which tests were used to 

study the behavior of soft textured cheese at low temperatures (Charalambides 

et al., 1995), (Charalambides et al., 2001), Imbeni et al., 2001), (Walstra &Van 

Vliet 1982). 

For optimal development of the shredding system, are needed data on 

cutting forces for different types of food and how they may vary depending on 

various factors such as cutting force, cutting speed, type of device cutting. 

2. Research Method and Experimental Results 

Studies and experimental research in shred by cutting domain of food 

materials with soft texture are particularly important; their results can provide 

more clarification in terms of shred food material with soft texture on the cut 

results, analysis of quantitative and qualitative losses of material, determine the 

forces necessary to shred at different speeds (Olaru et al., 2011) 

The experiment was conducted under laboratory conditions, and as a 

working method was used texture analysis method (Olaru, 2012). 

With TA.HDPlus texture analyzer (Fig. 1) and specialized software 

Texture Exponent could get a full three-dimensional analysis of processed 

samples (different materials with soft texture) by force, distance and time. 

TA.HDPlus texture analyzer works through a cell resistant to a 

temperature field varying between (50…200) 0 C. The "Texture Exponent" runs 

under Windows and purchase view and analyzes data from TA.HDPlus texture 

analyzer. One of the most important characteristics of the analyzer is able to 

store as one of the procedures and have immediate access to it.

PC 

Cell calibration 


Fig. 1 – Texture analyser - TA.HDPlus. 

To determine the characteristics of texture, the "Texture Exponent" 

covers a wide range of probes and cutting devices (needles, knives of various 

shapes). To achieve experience was used probe HDP/BSK (with blade knife set) 

which has the classical form of knife which is suitable for shredding to cutting 

of food products with soft textures. 

Food materials studied were: curdled cheese, strawberries, peeled 

bananas and Turkish delight. For analyzing the crack propagation for soft 

textured materials studied on texture analyzer was used different values of 

speed range: 40 mm/s, 10 mm/s and 5 mm/s. 

In Figs. 2, 3, 4 and 5 is presented the variation of cutting force while the 

textured soft food shredding materials studied for different values of cutting 

speed. 

Force, [N] 

Time, [s] 

Fig. 2 – The cutting force variation in time to shredding cheese for different 

cutting speeds: 40 mm/s, 10 mm/s, 5 mm/s.


Force, [N] 

Time, [s] 

Fig. 3 – The cutting force variation in time to shredding strawberry for different 

cutting speeds: 40 mm/s, 10 mm/s, 5 mm/s. 

Force, [N] 

Time, [s] 

Fig. 4 – The cutting force variation in time to shredding peeled bananas for 

different cutting speeds: 40 mm/s, 10 mm/s, 5 mm/s. 

Force, [N] 

Time, [s] 

Fig. 5 – The cutting force variation in time to shredding Turkish delight for 

different cutting speeds: 40 mm/s, 10 mm/s, 5 mm/s. 

The analysis of graphs obtained can be seen that high-speed shredding 

resulting a high cutting forces and small times.


As can be seen from analysis of data obtained cutting force varies directly 

with changes in cutting speed, that at high cutting speeds are required high 

cutting forces. If shred strawberries (we used baked strawberry-sized) found 

that the highest values of cutting force were recorded for cutting speed of 40 

mm/s, the cutting force decreased gradually with decreasing cutting speed. 

As with other types of material and bananas have used the same cutting 

speed, thus observing that due to their homogeneous structure cutting force 

varies directly with the cutting speed. In case of Turkish delight shred highest 

values of cutting forces were obtained for small cutting speeds, this being due to 

the relatively homogeneous composition of these materials. 

Table 1 

Working indices obtained for the shredding of soft –textured food materials through 

probe HDP/BSK, with v = 40 mm/s. 

Food 

material 

Curdled 

cheese 

Average 

cutting 

surface 

a× b(d 2 ) 

mm 2 

Weight 

before 

cutting 

g 

Weight 

after 

cutting 

g 

Percentage 

of losses 

% 

Adherence 

to the 

cutter 

30 x 30 35.8 35.55 0.70 Small 

Strawberries 25 x 25 21.1 21.03 0.36 Small 

Peeled 

bananas 

Turkish 

delight 

Cutting 

aspect 

With 

roughness 

Few 

roughness 

30 x 30 35.16 35.04 0.34 Small Regular 

35 x 35 44.32 44.25 0.17 High 

Few 

roughness 

Table 2 



Food 

material 

Curdled 

cheese 

Average 

cutting 

surface 

a× b(d 2 ) 

mm 2 

Weight 

before 

cutting 

g 

Weight 

after 

cutting 

g 

Percentage 

of losses 

% 

Adherence 

to the 

cutter 

30 x 30 38.1 37.85 0.66 Small 


Peeled 

bananas 

Turkish 

delight 

Cutting 

aspect 

With 

roughness 

Few 

roughness 


35 x 35 45.5 45.44 0.13 High 

Few 

roughness


With the analytical balance and visual analysis of food subject to shred 

material appearance were followed a series of work indices such as: adherence 

to the working body, the cut, loss of juice. In Tables 1, 2 and 3 present the 

results obtained. 

Table 3 



Food 

material 

Curdled 

cheese 

Average 

cutting 

surface 

a× b(d 2 ) 

mm 2 

Weight 

before 

cutting 

g 

Weight 

after 

cutting 

g 

Percentage 

of losses 

% 

Adherence 

to the 

cutter 

30 x 30 34.6 34.4 0.58 Small 


Peeled 

bananas 

Turkish 

delight 

Cutting 

aspect 

With 

roughness 

Few 

roughness 


35 x 35 43.12 43.08 0.10 High 


Few 

roughness 

From the study the shred by cutting of food material with soft texture can 

be seen that: 

i) to high cutting speed and high cutting forces resulting small times; 

ii) with decreasing speed cutting below 10 mm/s due to specific 

properties of each material cutting times increase than in conjunction with lower 

cutting forces; 

iii) in terms of the cut body cutter used experiments conducted have 

shown a number of disadvantages related to product quality resulting from 

shred, as follows: a big loss especially for strawberry (0.24% ÷ 0.36%) and loss 

of material for cheese (0.58% ÷ 0.70%) and lower for bananas (0.20% ÷ 0.34%) 

and Turkish delight (0.10% ÷ 0.17%); large frictional force especially for 

Turkish delight and even bananas, material adhering to shred body; the cut 

aspects with rough one particularly curdled cheese; loss of the initial form of 

materials subject to shred in particular at cheese; 

iv) thus improving the quality of shred by cutting at food materials with 

soft texture in the further research is necessary to establish a correlation 

between the working parameters, textural characteristics of the products subject 

to shred, cutter geometry to obtain more regular cutting surfaces while reducing 

juice and material losses resulting from shred.


REFERENCES 

Brown T., Cutting Forces in Foods: Experimental Measurements. Journal of Food 

Engineering, 199-204 (2005). 

Brown T., Gigiel A.J., Swain, M.V., James C., Practical Investigations of Two-stage 

Bacon Tempering. International Journal of Refrigeration, 690-697 (2003). 

Brown T., James, S.J., Purnell G., Swain M. J., Improving Food Cutting Systems. 

Proceedings of IChemE Food and Drink, 2000, pp. 103-106. 

Charalambides M.N., Goh S.M., Lim S.L., Williams J.G., The Analysis of the Frictional 

Effect on Stress–strain Data from Uniaxial Compression of Cheese. Journal of 

Material Science, 13-21 (2001). 

Charalambides M.N., Williams J.G., Chakrabarti S., A Study of the Influence of Ageing 

on the Mechanical Properties of Cheddar Cheese. Journal of Material Science, 13 

-21 (1995). 

Ciumac J., Merceologia produselor alimentare. Ed. Tehnică, Chişinău, 1995. 

Fellows P., Food Processing Technology. Principles and practice. Sec. Ed., Woodhead 

Publishers, Gainthersburg, 1999. 

Imbeni V., Atkins A.G., Jeronimidis J., Yeo J., Rate and Temperature Dependence of 

the Mechanical Properties of Cheddar Cheese. Proceedings of 10th International 

Conference on Fracture, 2001 . 

Nedeff V., Procese de lucru, maşini şi instalaţii pentru industria alimentară. Bucureşti, 

Redacţia Revistelor Agricole, Ed. Agris, 1997. 

Olaru C., Contribuţii la studiul regimului de mărunţire a materialelor agroalimentare 

cu textură moale. Teză de doctorat, 2012. 

Olaru C., Nedeff V., Panainte M., Studies and Research on the Possibilities of Cutting 

Analysis of Food Materials With Soft Texture. Journal of Engineering Studies and 

Research, 1, 17, 70-76 (2011). 

Panainte M., Mosneguţu E., Savin C., Nedeff V., Echipamente de proces în industria 

alimentară. Mărunţirea produselor agroalimentare, Ed. Meronia şi Rovimed 

Publishers, 2005. 

Raven T., The Future of Texture Analysis. International Food Ingredients, 1, 42-44 

(1998). 

Reilly G.A., McCormack B.A.O., Taylor D., Cutting Sharpness Measurement: A 

Critical Review. Journal of Materials Processing Technology (2004). 

Shmulevich I., Nondistructive Texture Assesment of Fruitss and Vegetables. Acta 

Horticulturae, 599, 289-296 (2003). 

Walstra P., Van Vliet T., Rheology of Cheese. Bull. IDF, 22-27 (1982). 

STUDIUL COMPORTĂRII PRODUSELOR ALIMENTARE SUPUSE MĂRUNŢIRII 

PRIN TĂIERE PRIN METODA ANALIZEI DE TEXTURĂ 

(Rezumat) 

Lucrarea prezintă rezultatele unui studiu în care s-a urmărit realizarea unei 

analize în condiţii de laborator a procesului de mărunţire prin tăiere a produsele 

alimentare cu textură moale funcţie de: modului în care se comportă materialul supus 

mărunţirii, forţele necesare mărunţirii, aspectul tăieturii.


Metoda folosită în studiu a fost metoda analizei de textură, metodă ce permite 

studierea comportamentului vîscoelastic a produselor vegetale cu textură moale. 

Cercetările s-au realizat pe o gamă largă de produse alimenatre cu textură moale: 

brânză închegată, căpşuni, banane decojite, rahat alimentar. 

Rezultatele experimentale obţinute evidenţiază faptul că forţa de tăiere variază 

direct proporţional cu variaţia vitezei de tăiere, adică la viteze mari de tăiere sunt 

necesare forţe de tăiere mari, excepţie făcând rahatul alimentar unde valori mari ale 

forţei de tăiere s-au obţinut la viteze mici de tăiere acest lucru datorându-se structurii 

omogene a acestui tip de material. 

Textura produsului precum şi tipul, forma organului de lucru influenţează 

semnificativ produsul final, astfel din rezultatele obţinute în urma cercetărilor 

experimentale dezvoltate au arătat că: pierderi mari de suc s-au obţinut la mărunţirea 

căpşunilor iar la brânzeturi s-au înregistrat pierderi semnificative de material, de 

asemeni pentru această categorie de produse suprafaţa obţinută în urma mărunţirii 

prezintă multe asperităţi. Aderenţa materialului la organul de lucru s-a putut observa la 

mărunţirea rahatului şi chiar a bananelor. 

THE STUDY TO OBTAINING THE TRAJECTORIES OF SOLID 

PARTICLES ON AN OSCILLATORY FLAT SURFACE 

BY 

EMILIAN MOŞNEGUŢU ∗ , VALENTIN NEDEFF, OVIDIU BONTAŞ, 

NARCIS BÂRSAN and DANA CHIŢIMUŞ 

Received: 

Accepted for publication: 

„Vasile Alecsandri” University of Bacău, 

Abstract. In this article presents a method to determine the trajectory of a 

real particle on an oscillating surface. A special attention was given to recording 

the mode of trajectory and for this purpose we used two cameras of the same 

type. Tridimensional trajectory of the particle on the oscillating surface was 

achieved through SyntEyes program. And been it was also determine the real 

speed of the particle. 

Key words: flat surface oscillatory motion, the particle trajectory. 

∗ Corresponding author: e-mail: emos@ub.ro



Vegetable products from agriculture and some of the products resulting 

from various industrial processes (grinding, granulating, briquetting etc.) is a 

heterogeneous mixture. To separate such a mixture may use different methods, 

which are chosen taking into account the nature of the system and its properties 

(phase discontinuous nature, size, temperature etc.) (Moşneguţu et al., 2007). 

Separation of a mixture of solid particles on their size difference is the 

most popular method, used to both cleaning and sorting mixtures of particles 

(Nedeff et al., 2001). 

The separations of solid particles by size represent the most old 

separation technology and we find many studies about it. But the most 

significant studies have been done by Brereton and Dymott in 1973, Rose and 

English in 1973, De Pretis in 1977 and Ferrara in 1988, studies which aim, in 

especially, to increase the efficiency of separation (Tsakalakis, 2001), 

(Soldinger, 1999). 

However the separation of solid particles by size is not well known, in the 

specialty literature, have appeared three trends to study this process: 

- experimental studies (Trumic, 2011); 

- theoretical studies (Soldinger, 2002), (Li et al., 2003), (Stoicovici et 

al., 2008), (Szymański et al., 2003), (Xie, 2007); 

- studying the process through the simulation (Dong, 2009), (Zhao, 

2010), (Jianzhang & Xin , 2012) . 

From specialty literature study is established that the process separation 

of solid particles after their size depends on several factors (Tsakalakis, 2001): 

- size and shape aperture sieve; 

- particle size and form; 

- moisture particles undergo the process of separation; 

- intensity of vibration, respectively frequency and amplitude; 

- quantity of material that is fed sieve; 

- sieve length. 

These factors also influence the behavior of solid particle on to surface 

oscillating respectively its trajectory. 

2. Materials and equipment 

In the specialty literature there are very few experimental determinations 

which involving the use of real particles to determine the trajectory and speed of 

movement thereof on an oscillating surface. Therefore, the experimental 

determinations have performed using real particles, respectively particles of 

beans and using the worksheet - Triplot was determined form of these types of 

solid particle (Fig. 1).


To generate oscillatory movement was used laboratory stand equipped 

with a crank rod mechanism ( 

Fig. 2). Because the laboratory stand has tyrants horizontal and vertical 

the sieve block executes oscillations on three directions. 

It is very difficult to follow the solid particle on the surface oscillating, so 

it was used two video cameras, Sony DCR-SR 36, which has record speed is 25 

frames / second. Cameras have been placed in two perpendicular planes (Fig. 

3), in order to track particle movement on plans: 

- XOY – camera no. 1; 

- XOZ – camera no. 2. 

1 

2 

3 

4 

Fig. 1. Determination of solid particle shape: 

a - length, b - width, c – thickness. 

13


6 

5 7 8 

Fig. 2. Laboratory stand components: 

1 - flow control system, 2 - block the sieve, 3 - device to control of block angle of sieve, 

4 - frameworks, 5 - counterweight, 6 - eccentric the shaft 7 - horizontal longitudinal 

thrusts, 8 - belt, 9 - vertical rod, 10 - transverse horizontal thrusts, 11 - electric motor, 

12 - collection cassette of fractions, 13 - evacuation chute. 

During these measurements we used a blind screen inclined to the 

horizontal at an angle of 7. Inclined flat surface was subjected to oscillatory 

motion of which amplitude was: 

- on axis OX on 0.723 mm; 

- on axis OY on 0.38 mm; 

- on axis OZ on 1.01 mm. 

Values have been determined using a Vibrotest 60. 

To follow as possible the path exactly of particle solid, on the working 

surface was limited the distance on which has followed the travel of particle and 

was draw a marker line in order to be view the particle movements in sideways 

(Fig. 4). 


Z 

O 

X Y 

Fig. 3. The video cameras location. 

9 


10 

11 

12


Fig. 4. Marks used on oscillating surface. 

For obtaining the coordinates of the point followed, regardless of video 

camera position, was used software SynthEyes, such: 

- values obtained from the analysis of motion in the plane XOY (Fig. 

5) will be determine: 

- the travel time of solid particle on monitoring distance; 

- the solid particle trajectory on oscillating surface; 

- velocity of solid particle. 

- the second video camera (shooting in the plane XOZ) aims to 

determine the solid particle jumps on which made them at the time they travel 

on blind sieve (Fig. 6). 

Fig. 5. Manner of determining the 

coordinates of point in the plane XOY. 

3. Experimental results 

Fig. 6. Manner of determining the 

coordinates of point in the plane XOZ. 

In the processing of experimental data was taken into account:


- focus distance, aiming at the values obtained from the calculation 

correspond to real coordinates; 

- the angle of the plane surface. 

So after processing the data obtained with the help of SynthEyes, it could 

determine: 

A) Solid particle trajectory on the two planes tracked by video cameras 

(Fig. 7). 

B) Tridimensional trajectory of the solid particle moving on the 

oscillating flat surface, obtained by combining the two trajectories previously 

presented (Fig. 8); 

C) Knowing the time to follow the solid particle on oscillating surface 

and the distance traveled of particle after each frame it’s possible to determine 

the speed of movement (Fig. 9); 

Fig. 7. Moving the particle solid on oscillating surface. 

11.0 

Axis OY movement (mm) 

8.8 

6.6 

4.4 

2.2 

0.0 

0 

100 

200 

6.5 

0.0 

Axis OX movement (mm) 

Fig. 8. 3D trajectory of a particle on the oscillating surface. 

13.0 

300 

32.5 

26.0 

19.5 

Axis OZ movement (mm)


Fig. 9. Velocity of solid particle on surface oscillating 


The separation of solid particles after size is a complex process 

influenced by several factors, which depend of the particles and characteristics 

of equipment used. 

In the experimental measurements were used real particles and from 

analysis of the Triplot chart shows that particle is used as lamellar particle. 

Because tyrants, stand used in the experimental determinations executed 

oscillations along three axes. 

By using two video cameras with purpose to tracking the solid particle 

trajectory was be achieved, with help of SynthEyes program, obtaining a threedimensional 

trajectory. 

Also from experimental data could determine the variation of solid 

particle velocity on the oscillating surface. 

REFERENCES 

Dong K.J, Yu A.B., Brake I., (2009), DEM simulation of particle flow on a multi-deck 

banana screen, Minerals Engineering, 22, 910 – 920. 

Li J., Webb C., Pandiella S.S., Campbell G.M., (July 2003), Discrete particle motion on 

sieves—a numerical study using the DEM simulation, Powder Technology, 133, 1– 

3, 30, 190–202. 

Moşnegutu E., Panainte Mirela, Savin Carmen, Măcărescu B., Nedeff V., (2007), 

Separarea amestecurilor de particule solide în curenţi de aer verticali, Ed. Alma 

Mater Bacău. 

Nedeff V., Moşneguţu E., Băisan I., (2001), Separarea mecanică a produselor 

granulometrice şi pulverulente din industria alimentară, Ed. Tehnica-Info, 

Chişinău. 

Shilin Xie, Siu Wing Or, Helen Lai Wa Chan, Ping Kong Choy, Peter Chou Kee Liu, 

(January 2007), Analysis of vibration power flow from a vibrating machinery to a


floating elastic panel, Mechanical Systems and Signal Processing, 21, 1, , 389– 

404. 

Soldinger M., Transport velocity of a crushed rock material bed on a screen, (January 

2002), Minerals Engineering, 15, Issues 1–2, 7–17. 

Stoicovici D.I., Ungureanu M., Ungureanu N., Bănică, M., (2008), A computer model 

for sieves’ vibrations analysis, using an algorithm based on the “false-position 

method, American Journal of Applied Sciences, 5, 12, pag.48-56. 

Szymański T., Wodziński P., Screening on a screen with a vibrating sieve, (2003), 

Physicochemical Problems of Mineral Processing, vol. 37, 27-36. 

Trumic Milan and Magdalinovic Nedeljko, (2011), New model of screening kinetics, 

Minerals Engineering, 24, 42 – 49. 

Tsakalakis K., (2001), Some basic factors affecting scree performance in horizontal 

vibrating screens, The European Journal of Mineral Processing and Environmental 

Protection, I, 42 – 54. 

Xiao Jianzhang, Tong Xin, (2012), Particle stratification and penetration of a linear 

vibrating screen by the discrete element method, International Journal of Mining 

Science and Technology. 

STUDIUL OBŢINERII TRAIECTORIEI UNEI PARTICULE SOLIDE PE O 

SUPRAFAŢĂ PLANĂ OSCILANTĂ 

(Rezumat) 

O parte din produsele din agricultură sau rezultate din diferite procese industriale 

sunt supuse separării. Cea mai utilizată metodă de separare a unor astfel de produse o 

reprezintă separarea după dimensiunile particulelor care alcătuiesc amestecul eterogen. 

Cu toate că procesul de separare după dimensiuni este unul dintre cel mai răspândit, în 

literatura de specialitate sunt puţine studii care să prezinte modul de determinare a 

traiectoriei unei particule pe suprafaţa oscilantă a blocului de site. Articolul de faţă 

prezintă o metodă de determinare a acestei traiectorii, pentru o particulă reală, utilizând 

o sită oarbă care execută o mişcare oscilantă plană. Pentru determinarea traiectoriei 

particulei reale s-au utilizat două camere video care au avut drept scop filmarea 

deplasării particulei solide după planurile XOY şi XOZ. Cu ajutorul soft-ului SynthEyes 

s-au extras coordonatele punctului urmărit. Datele obţinute au fost prelucrate ţinându-se 

cont de distanţa de focalizare şi de unghiul de înclinare al suprafeţei plane. Prelucrarea 

datelor a avut drept scop determinarea traiectoriei reale a particulei şi viteza de 

deplasare a acesteia. 

THEORETICAL ECO-EFFICIENCY COMPARATIVE STUDY 

CASE, HYDROCARBONS, AMMONIA AND HFC MIXTURE 

ALTERNATIVES RETROFIT


BY 

GRAŢIELA MARIA ŢÂRLEA ∗2 , ION ZABET 1 , MIOARA VINCERIUC 2 

and ANA ŢÂRLEA 2 

1 Romanian General Association of Refrigeration of Bucharest 




Abstract. Several natural refrigerant alternatives are compared on the basis 

of their cycle coefficient of performance and TEWI factor (Total Equivalent 

Warming Impact – in respect with SR EN 378-1). 

These natural alternatives: ammonia and mixtures have zero or low global 

warming potential (GWP). 

The theoretical study uses as reference a single stage refrigeration system 

which works with R 404A. To implement the international Legislation, in the 

future it is necessary to retrofit HFC refrigerants with ecological refrigerants 

(with zero ODP and zero GWP). 

Energy efficiency is directly related to global warming and greenhouse 

gases emissions 

Key words: energy efficiency, TEWI, natural refrigerants, refrigeration 

system. 


The retrofitted single stage refrigeration system is used for chilled 

products in a cold storage (0…4°C). Initial refrigeration system is composed by: 

condenser, scroll compressors, evaporators, refrigerant vessel and thermostatic 

valves. 

The single stage refrigeration system was upgraded by replacing the 

scroll compressor (working with R404A) with open drive screw compressor 

(working with R717). The new refrigeration system is composed by: condenser, 

open drive screw compressor, ammonia vessel, ammonia pump, cooling oil 

vessel, and equalization column and ammonia evaporators. 

In the following lines the eco-efficiency comparative study between three 

refrigerants (R717, R404A and R507A) are described. The eco-efficiency 

comparative study consists in performance coefficient, annual energy 

consumption and TEWI factor (Ţârlea et al., 2010). 

2. Study Case 



In this section, the method and step calculation for eco-efficiency of 

single stage refrigeration system will be described. Fig. 1 shows the schematic 

representation of the single stage refrigeration system which must be retrofitted 

from R404A to R717. The R404A refrigerant from the discharge compressor 

port goes to the condenser (where appear the condensation process), then from 

the condenser outlet it goes into liquid receiver. From the liquid receiver the 

refrigerant enter in evaporator 1 and 2 (where appear evaporation process) and 

after the evaporator the refrigerant goes to suction port of the compressor. 

First the working conditions are described; secondly the theoretical ecoefficiency 

parameters (presented in the introduction) and thirdly the results of 

the theoretical study are shown. 

The working conditions are given in Table 1. The refrigeration system 

works 8 hours/day (nhours) and 261 days/year (ndays). The life cycle of the 

refrigeration system is 15 years (nyears). The refrigerant charge (m) is: R404A – 

14kg, R717 – 6.45 kg and R507A – 14.11 kg. The requirement electrical power 

(Pe) is shown in Table 1. The carbon dioxide emission (β) is 0,6kg/kWh. The 

recovery factor (αrec) is 75%. The refrigerant leak (msc) is 8%. 

The electrical power values were calculated and established by each 

refrigeration system, depending on properties and evaporation temperature 

(Zabet, 2011). 

Table 1 

Working conditions 

Refrigerant Φ0 (kW) 

M 

(kg) 

Tc (K) 

TSH (K) 

TSUB (K) 

T0 (K) 

263.2 273.2 

Pe (kW) 

278.2 

R404A 12.3 14 3.70 5.25 5.47 

R717 10 6.45 308.2 10 3 3.55 4.00 3.85 

R507A 10 14.11 

3.80 3.94 4.07 

where: Φ0 is refrigerant capacity, T0 evaporation temperature; TC condensing 

temperature, TSH superheating temperature; TSUB subcooling.


Fig. 1 – Schematic representation of refrigeration system: where: T1 1 – refrigerant 

temperature at suction compressor [ºC]; T2 T2 – refrigerant temperature at discharge 

compressor [ºC]; T3 T3 refrigerant temperature at condenser inlet [ºC]; T4 T4 – refrigerant 

temperature at condenser outlet [ºC]; T5 T5 – air temperature at inlet evaporator 1 [ºC]; T6 T6 

–air temperature at outlet evaporator 1 [ºC]; T7 T7 – refrigerant temperature at discharge of 

evaporator 1 [ºC]; T8 T8 – air temperature at inlet evaporator 2 [ºC]; T9 T9 – air temperature at 

outlet evaporator 2 [ºC]; T10 T10 – refrigerant temperature at discharge of evaporator 2 [ºC]; 

T11 T11 – air temperature at condenser inlet [ºC]; T12 T12 – air temperature at condenser 

outlet [ºC]. 

2.1. Calculation of the Theoretical Eco-efficiency Parameters 

The theoretical eco-efficiency parameters discussed in this paper are: the 

performance coefficient, annual energy consumption and TEWI factor. For 

finding the eco-efficiency parameters, classic thermodynamics thermodynamics relations 

together with parameters defined defined in Table 1 were used (Ţârlea (Ţârlea et al., 2010). 2010). 

Finding the performance coefficient: 

First of all was calculated the refrigerant mass flow by 

M& 

Φ 

= 

Δh 

0 , 

ev 

(1)


where Δhev 

is enthalpy difference at evaporation working condition [kJ/kg]; M 

is the refrigerant mass flow rate [kg/s]. 

Secondly, the mechanical power for the compressor was found, by 

(2) 

, 

W& = M& Δh 

cp cp 

where: Wcp is the mechanical power [kW]; 

& 

compression working conditions [kJ/kg]. 

Δ hcp 

is the enthalpy difference at 

Finally, the results of the coefficient of performance (COP) by 

Φ0 

COP = . 

W& 

Finding the annual energy consumption: 

The annual energy consumption (Eannual) is calculated according SR EN 

378-1 by 

E = n n P 

(4) 

cp 

annual hours days e. 

Finding the TEWI factor 

The TEWI factor is calculated by SR EN 378-1 by 

TEWI = GWP ⋅ Ln + GWP ⋅m1− α + n E β (5) 

( ) 

years rec years annual 

where: GWP – global warming potential [-]; L – refrigerant leak during a year 

working [kg]. 

Table 2 shows the global warming and ozone depletion potential values 

for analysed refrigerants (Ţârlea et al., 2011). 

Table 2 

Global warming and ozone depletion potential values 

Refrigerant GWP ODP 

R404A 3260 0 

R717 0 0 

R507A 3300 0 

(3) 

L= mscm. (6) 

3. Theoretical Eco-Efficiency Study Results 

The theoretical results are presented in the next lines by tables and 

figures. Table 3 shows the study results for the coefficient of performance, 

annual energy consumption and TEWI factor (Zabet & Ţârlea, 2011). 

As it is shown in Fig. 2, the electrical power for R717 refrigerant is 15% 

smaller than R404A and R507A and this electrical power savings has a great 

impact over the environment pollution and maintenance cost.


Table 3 

The study results 

Refrigerant COP [-] E annual [kWh] TEWI [tonsCO 2] 

R404A 4.484 6.328 7.724 10962 8352 8227 164.8 141.3 140.2 

R717 4.812 6.689 8.101 7726 7412 7934 69.5 66.7 71.4 

R507A 4.461 6.273 7.639 11421 8039 8498 170.3 139.9 144 

Fig. 2 – Electrical power versus evaporation temperature. 

Fig. 3 shows the variation of the coefficient of performance’s versus 

evaporation temperature. One could observe that the coefficient of performance 

for R717 refrigerant is higher by 6% than R404A and by 7% than R507A 

(Ţârlea et al., 2010). This is happens due to convenient thermodynamics 

properties of R717. 

Fig. 3 – Coefficient of performance versus evaporation temperature. 

Fig. 4 presents the variation of TEWI factor versus evaporation 

temperature. As shown, the TEWI factor for R717 refrigerant is 60% smaller 

than R404A and R507A. This implies that R717 refrigerant is the most suitable 

for environment from the global warming point of view.


Fig. 4 – TEWI factor versus evaporation temperature. 

Fig. 5 shows the annual energy consumption versus the evaporation 

temperature. It can be observed that the annual energy consumption for R717 

refrigerant is 15% smaller than R404A and R507A. 

Fig. 5 – Annual energy consumption versus evaporation temperature. 


1. The theoretical study analysed a single stage refrigeration system with 

R 404A as refrigerant. 

2. To improve the eco-efficiency, the HFC refrigerant was replaced with 

an ecological refrigerant R717. 

3. The comparative study of these facilities was based on the coefficient 

of performance of a refrigeration system and the TEWI factor.


4. The refrigerant R717 is more eco-efficient than R404A and R507A 

with a 15% lower electrical power consumption, 6% higher coefficient of 

performance, 60% lower TEWI factor and 15% lower annual energy 

consumption. 

5. Retrofitting the single stage refrigeration system by replacing R404A 

refrigerant with R717 it is proven that the savings are higher from the ecoefficiency 

point of view and the costs are lower (more equipment in the case of 

R717 than R404A). 

REFERENCES 

Ţărlea G.M., Vinceriuc M., Zabet I., Ţărlea A., Ecological Alternative for R404A 

Refrigerant. The 41st HVAC&R Congress KGH 2010 tome “Heating, 

Refrigerating and Air-Conditioning” 3-5 December, Belgrad, Serbia, 27-33 

(2010). 

Ţărlea G.M.,Vinceriuc M., Zabet I., Ţărlea A., Theoretically Study of Ecological 

Alternative for R404A, R507A and R22. The 42nd International Congress and 

Exhibition Heating, Refrigeration and Air-Conditioning 30 November – 2 

December, Belgrade (2011). 

Ţărlea G.M., Vinceriuc M., Zabet I., Ammonia as a Very Eco-efficient Romanian 

Refrigerant. International Congress, 5-7 May COFRET Iasi (2010). 



STUDIU TEORETIC COMPARATIV PRIVIND ÎMBUNĂTĂŢIREA DIN PUNCT 

DE VEDERE AL ECO-EFICIENŢEI UNUI SISTEM FRIGORIFIC CA URMARE A 

ÎNLOCUIRII HIDROCARBURILOR ŞI AMESTECURILOR DE HFC CU 

AMONIAC 

(Rezumat) 

Criza energetică majoră precum şi încălzirea globală care afectează în prezent 

economia mondială şi viitorul societăţii umane, impun creşterea performanţelor 

energetice şi ecologice ale echipamentelor şi sistemelor frigorifice şi de aer condiţionat. 

În acest scop pe plan mondial există un efort susţinut pentru reducerea emisiilor de 

dioxid de carbon rezultate din arderea combustibililor fosili şi a altor emisii de gaze cu 

efect de seră. 

In aceasta lucrare se prezintă un sistem frigorific intr-o treapta de comprimare cu 

vapori funcţionând cu agentul frigorific R404A, instalaţie care a fost optimizata prin 

înlocuirea cu agentul frigorific R717. 

Structura aleasa la redactarea lucrării prezente este următoarea: in prima parte s-a 

făcut o scurta descriere a condiţiilor de lucru, in partea a doua s-au calculat parametrii 

teoretici de eco-eficienţă (TEWI, COP, puterea electrica consumata si consumul anual 

de energie electrica) si in a treia parte s-au prezentat rezultatele obţinute in urma 

studiului teoretic.


Conform celor arătate in Fig. 2 consumul electric al sistemului frigorific 

funcţionând cu agentul frigorific R717 este cu 15% mai mic decât in cazul utilizării 

agentului frigorific R404A si R507A. Aceasta reducere aduce un câştig semnificativ din 

punct de vedere al costurilor precum si al impactului ecologic asupra mediului 

înconjurător. 

Conform celor prezentate in Fig. 3 se poate observa un coeficient de performantă 

mai mare in cazul agentului frigorific R717: cu 6% fata de R404A si cu 7% fata de 

R507A. 

In concluzie se poate afirma ca optimizarea sistemului frigorific cu vapori intr-o 

treapta de comprimare prin înlocuirea agentului frigorific R404A cu agentul frigorific 

R717 aduce un câştig din punct de vedere al eficientei energetice, costurilor si un 

avantaj din punct de vedere al mediului înconjurător prin reducerea gradului de poluare.

buletinul institutului politehnic din iaşi - Universitatea Tehnică ...

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