contents · significance. their further development is of great practical importance to improve the...

Fiabilitate si Durabilitate - Fiability & Durability No 1/ 2015 Editura “Academica Brâncuşi” , Târgu Jiu, ISSN 1844 – 640X

1

CONTENTS

Pag. 1. Feodor NOVIKOV - ANALYTICAL PARAMETERS DETERMINATION

OF CUTTING POWER TENSION……………………………………………………………………………….. 3

2. Sergey BRATAN, Vladimir BOGUTSKY, Alexander BUZKO, Yury GUTSALENKO - ANALYSIS AND COMPARATIVE DESCRIPTION OF PREPARATION METHODS OF SMALL CYLINDRICAL CONNECTIONS PRECISION SURFACES TO THE ASSEMBLY………………………

9

3. Liliana LUCA, Iulian POPESCU, Constanța RĂDULESCU - STRUCTURE AND KINEMATICS OF A MECHANISM WITH STOPS............................................................................................................

21

4. Iulian POPESCU, Liliana LUCA - OPTIMIZATION OF A MECHANISM WITH STOPS 29 5. Dan ILINCIOIU, Cosmin MUSCALAGIU, Cosmin-Mihai MIRIŢOIU - ANALYSIS OF THE

MECHANICAL STRENGTH OF A DRIVING MECHANISM CALLED SHOCK………………………… 37

6. Bianca Silvia ORHEI, Daniel APOSTOL - THE STUDY ON INCREASING THE RELIABILITY OF WHEELS AT RAIL VEHICLES WITH DIESEL-ELECTRIC TRACTION ...............................................

43

7. Cătălin ALEXANDRU - OPTIMAL KINEMATIC DESIGN OF A CAR AXLE GUIDING MECHANISM IN MBS SOFTWARE ENVIRONMENT………………………………………………………..

49

8. Cosmin-Mihai MIRIŢOIU - LOSS FACTOR AND DYNAMIC YOUNG MODULUS DETERMINATION FOR COMPOSITE SANDWICH BARS REINFORCED WITH STEEL FABRIC.....

56

9. Cosmin-Mihai MIRIŢOIU - RESEARCH REGARDING THE BREAKING STRENGTH AND THE SECTION MACROSCOPIC SHAPE FOR COMPOSITE SANDWICH BARS REINFORCED WITH STEEL WIRE MESH...................................................................................................................................

63

10. Păun ANTONESCU, Ovidiu ANTONESCU, Constantin BREZEANU - THE GEOMETRY OF THE SPATIAL FOUR-BAR MECHANISM AND OF ITS PARTICULAR FORMS…………………………

70

11. Ioana POPESCU, Ovidiu ANTONESCU, Păun ANTONESCU - STRUCTURAL AND GEOMETRICAL ANALYSIS OF THE LIFTING MANIPULATORS FOR A GREEN ENVIRONMENT...

77

12. Ovidiu ANTONESCU, Viorica VELISCU, Daniela ANTONESCU - PLANAR MECHANISMS USED FOR GENERATING CURVE LINE TRANSLATION MOTION.....................................................

84

13. Ovidiu ANTONESCU, Viorica VELIȘCU, Constantin BREZEANU - MAIN TYPES OF MECHANISMS USED AS WINDSHIELD WIPER……………………………………………………………

91

14. Constantin BREZEANU, Ioana POPESCU, Păun ANTONESCU - TOPOLOGICAL STRUCTURE OF CONNECTING MECHANISMS IN THE ELECTRIC GRID...............................................................

99

15. Dan MESARICI, Viorica VELIȘCU, Daniela ANTONESCU - CABLE MECHANISMS USED FOR ACTUATING CAR ELEVATORS WITH 2 AND 4 POLES.........................................................................

106

16. Catălin ROŞU - EXPERIMENTAL INVESTIGATION REGARDING THE STRESS VALUES FROM LATHE CUTTERS DURING THE MANUFACTURING OF STEEL SHAFTS ........................................

112

17. Catălin ROŞU - DIRECT CALCULUS FORMULAS FOR THE LATHE TOOL EQUIVALENT STRESS VALUES DURING THE MANUFACTURING OF STEEL SHAFTS .........................................

119

18. Viorica VELIȘCU, Daniela ANTONESCU, Dan MESARICI - MECHANISMS USED FOR DRIVING WINDOWS OF CAR SIDE DOORS………………………………………………………………..

126

19. Viorica VELIȘCU, Dan MESARICI, Păun ANTONESCU - TOPOLOGICAL STRUCTURE AND MOBILITY OF THE MECHANISMS USED IN CAR MECHANICAL JACKS.........................................

133

20. Mihaela ISTRATE, Monica BÂLDEA, Ancuța BĂLTEANU, Jan Cristian GRIGORE - STUDY ON THERMAL DEFORMATIONS OF THE PRIMARY SEALING OF FRONT SEALING

140

21. Monica BÂLDEA, Mihaela ISTRATE, Alexandru BOROIU, Andrei Alexandru BOROIU - STUDY ON MECHANICAL DEFORMATIONS OF THE PRIMARY FRONTAL SEALING……………

147


2

22. Jan Cristian GRIGORE, Monica BÂLDEA, Mihaela ISTRATE, Ancuţa BĂLTEANU - VIBRATIONS AND EQUILIBRIUM OF THE PLANAR KINEMATIC CHAINS WITH ROTATIONAL KINEMATICAL LINKS WITH CLEARANCES..........................................................................................

154

23. Mihaela ISTRATE, Jan Cristian GRIGORE, Monica BÂLDEA, Ancuța BĂLTEANU - STUDY ON THE THERMODYNAMIC ASPECTS OF FRONT SEALING…………………………………………….

158

24. Gabi ROSCA FARTAT, Constantin POPESCU, Constantin D. STANESCU - CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS IN THE CANDU 6 NUCLEAR REACTOR. PART 6 - PRESENTATION OF THE DECOMMISSIONING DEVICE ...................................................

165

25. Gabi ROSCA FARTAT, Constantin POPESCU, Constantin D. STANESCU - CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS IN THE CANDU 6 NUCLEAR REACTOR. PART 7 - FUNCTIONING OF THE DECOMMISSIONING DEVICE ....................................................

172

26. Constantin POPESCU, Gabi ROSCA FARTAT, Constantin D. STANESCU - CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS IN THECANDU 6 NUCLEAR REACTOR. PART 8 - PRESENTATION OF THE CUTTING AND EXTRACTING DEVICE......................................

181

27. Constantin POPESCU, Gabi ROSCA FARTAT, Constantin D. STANESCU - CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS IN THE CANDU 6 NUCLEAR REACTOR. PART 9 - CUTTING AND EXTRACTING DEVICE FUNCTIONING......................................................

189

28. Constantin D. STANESCU, Gabi ROSCA FARTAT, Constantin POPESCU - CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS IN THE CANDU 6 NUCLEAR REACTOR. PART 10 - PRESENTATION OF THE DECOMMISSIONING DEVICE OPERATING ...........................

195

29. Constantin D. STANESCU, Ing. Constantin POPESCU, Gabi ROSCA FARTAT - CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS IN THECANDU 6 NUCLEAR REACTOR. PART 11 - PRESENTATION OF THE CUTTING AND EXTRACTING DEVICE OPERATING..............

205

30. Stefan GHIMIȘI - CONSIDERATIONS REGARDING THE FRETTING PHENOMENON USING LEAF SPRINGS............................................................................................................................. ............

212

31. Corneliu MOROIANU - STRESSES DETERMINATION METHOD IN MOVING PARTS OF THE MARINE ENGINES.................................................................................................................................

220

32. Corneliu MOROIANU - THE MARINE HEAVY FUEL IGNITION AND COMBUSTION BY PLASMA.....................................................................................................................................................

225


3

ANALYTICAL PARAMETERS DETERMINATION OF CUTTING POWER TENSION

Prof. PhD. Eng. Feodor NOVIKOV

Kharkov National University of Economics [email protected]

Abstract: analytically identified components of the cutting force, cutting and conventional stress conditions reduction for blade and grinding. Substantiated conditions for transition from the cutting process to the process of plastic deformation of the material being processed without the formation of chips.

Keywords: cutting process, grinding, cutting force, processing productivity

1. INTRODUCTION Of the various materials processing technologies based on the use of various forms of

energy, it is necessary to highlight the mechanical machining materials processing technologies [1, 2], which are characterized by the lowest power consumption and highest processing capacity and ensure high quality and accuracy of machined surfaces. Cutting processes materials are widely used in manufacturing and in the coming years won't lose its significance. Their further development is of great practical importance to improve the accuracy, quality and machining performance. Accordingly in the work considered questions analytical determination of parameters of the power intensity of the cutting process and the conditions of reduction applied to the blade and grinding.

2. ANALYTICAL RESEARCH The cutting process is directed destruction of the surface layer of the treated material

under the action of the load arising from contact with the cutting tool. This load is called the theory of cutting the cutting force. Usually it is resolved into two components - the tangential

zP and radial yP (fig. 1):

О

V

АС

1

2

3

Nа

yPР

zP

Nf

Fig. 1: Design scheme of the cutting process: 1 - processed material, 2 - chips, 3 - cutting tool

mailto:[email protected]


4

cutz V

QSP ; (1)

cutcutcut

zy V

Q

KK

PP

, (2)

where j - conventional voltage cutting; S - the cross sectional area of cut; cutVS=Q -

processing performance ; cutV - cutting speed; yzcut P/P=К - coefficient of cutting.

The most important condition of efficiency of the cutting process is to increase

processing performance Q . On the basis of the converted according (1): jVP

=Q cutz , this is

achieved by increasing parameters zP , cutV and reduction j . Obviously, the increase of the

tangential component cutting force zP limited durability of the cutting tool, and increase the

cutting speed cutV - his resistance due to the temperature factor. The possibility of reduction

of nominal voltage cutting j connected with the contact management processes, taking place on the front and back surfaces of the instrument. Therefore, parameters j and zP due

mechanics, the cutting speed cutV - the thermophysics of the cutting process. However, this

separation is very relative, because the parameters j and zP also temperature depends on cutting and they must learn how to position mechanics, and from a position of Thermophysics of the cutting process. About the effectiveness of processes for mechanical and physical-technical materials processing judged by the value of energy consumption Е , equal to the work expended on a unit volume of material removal V . Presenting the work А , necessary for removing bulk material V , at machining and volume V as lP=A z ; Sl=V (where l - length of cutting

path , м; S - the cross sectional area of cut , м2), energy consumption V/A=Е determined

j=S

P=Е z . (3)

As can be seen, when machining (cutting, cutting and abrasive tools) energy intensity Е numerically equal to the conventional stress cutting j .

To determine j should consider a simplified analytical model of the cutting process (fig. 1), according to which the material removal occurs by shifting elements formed along the conditional chip shear plane, located at the angle く to the direction of movement of the

tool. Under the influence of forces zP and yP in the shear plane shear stress occurs k . To

determine the position of the shear plane of the material conditional (contingent shift angle material く ) it is necessary to project the components of the cutting force zP , yP and set the

shear stress k : くsinsin2くК0,5вa

P=くsinPくcosP

вa

くsin=k 2

cuty

yz , (4)


5

where ва, - accordingly, thickness and width of cut , м.

Shear stress k , based on dependencies (4), has the most from conditional shift material く . On the fig. 2 graphically the character of change

of the function くsinsin2くК0,5=k 2cut0 , a

member of the dependence (4). The extreme value く is determined from

the necessary conditions of extremum 0=k'く :

cutK=2くtg . (5)

As you can see, the relative shift of the material clearly depends on the coefficient of cutting cutК : with increasing angle く increases.

Radial component cutting force yP is determined from the dependence (4) with

relation (5) on condition shk=k :

2

cut2cut

sty К+1+1

Квa

=Pσ

, (6)

where stsh j0,5k - tensile strength the shift of the processed material, Н/м2; stj – tensile

strength at compression of processed material , Н/м2. Tangential component cutting force ycutz PК=P and conditional voltage cutting

вa/P=j z are defined:

2

cutcut

stz К+1+1

Кjвa

=P ; (7)

2cut

cut

st К+1+1Кj

=j . (8)

Specific components of the cutting forces describes dependencies:

2cut

cutst

zspz К+1+1

К1

=jва

P=P ; (9)

2cut2

cutst

y

spy К+1+1К

1=

jваP

=P . (10)

From dependences (9) and (10) it follows that the parameters spzP and j identical,

asdescribes the same addiction. The analysis calculated based on the dependencies (9) and (10) values

spzP and spyP , which is shown in fig. 3, shows that, subject to the cutК =

1parameters spzP and

spyP equal, as provided cutК <1 и cutК >1 fair accordingly conditions

spzP <spyP и

spzP >spyP . As is known, the condition cutК < 1 implemented for abrasive

0 10 20 30

0,05

0,10

0,15

0,20

0

.ÇëаÑ,

1

2

Fig. 2: Dependency 0k on the angle く :

1 - cutК =0,5; 2 - cutК =1.


6

processing, but the condition cutК >1 – at the blade processing. Therefore, when sanding the

greatest influence on the process parameters has a radial yP component cutting force, but

when the blade processing - tangential zP component cutting force. Thus the components of

the forces at the edge processing ( cutК >1) less than for abrasive processing ( cutК <1), that

indicates the possibilities of improving the accuracy and quality of cutting edge tools. This pattern is due to the smaller values of the conditional voltage cutting j =

spzP , which

provided cutК seeks to adopt the values stj (fig. 3). In this case the conditions of chip

formation comply with the terms of straightforward destruction of the sample at its compression.

Express the ratio of cut cutК through the components of the cutting force N и Nf ,

appearing on the front surface of the tool (fig. 1) [3]:

,cosNfsinNP

,sinNfcosNP

y

z

(11)

where f - the coefficient of friction on the front surface of the tool; け - cutting angle of the tool. Whence

けψctg=Kcut tgけf

tgけf+1=

, (12)

where tgψ=f ; ψ - conditional friction angle on the front surface of the tool.

Graphically, the dependence (12) shows on the fig. 4. The coefficient cutting

cutК - plus, changes in limits from 0 up to 0.

Comparing dependence (5) and (12), the obtained dependence for calculation of the conditional shift material く :

2

ψけ+45°=く

. (13)

Angel けψ=の is called the angle of

the. So the options zP , yP and j can be

expressed also through the angle of the けψ=の :

1+けψtg+けψtgjвa=P 2stz ; (14)

1+けψtg+けψtgけψtgjвa=P 2sty ; (15)

0 1 2 3 4

1

2

4

6

,Päóтz

äóтyP

ëі£К Fig. 3: Dependency

spzP (the solid line) и spyP

(the dotted line) rate cuts cutК .


7

1+けψtg+けψtgj=j 2st . (16)

Obviously, edge processing is implemented,

provided 045<けψ=の , and abrasive treatment on

condition 045>けψ=の (given the negative values

of angle け – condition 045>け+ψ=の ). From dependence (5) and fig. 4 it follows that

the condition cutК (or stjj ) performed at an

angle of actions けψ=の 0 (or condition tgけ=f ). Most simply, this condition is realized when

processing the diamond tool, because diamond has the lowest coefficient of friction f with the processed material. This explains the possibility of a significant reduction of forces in diamond turning, which takes place in practice.

In dependencies (14) – (16) the angle of the けψ=の you must look at the positive.

Obviously, with the increase in the angle-action けψ=の parameters yP , zP and j

increase. For them reduce it is necessary the positive angle of the tool け increase, and the

friction angle ψ (coefficient of friction f ) reduce.

At values 1>けψtg radial component cutting force yP more tangential component

cutting force zP . This is possible when the angle of action °45>けψ , sowith a relatively

large contingent angle of friction °45>ψ and small value 0け , or negative values of the

front angle け , then the angle actions けψ=の will be obviously more °45 . The last case is implemented, for example, when cutting with abrasive and diamond grains grinding wheel [4]. The more the degree of dulling of grains, the more negative the front angle け and the higher the ratio zP / yP . This is consistent with the practice of cutting. It

is established experimentally that when the blade processing is executed, as a rule, condition

zP > yP , so 1>け+ψtg .

At the initial moment of treatment (at the moment of cutting blade tool in the processed material), when almost no friction generated chip with the front surface of the tool( 0ψ ), tangential component cutting force zP more, than at the steady state cutting

process, when there is friction chip with the front surface of the tool and 0>ψ . This is

because in the first case, when °40<け<10° values けψtg more, than in the second case,

at °40<ψ<10° – 0けψtg .The lowest value zP is reached at け=ψ .

Increase zP at the moment of cutting tool in the material can be one of the factors determining the low efficiency of the instrument for discontinuous cutting. The presence of

045 0900180

ëі£К

1

0

Fig. 4: Dependency cutК from the

angle けψ


8

forces yP at the steady state cutting process allows essentially to reduce zP and improve the

efficiency of the instrument. Negative values of the front corner of the instrument け (and zero) tangential

component cutting force zP at the moment of cutting is less, than at a steady process of cutting. Consequently, the use of instruments with a negative front corners of the fundamentally changes regularities tensions process, excludes increase zP at the moment of cutting. According to (14) - (16) with relation (5) and trigonometric transforms simplified and take the form:

tgく2くtg

jвa=P st

y

; tgく

jвa=P st

z

; tgくj

=j st . (17)

As you can see, the options yP , zP , j quite clearly depend on conditional shift

material く . The more く , the less yP , zP , j and effectively cutting process.

At grinding ( cutК < 1), looking forward corner cutting grain circle け negative [5],

dependency (14) – (16) can be simplified

cut

stz К

jвa2=P

;

2cut

sty К

jвa2=P

; け+ψtgj2=

Кj2

=j stcut

st . (18)

on condition 090け+ψ rightly j . In this case the cutting process (chip disposal) is not implemented, only the elastic-plastic deformation of the material without the chip formation. Therefore, the process of cutting edge cutting and abrasive tools may be made,

provided 090<け+ψ=の . Using this condition, you can scientifically grounded approach to the selection of optimum processing conditions, including processing methods, mode settings cutting tool characteristics, etc. Thus, in the solution of important topical the problem of determination of parameters of power tension cutting process, and the conditions of decrease in relation to edge cutting machining and grinding.

REFERENCES [1] Bobrov, V. F. Fundamentals of the theory of metal cutting. – Moscow: Mashinostroenie,

1975. – 343 p. – In Russian.

[2] Granovsky, G. I. Cutting of metals. – Moscow: Higher school, 1985. – 304 p. – In Russian.

[3] Rosenberg, A. M., Eryomin, A. N. Elements of the theory of the process of metal cutting. – Moscow-Sverdlovsk, 1956. – 318 p. – In Russian.

[4] Maslov, E. N. Theory of grinding of metals. – Moscow: Mechanical Engineering, 1974. – 319 p. – In Russian.

[5] Novikov, F. V. Physical and kinematic bases of high-performance diamond grinding / Tesis for a Doctor's degree. − Odessa, 1995. – 36 p. – In Russian.


9

ANALYSIS AND COMPARATIVE DESCRIPTION OF PREPARATION METHODS OF SMALL CYLINDRICAL CONNECTIONS

PRECISION SURFACES TO THE ASSEMBLY

Prof. PhD. Eng. Sergey BRATAN1, Sen. Lecturer PhD. Vladimir BOGUTSKY1, Eng. Alexander BUZKO1, Sen. Staff Scientist Yury GUTSALENKO2

1Sevastopol State Univ., [email protected] 2Nat. Tech. Univ. “Kharkov Polytechnic Inst.”, [email protected]

Abstract: The article describes the methods of finishing precision surfaces of small cylindrical connections. Keywords: precision surface, confidence limit, wear resistance of surface, finishing, spinning 1. INTRODUCTION Surface preparation methods of small cylindrical connections to the assembly of high

accuracy have its specialty. All finishing operations on the seating surfaces are performed on the almost ready, highly complex parts or the whole subassembly, the cost of which at this stage of the process is very high. It currently used methods are primarily lapping and machining of parts by plastic surface deformation (PSD).

Lapping provides surfaces precision in grade 6 and above and roughness to Ra = 0.063 microns. Available literature data about lapping (including the equipment, tools, modes) are mainly cases of finishing medium and higher size parts [1, 2]. So in this work we paid attention to the small (up to 8 microns) diameter surfaces lapping.

The purpose of the article is a comparative evaluation of surface preparation for assembly methods of small cylindrical connections increased accuracy.

2. THE MAIN CONTENT AND RESULTS Processing parts by PSD method is used to improve the microhardness, increase

surfaces and improve the wear resistance of surfaces [3, 4]. Precision machining for PSD-hanging depends on the design features of parts, shape and quality of the original surface, tools, modes of processing, precision size. Original surfaces with 5...7 accuracy degree are subjected of burnishing and spinning, so that allows to advance dimensional accuracy to 10...20% and to reduce the shape deviations to 10... 30 microns.

Finishing allowance can reach relatively large quantities. Figure 1 shows the allowances Z landfill distribution, skimmed in finishing micromachines shafts trunnions diameter 4...6 microns (Fig.1,a), and holes with diameter 11...15 microns (Fig.1,b) made of material 14X17H2 and 17X18H9. Pretreatment is a grinding to Ra = 3.2 ... 1.6 microns.

Confidence limits for the mean values are built on t - t-test with a confidence level of 0.95 [3].

To investigate the intensity of material removal during lapping, a number of statistical tests were made. Size measurements were made with optimeter ICG. The test results were summarized in the correlation table 1. Further calculations were carried out in accordance with [3].




10

Fig. 1: Allowances allocation polygons for finishing

Table 1: Correlation table of test results Time of

grinding t, sec Allowance of Z, micron Numbers of columns

0,5 1 2 3 4 5 6 7 8 1 2 3 mt mt∙t mt∙t2

2 1 - - - - - - - - 1 2 4 4 1 1 1 1 - - - - - 4 16 64 6 - 2 3 - 1 - - - - 6 36 216 8 - - 1 1 - - - - - 2 16 128

10 - - - - 1 1 1 - 1 4 40 400 12 - - - - 1 1 1 - 2 6 72 864 14 - - - - - 1 1 - 2 28 392 16 - - - - 2 - - - - 2 32 512 18 - - - - - - 1 - - 2 36 648 20 - - - - - - - - 1 1 20 400

1 nz 2 3 5 3 5 2 4 1 4 ∑(1)=29 ∑(2)=298 ∑(3)=3628 2 nz∙z 1 3 10 9 20 10 24 7 32 ∑ (2) = 116 3 nz∙z2 0.5 3 20 27 80 50 144 49 144 ∑ (3) = 522.5 4 ∑nzt∙t 6 16 30 24 60 22 54 14 54 ∑ (4) = 280 5 Z∑nzt∙t 3 16 60 72 240 110 324 98 432 ∑ (5) = 1355 6 3 5.34 6 8 12 11 13.5 14 13.5

Using data from Table 1 we find the average values of Z and removes the allowance of

time for this procedure t, correlation Czt, average square deviations σz and σt, as well as the correlation coefficient Rzt.


11

microns96,329

115 z

z

n

znz ; sec25,10

29

298 y

i

m

tmt ;

3,625,1096,3

29

1355 tZn

tnZで

z

ztzt

;55,196,329

5,522 222 Z

n

Zn

z

zz

70,425,1029

3628t

m

tm22

t

2t

z

86,070,455,1

3,6CR

tz

ztzt

The correlation coefficient characterizing the relationship between the lapping machine time and allowance is not equal to one. This is due to the variation of finishing details microhardness and the operator’s fear to remove too much envelop of metal. According to the industry, which produces equipment for shipbuilding, attempts to mechanize the process did not lead to positive results due to weak coupling Z and t.

Due to the fact that the ratio of t and Z isn’t straight, we need to verify a process to the presence of the curvilinear link. Suppose there is a parabolic relationship, described by the equation:

2210 tataaZ (1)

where: a0, a1, a2 – constant coefficients. To find the unknown parameters of equation (1) - a0, a1, a2 we use the Chebyshev

interpolation formula, where the argument is the value: tt , ÇÑñ n

tt i

i ( n - the

number of tests): (x)qk(x)qk(x)qk y 221100 (2)

Experimental data for the calculation and preliminary calculations are summarized in

Table 2. The definition of a zero degree parabola is:

85,311

3,420

n

Zk i 3.85(x)qk of(x) 1,x(x)q 00

00

The fundamental error of σ0 is:


12

71,72

3,4271,235

22

2

nn

ZZ i 67,2

111

71,72

10

0 n

Parameter determination of the first degree parabola:

;1011

110 n

tt ii ; ;,

χχZ

ki

ii 350440

15021

)(q 11 ; 0.353.85 )(qk)0f( )1f( 11

The fundamental error σ1 is:

18.71=4400,35 - 72.71=k-= 22i

2101 44,1

211

71,18

21

1 n

Due to the fact that σ0 considerably exceeds σ1, we’ll continue interpolation. The definition of the second degree parabola is:

;4011

4402

2 n

A i ;0440

02

3

2 i

ib

137284404003132822

32

42 iii χAχbχで

00845,013728

044085,32,1579

2

31

20

22

C

kkzk iii

40;-xA-xb- x (x)q 222

22 40)-0.00845(x(x)qk 2

22

Table 2: Test data Time of

grinding t, sec

Allowance of Z, micron tm Z 0.5 1 2 3 4 5 6 7 8

2 1 - - - - - - - - 1 0,5 4 1 1 1 1 - - - - - 4 1,4 6 - 2 3 - 1 - - - - 6 2,0 8 - - 1 1 - - - - - 2 2,5

10 - - - - 1 1 1 - 1 4 5,7 12 - - - 1 1 1 1 - 2 6 5,7 14 - - - - - - 1 1 - 2 6,5 16 - - - - 2 - - - - 2 4,0 18 - - - - - - 1 - - 1 6,0 20 - - - - - - - - 1 1 8,0 nz 2 3 5 3 5 2 4 1 4 29

∑ nz∙t 6 16 30 24 60 22 54 14 54 280

3.0 5.34 6.0 8.0 12.0 11.0 13.50 14.0 13.5


13

Make the substitution in accordance with (2) to find f(x). After transformations we obtain:

2

2 0.00845x-0.35x4.188 f(x) (3)

The fundamental error σ2 is:

∑2=∑1-k22C2=1871-(-0.00845)2∙13728=17.735 42,1

311

735,17

32

2 n

Carry out replacing x = t-10 in the equation (3) to find Z (t). After conversion, we get:

0.00845t0.52t0.157- Z(t) 2

10;0

20;1

t

t (4)

Fig. 2: The material removal rate during lapping

Graphical interpretation of Table 3 is shown in Fig. 2. Table 3 shows the comparison of experimental and calculated values of Z.

Table 3: The experimental and calculated values of Z

t 2 4 6 8 10 12 14 16 18 20 zexp. 0,5 1,4 2,0 2,5 5,7 5,7 6,5 4,0 6,0 8,0 zcаlc. 0,85 1,79 2,66 3,49 4,20 4,86 5,48 6,00 6,45 6,86

.exp

.exp

z

zz ca lc -0,7 -0,2 -0,33 -0,4 0,26 0,15 0,15 -0,5 -0,08 0,14


14

Figure 2 shows the removal stock swings ΔZ at fixed values of time t with a confidence level of 0.95.

ΔZ=C∙σ2 (5)

where: ΔZ is a deviation of the average values of the removed metal in the interest

intervals, C - ratio of the average square deviation to the interval swing (in connection with the fact that in some cases the number of intervals is less than 6, then C = 2,3), σ2 - average square deviation counted in equation (4) development.

The wide confidence zone indicates instability of the lapping process in time. This requires operator great skill and greatly complicates the process of mechanization.

Similar experiments were carried at the facility for finishing and restoration rotors pins developed at the Department of Manufacturing Engineering FSFEI HPE "Sevastopol State University" (Fig.3) PSD was using as reworking operation.

Chiseled and polished pins Ø4 ... 6 mm, a length of 4...5 mm of material 30X13, 14X17H2, 17X18H9 (HRC = 26...30) with the initial roughness Rz = 3,2 ... 1,6 microns were chosen for the experimental spinning. A total number of 2,000 pins was spinned, at the rate of 27 ... 30 pins on one mode of the spinning, which corresponds to a confidence level of β = 0,95.

By analogy with the lapping amount of the diameter change in a spinning was conditionally named removable allowance.

Fig. 3: General view of the complex for finishing and restoration micromachines rotors pins by plastic deformation

Spinning sharpened pins. Correlation evaluation such as the above was made. It was found

that the value of the allowance Z and the spinning time t are related with curvilinear relation. Moreover, the maximum correlation coefficient is observed when spinning force is P = 60...90N.


15

Table 4: The average experimental Z (microns) allowance removed values for spinning sharpened pins Ø4 ... 6 mm

t, sec Р, N

0 3 6 9 12 15 20 25 40 60 rzt confidence limit at β=0,9

30 0 0,5 0,8 0,9 1,0 1,1 1,2 1,2 1,2 1,2 0,52

0,57÷0,85

60 0 1,2 1,5 2,0 2,1 2,3 2,3 2,3 2,3 2,3 0,78 90 0 1,5 1,8 2,1 2,5 2,9 3 3 3 3 0,71 120 0 1,9 2,6 3,3 3,5 3,9 4,0 4,0 4,0 4,0 0,58 150 0 2,1 3,5 3,9 4,7 5,0 5,1 5,1 5,1 5,1 0,56 300 0 5,5 6,0 9,0 10,0 10,0 10,0 10,0 10,0 10,0 400 0 11 11,5 12,0 12,0 12,0 12,0 12,0 12,0 12,0 500 0 12,5 12,7 12,7 12,7 12,7 12,7 12,7 12,7 12,7

Table 4 shows the average experimental Z allowance removed values for spinning,

correlation coefficients rzt and confidence limit for the maximum rzt, calculated in accordance with [3].

Table 4 shows that Z allowance taken from the operating force P <300N depends on the spinning duration only the first 15 seconds. After this period Z allowance is almost unchanged. Due to the fact that, for t> 15 sec. Z and t values are uncorrelated, the rzt correlation coefficients were found for spinning process with t≤ at 15 seconds. Allowance are removing almost in the first 3 ... 6 seconds if the rollers pressing efforts was above 300N. Calculations of rzt correlation coefficients of confidence limits for the allowance Z are similar to the previous estimates.

Fig. 4 is a graphical depiction of relation Z with the spinning time, and confidence limits for the Z allowance are shown for spinning force 60N.

Further analysis of the experimental data from table 4 makes possible to calculate the dependence of Z (P) for the duration of the spinning t = 15 sec and allowance confidence limits with the method of calculation used above.

Z(t) = 0.35t – 0.015t2 (6)

sec15;0

60sec;15;1

t

NPt


16

Fig. 4: Comparison data of experimental and calculated by the formula (6) values of removed

during spinning allowance Z microns

Fig. 5 is a graphical interpretation of the calculated dependence Z (P), and average experimental values and confidence limits for the allowance Z are applied.

Z (P) = 0.18+0.0323P (7)

оthertNP

;0

sec15;30030;1

Fig.5: Comparative data of experimental and calculated according to the formula (7) value of removed during spinning allowance Z microns

Based on the deduced data (in) we could build a multiple regression equation:

z=a1t+a2P+a3t∙P+a4t2+a5P

2

where: a1, a2, a3, a4, a5 – regression coefficients; t – machining time; P – spinning effort.

Regression coefficients are found by the least squares method:

min2

i1i

UZZU


17

where: Zi - the Z experimental value.

The first derivatives for this equation are zero. We could write the last expression in the following form: min

225

24321 UZPatatPaPataU

We find the partial derivatives of functions on a1, a2, a3, a4, a5 and equate them to zero.

0PZPatatPaPata2a

U

0tZPatatPaPata2a

U

0tPZPatatPaPata2a

U

0PZPatatPaPata2a

U

0tZPatatPaPata2a

U

225

24321

5

225

24321

4

25

24321

3

25

24321

2

25

24321

1

Reducing equations system to the normal form and solving it, we could find the values of the regression coefficients.

Multiple regression equation is as follows:

Z∙103 = 96t+7P+1,5t∙P-5.5t2+0.02P2 (8)

Нづсекt

HPсекt

150;15;0

150;15;1

Figure 6 shows the graphical interpretation of the expression (8) and Z experimental

points. Polished trunnions spinning. Correlation analysis reveals a curvilinear relationship

between the spinning time t and removes oversize Z. Moreover, the maximum correlation coefficient is observed at spinning force P = 120N. For this case the dependence removed allowance Z from duration of spinning was found using the Chebyshev method. In Table 5, Figure 7 and Figure 8 the results of investigations carried out by the above procedure are shown.


18

Fig.6: Theoretical and experimental Z (t, P) dependences during sharpened samples machining

Table 5: The average experimental values of remove oversize Z microns during spinning polished pins Ø4 ... 6 mm

Spinning force P, H

Spinning time t, sec rzt confidence limit at β=0,9 5 10 15

90 0,9 1,2 1,35 0,7 0,34÷0,93 120 1,2 1,35 1,4 0,77 150 1,35 1,5 1,5 300 3,00 3,00 3,00

Z(t) = 1.04+0.02t (9)

оtherNづt

;0

120sec;155;1


19

Fig.7: Comparative data of dependence experimental and calculated values of the spinning removed allowance to the duration of this process

Z(P) = 0.04+0.011P (10)

оthertNづ

;0

sec10;30090;1

The Z (P) dependence is calculated for the case when the maximum force of rollers pressing during spinning does not exceed 300 ... 400N. When P≥500N some deteriorations of trunnions macrogeometry and parts straightness exceeds the allowance for this parameter are observed.

Fig.8: Comparative data of experimental and calculated according to the dependence (10) values of removed allowance Z during spinning with various efforts P.


20

3. CONCLUSION Firstly, researching of finishing methods entering to the assembly parts of small

diameters precision pairs have showed that rollers spinning should be encouraged as a reworking operation. Of all others this method allows to get the items that provide higher performance characteristics using a similar amount of efforts. Secondly, elucidated that spinning finishing is a more controlled process than grinding (correlation coefficients values are larger). Thirdly, the analytical expressions depending value of the removed allowance from the mode finishing were elucidated.

Presented in this work results are valid for steel 14X17H2 and 17X18H9, using other materials more research is needed, which is a further object of research in this area. REFERENCES [1] Bratan, S. M. Technological basis for the quality ensuring and stability improvement of

high-performance finishing and fine polishing / Dissertation for a doctor of sciences degree: Manuscript. − Odessa, 2006. – 321 p.

[2] Novoselov, Yu., Bratan, S., Bogutski, V., Gutsalenko, Yu. Calculation of surface roughness parameters for external cylindrical grinding // Fiability & Durability. – Targu Jiu: Editura “Academica Brancusi”. – No. 1/2013. – PP. 5-15.

[3] Pisarevsky, M. I. Rolling of precision threads, splines and teeth. – Leningrad: Engineering, 1973. – 200 p.

[4] Kremer, N. Sh. Probability theory and Mathematical statistics – Moscow: Unity-Dana, 2002. – 343 p.


21

STRUCTURE AND KINEMATICS OF A MECHANISM WITH STOPS

Prof.PhD. Liliana LUCA, University Constantin Brancusi of Targu-Jiu, [email protected]

Prof.PhD. Iulian POPESCU, University of Craiova, [email protected] Lecturer PhD. Constanța RĂDULESCU, University Constantin Brancusi of Targu-Jiu,

[email protected]

Abstract: It is structurally and kinematical analyzed a mechanism that ensures the switching off of the final driven element, for a subinterval of the cycle in which the crank rotates further. Two variants are analyzed, leading to the conclusion that the first option is more convenient on accuracy. The mechanism serves the purpose, the errors being small. Keywords: mechanism with stops, precision of mechanisms. INTRODUCTION Mechanisms with stops are known for a long time. They have the property that in a subinterval of the cycle of movement of the crank, the movement of the final driven element stops, although leading element is still moving. Cebâşev [1] also built many mechanisms based on geometrical considerations. Other researchers have also designed such mechanisms (Reuleauax, Evans, Silvester, Artobolevskii). The best known mechanism of this type is the mechanism of Geneva (cross of Malta). In [2] it is shown, through animation, how this mechanism works. There are numerous gear mechanisms that meet this condition, based on alternating teeth with arcs on the periphery of the wheels; in [3] it is shown the animation of such a mechanism, which, at one time, from a simple gear, it becomes a planet gear. [4] shows some mechanisms with stops used at textile machines. In [5] there are some mechanisms with bars having breaks in operation. Below, a mechanism from [6] is studied, which can be useful in textile and printing machinery and in other areas. VERSION WITH CONSTANT It went from kinematical diagram of FIG. 1 [Kojevnikov, pg . 468], which states that at the rotation of element 1 in the subinterval given by the angle, section C slides on guide 2, the vertical movement of 2 being stopped (with some approximation).The figure shows that AB is welded to the BC , i.e. AB and BC will tilt simultaneously.


22

Fig. 1 The kinematical scheme of the mechanism is shown in FIG.2. Structurally (fig.3), the mechanism is composed of the leading element with rotating movement AB (R), BCC dyad type RRP and the DDE dyad type RPP, so it is R- RRP – RPP. Fig. 2


23

Fig. 3 For the kinematical analysis of the mechanism, the following relations are written: XB=ABcosφ ; YB=ABsinφ XC=XB+BCcosλ ; YC=YB+BCsinλ=0 δ=λ-180-α XD=XC+CDcosδ ; YD=CDsinδ XF=XD+S4cos(90-γ+180)=const. YF=YD+S4sin(90-γ+180) XE=XF=const. YE=YF+S5=const.

The following initial values were taken: AB=32: BC=43: CD=42: XE=40: YE=64: ALFA=66: GAMA=66: BC=100: AB=15. In Fig. 4 the mechanism is shown in a position. In Fig. 5 the mechanism is shown in successive positions.


24

Fig. 4 Fig. 5

Fig. 6 notes that all races: XC, S4 and S5 have symmetrical variations. It also notes that the curve of S5 has the smallest amplitudes. Further, it has been made the crank angle variation as in Fig. 1, and it is determined that = - 45…+45, so that the successive positions have been obtained in Fig. 7. It is noticed that the FD segment positions are almost overlapping, which means that the variation of the S5 race on this subinterval is very small, that is exactly what is required in this mechanism.

0.0 100. 200. 300. 400.

Fi

20.

40.

60.

80.

100.

120.

X CS4S5

Fig. 6 Fig. 7

In Fig. 8, the symmetry of the curve is seen and large variations. In reality, the ordinate data show small variations, up to 1.95 mm, which is considered admissible; there are low variations on the ordinate.


25

-60.00 -40.00 -20.00 0.0 20. 40. 60.

S5

49.

49.5

50.

50.5

51.

51.5

Fig. 8 The values obtained are given in Table 1.

Table 1 Fi S5 -45 49.46939 -40 49.86157 -35 50.21485 -30 50.5266 -25 50.79444 -20 51.01639 -15 51.19076 -10 51.31625 -5 51.39193 0 51.41722 5 51.39192 10 51.31624 15 51.19074 20 51.01636 25 50.79441 30 50.52655 35 50.2148 40 49.86151 45 49.46932


26

VERSION WITH CD WELDED TO SLIDE 3 It was studied variant where the mechanism of fig. 2 is amended so that the CD is not welded to BC, but the slide 3. In this case the structure of the mechanism is the one of Fig. 9, the mechanism being R- RRP - RPP type, but the RPP dyad changes. In this case, the successive positions of the mechanism are given in Fig. 10, ascertaining that, in contrast to Fig. 5, the CD tray always stays parallel to the ordinate. Fig. 9 In Fig. 11 are given only the positions for = - 45…+45, observing variations of the S5 race. In Fig. 12 the S5diagram ( ) is given, noticing the higher values of S5 for this subinterval, in comparison with the diagram of Fig. 8.

Fig. 10 Fig. 11


27

-60.00 -40.00 -20.00 0.0 20. 40. 60.

Fi

53.

53.5

54.

54.5

55.

55.5

S5

Fig. 12 Numeric values are given in Table 2. Here, the maximum difference is 2.21, as opposed to the previous case when it was 1.95, therefore the first mechanism is more convenient.

Table 2 Fi S5 -45 53.1847 -40 53.62203 -35 54.01904 -30 54.37177 -25 54.67664 -20 54.93051 -15 55.13078 -10 55.27533 -5 55.36268 0 55.3919 5 55.36268 10 55.27533 15 55.13078 20 54.93051 25 54.67664 30 54.37177 35 54.01904 40 53.62203 45 53.1847


28

Referring to FIG. 13, it is noted that the solution to the results in Table 1 (first mechanism) provides lower values for S5 compared to the second mechanism, with the results given in Table 2.

0.0 100. 200. 300. 400.

Fi

35.

40.

45.

50.

55.

T abelul 1T abelul 2

Fig. 13 CONCLUSION - The structure of the mechanism from [6] was studied in two versions. - S5’s race variation was established for a particular subinterval of the crank rotation cycle for both structural variants. - It appears that the more precise is the first structural variant. - The mechanism serves the purpose, the errors being small. REFERENCES [1]. Cebâşev, P. L. – Izobrannâe trud, Izd. Nauka, Moskva, 1953. [2] . http://en.wikipedia.org/wiki/Geneva_drive [3]. http://www.mekanizmalar.com/beaver_tail.html [4]. Choogin, V., Bandara, P., Chepelyuk, E. – Mechanisms of flat weaving technology. The Textile Institute, Woodhead Publishing, 1913. [5]. http://blog.rectorsquid.com/linkage-mechanism-designer-and-simulator/ [6]. Kojevnikov, S. N., Esipenko, Ia.,I, Raskin. Ia., M. – Mehanizmî. Spravocinoe posobie. Izd. Maşinostroenie, Moskva, 1976.

http://en.wikipedia.org/wiki/Geneva_drive


29

OPTIMIZATION OF A MECHANISM WITH STOPS

Prof.PhD. Iulian POPESCU, University of Craiova, [email protected] Prof.PhD. Liliana LUCA, University Constantin Brancusi of Targu-Jiu,

[email protected]

Abstract: The optimization of the two angles of a mechanism with stops is done by repeated analysis sequences. We study every angle influence on the precision with which it is ensured the switching off of the final driven element on a subinterval on which the leading element rotates. It reaches optimal solutions by analyzing associated positions and diagrams of the races. Keywords: optimization of mechanisms with bars, precision of the mechanisms, mechanism with stops. INTRODUCTION Optimizing mechanisms was addressed in more detail since the 70s when computers were introduced in the calculation mechanisms. Earlier attempts are made by graphic optimization mechanisms and especially by designer intuition. Computers have allowed the development of algorithms of increasing performance optimization. The difficulty in optimizing the mechanisms is that the mechanism is a system with variable geometry, which means that optimization is not static, but a dynamic one, which complicates the algorithms and programs. In [1] stepping mechanism was studied, using optimization techniques. In [2] is done the optimal synthesis of the articulated quadrilateral through linear programming, for trajectories imposed. In [3] the Watt's mechanism is optimized for drawing a straight line by a connecting rod point, using genetic algorithms. In [4] the optimization of several mechanisms with bars is studied for trajectories or required associated positions, using different algorithms from numerical analysis. Below, a mechanism from [5] is optimized to obtain maximum accuracy. INITIAL DATA It went from the kinematical diagram of FIG. 1 of a mechanism kinematical and structurally studied in a paper aside. The following initial values were taken: AB=32:BC=43:CD=42:XE=40:YE=64: ALFA=66: GAMA=66: BC=100:AB=15.


30

Fig.1 The purpose of the mechanism is to stop the movement of element 5 for a subinterval where the crank AB rotates further, so the mechanism is with stops. Optimizing aims to establish values for angles and for which, in the subinterval = - 45 ... + 45 degrees, S5 to be minimal, to zero. RESULTS The optimization was done by a succession of repeated analysis sequences. The solutions have been verified by tracking the movement of segment FD, for the said sub domain, the good solutions being those where FD's positions are almost superimposed. First, successive positions of the movement of the mechanism for the entire cycle were drawn, as follows: fig. 2 ( =25), fig. 3 ( =40), fig. 4 ( =60), fig. 5 ( =100), fig. 6 ( =120). We observe the change in successive positions of the mechanism and S5 becomes negative for some values of . The angle positions the straight line FD to the vertical line FE. Referring to Fig. 2, it follows that S5 < 0.

Fig. 2 Fig. 3


31

Referring to Fig. 3, it follows that S5 has both positive and negative values.

Fig. 4 Fig. 5

Fig. 6 Referring to Fig. 4, 5, 6, it appears that S5 > 0. Referring to Fig. 2…5, it is observed that at the increase of , the S5's variation decreases and at the increase of over 100 degrees, S5 starts to increase. Further, angle has been cycled, resulting in: fig. 7 ( =10), fig. 8 ( =20), fig. 9 ( =40), fig. 10 ( =60), fig. 11 ( =90), fig. 12 ( =120). For the case of Fig. 7, S5 has both positive and negative values.

Fig. 7 Fig. 8


32

Fig. 9 Fig. 10

Fig. 11 Fig. 12 It is noted that only for = 20 ... 40, the variation of S5 decreases, otherwise it increases. Further, it took only successive positions for 45...45 , resulting, at the cycling of : fig. 13 ( =20), fig. 14 ( =40), fig. 15 ( =60), fig. 16 ( =70), fig. 17 ( =80), fig. 18 ( =90).

Fig. 13 Fig. 14

Fig. 15 Fig. 16


33

Fig. 17 Fig. 18 Referring to Fig. 13…18, it is observed that the most convenient are the values for = 60 ... 80 degrees. Then, was cycled, resulting: fig. 19 ( =35), fig. 20 ( =45), fig. 21 ( =60), fig. 22 ( =65), fig. 23 ( =70), fig. 24 ( =80).

Fig. 19 Fig. 20

Fig. 21 Fig. 22


34

Fig. 23 Fig. 24 The solutions with =60…80 and =65…70 are convenient. As a result of these findings, the values of and have been changed at the same time and, in suitable areas, resulting solutions close to the optimal solution: fig. 25 ( =60, =60), fig. 26 ( =65, =70), fig. 27 ( =70, =70), fig. 28 ( =70, =80).

Fig. 25 Fig. 26

Fig. 27 Fig. 28 New dimensions more convenient for the variation of S5 race were found. Referring to FIG. 29, corresponding to one complete revolution of the crank AB, with values 80;70 , the following is observed: - The curve of S4 is symmetrical; when increases from 0-180 degrees, D goes left, then right; - The variation of S5 is low, and in the ranges = 0 ... 45 and = 0 ... ( - 45 ), the curve is flattened.


35

0.0 100. 200. 300. 400.

Fi [ deg]

20.

30.

40.

50.

60.

70.

80.

S4 [mm ]S5 [mm ]

Fig. 29 From Table 1 and Fig. 30 of S5 is found that variations of S5 are very small, the curve is close to a straight line.

Table 1 Fi S4 S5 -45 71.12304 34.35019 -40 72.12045 34.52339 -35 73.02592 34.68062 -30 73.8304 34.82031 -25 74.52571 34.94105 -20 75.10471 35.04159 -15 75.56146 35.1209 -10 75.89115 35.17815 -5 76.09037 35.21275 0 76.15701 35.22432 5 76.09037 35.21275 10 75.89115 35.17815 15 75.56146 35.1209 20 75.10471 35.04159 25 74.52571 34.94105 30 73.8304 34.82031 35 73.02592 34.68062 40 72.12045 34.52339 45 71.12304 34.35019


36

-60.00 -40.00 -20.00 0.0 20. 40. 60.

Fi [ deg]

30.

40.

50.

60.

70.

80.

S4 [mm ]S5 [mm ]

Fig. 30 CONCLUSION

- A mechanism with stops was studied. - Its optimization was done through sequences of repeated analysis. - - We studied the influence of the variations of two constant angles on race S5, which

minimized. - - We established the optimum values of these angles, checking the operation of the

mechanism with the utmost precision. References [2].Marjanovic, D. , Four-bar linkage design using global optimization. În: Proceedings of Design, 2002, the 7th International Design Conference, Dubrovnik. [1].Ghassaei, A., Choi, P., Whitaker, D. – The design and optimization crank-based leg mechanism. Pomona College Dep. Phisics and Astronomy, 2, 2011 [http://www.amandaghassaei.com/files/thesis.pdf]. [2].Marjanovic, D. ,Four-bar linkage design using global optimization. În: Proceedings of Design, 2002, the 7th International Design Conference, Dubrovnik. [3].Mehdigholi H., Akbarnejad, S. , Optimization of Watt`s six-bar linkage tot generate straight and parallel leg motion. Sharif University of Technology, IN-Teh [http://cdn.intechopen.com/pdfs-wm/4294.pdf] [4].Popescu. I., Mîlcomete, D. C. , Cercetări privind sinteza şi optimizarea mecanismelor. Editura Sitech Craiova, 2006. [5].Kojevnikov, S. N., Esipenko, Ia.,I, Raskin. Ia., M. ,Mehanizmî. Spravocinoe posobie. Izd. Maşinostroenie, Moskva, 1976.


37

ANALYSIS OF THE MECHANICAL STRENGTH OF A DRIVING MECHANISM CALLED SHOCK

Prof. phd. eng. Dan ILINCIOIU University of Craiova, Faculty of Mechanics,Department of Applied Mechanics and Civil

Constructions, Calea Bucuresti Street, no.107, Craiova,Code 200512, Romania, [email protected]

Eng. Cosmin MUSCALAGIU University of Craiova, Faculty of Mechanics, Calea Bucuresti Street, no. 107,Craiova,Code

200512, Romania, [email protected] Assistant Phd. Eng. Cosmin-Mihai MIRIŢOIU

University of Craiova, Faculty ofMechanics, Department of Vehicles, Transports and Industrial Engineering, CaleaBucuresti Street, no. 107, Craiova,Code 200512, Romania,

[email protected]

Abstract. It evaluates the maximum static and dynamic stresses produced in the elements of a quadrilateral mechanism transporting a vehicle in the storage in an urban park. Determine multiplier shock hazard if the mechanism freezes and increases mechanical stress.

Keywords. Parking the car, actuators, shock requests, optimizations.

1. Defining the problem

Fig. 1. Car platform





38

It examines platform transport a car in the modular storage in a multi-storey car park. The system is essentially a two beams oscillating platform articulated according to data in figure 1. Scheme for calculating the strength of the structure is given in Figure 2. Kinematic parameters required are:

T

5,0, rv

notations: - speed swing angle bars;T- time lowering the car; r- oscillatory length;

Fig.2. Calculus scheme

The masses in motion are:

car table: M=g

F

g

FF 21 ; platform table: m=m0l; oscillating bar table: m1= rm 0 ;

Ratings: F-weight car; g-gravitational acceleration , m0-specific mass of the beams. It will define only the reactions required for resistance calculation (calculation is not complete).

2. Static analysis In terms of balancing the moments on the platform, the reactions are (neglecting the masses platform).

V1= )()(1

21 blFalFl

, F1= Fk , F2=(1-k)F

V2= bFaFl 21

1 (1)

It will take into account only the request in bending. Platform moment equations are: M12= xV 1 ; M23= )(11 axFxV ; M34= )()( 211 bxFaxFxV (2)


39

If either of the oscillating bars are locked in lower joint, the bar will be requested blocked and bending moment was: M1= xV sin1 sau M2= xV sin2 (3) It defining the mechanism bars: the moment of inertia Ix, modulus Wx, m0-specific mass. The masses bars are: m= 0ml , m1= 0mr (4)

Tensions are likely maximum:

xx W

aV

W

M 122 ;

xx W

abFbV

W

M )(1133

; xW

rV sin11 ;

xW

rV sin24 ; (5)

3. Shock calculation It requires a locking mechanism during collaboration, causing an increase in the mechanical stresses. Impact multiplier is calculated using the equation:

K 11 , Kpc

c

E

E ; (6)

notations are: Ec-kinetic energy of the moving masses, Ep- the potential energy stored in mechanism elements. Define 3 cases blocking the movement: 1.platform blocking horizontal direction; 2.rocker blocking A-1 in the lower joint; 3.rocker blocking B-4 in the lower joint; It will calculate the multiplier in 3 cases.

3.1.Blocking Case 1.

The kinetic energy is:

21 )(5,0 vm

g

FEc (7)

Potential energy due to bending deformation is:

a b

a

l

bxxp dxMdxMdxM

IEdxM

IEE

0

234

223

212

21 2

1

2

1 (8)

Replacing the timing relationship and make the integrals, parenthesis in equation (8) is calculated with equation (9).

klbabkblakbalbakbbkblakl

FI )()()2()()(

3232

2

2

1 +

+ 232

2

)()(3

bkbakbll

F (9)


40

Potential energy is:

xp IE

IE

21

1 (10)

E-modulus of elasticity of the material

11 11 K , K1=1

1

p

c

E

E (11)

Maximum dynamic tensions are:

122 d , 133 d (12)

3.2.Blocking Case2

The kinetic energy is:

Ec2=2

125,05,0 vmmg

F

(13)

Potential energy stored in the elements of the mechanism is:

Ep2= )(2

1

2 0

211

2 v

Apxx

dxMEIEIE

I (14)

After performing the integral bracket of equation (14) is calculated using the equation:

I2= 2

2

3 l

F

{(ka-l+b-kb)2(b3+v3sin2)+k(a-b)2[l(ka-l+b-kb)(a+2b)-kl2(a-b)]+(l-b)3(ka+b-kb)2(15)

Dynamic tensions are:

,11 22 K K2=2

2

p

c

E

E; 222 d ; 323 d ; 121 d ; (16)

3.3.Blocking Case 3

The kinetic energy is similar to the previous situation: Ec3=Ec2

Potential energy is:

Ep3= )(2

1

2 0

221

3 r

pxx

dxMEIEIE

I (17)

Parenthesis in equation (17) is calculated using the equation:

I3= 2

2

3 l

F

{(ka-l+b-kb)2b3+k(a-b)2[l(ka-l+b-kb)(a+2b)-kl2(a-b)]+(ka+b-kb)2[(l-b)3+r2sin2] (18)

Dynamic stresses are calculated:

33 11 K , K3=3

3

p

c

E

E (19)

232 d , 333 d , 434 d


41

3.4.Evaluation of a particular case We adopt dimensions: l = 400cm, a = 80cm, 300cm b = r = 180cm (20) Full displacement (vertical-horizontal) will be in T = 30s, resulting kinematic parameters:

= sradT

/052,05,0

, v=r=9,42cm/s (21)

Rolling mechanism is adopted to build IPE200 type that has the following characteristics: Ix=1940cm4, Wx=194cm3, m0=0,224 kg/cm (22) The position of the center of gravity of the vehicle and its weight are: F=2000daN, K=0,4, g=1000cm/s2, F1=KF=800daN, F2(1-K)F=1200daN (23) Maximum voltages are (it is thought that = 0,5): 1=872daN/cm2, 2=387daN/cm2, 3=546daN/cm2, 4=983daN/cm2 (24) Dynamic stresses calculated with shock multipliers are given in Table 1. Table 1

Cazul 1 2 3

4,9 4,17 4

1d(daN/cm2) - 3637 -

2d 1895 1616 1558

3d 2671 2278 2196

4d - - 3953

4. Conclusions

Defined mathematical model for assessing static voltages and multipliers shock for threecases blocking transport mechanism. Under quasi-static stress mechanical stresses are relatively small, lower value (considerated admissible) of 100daN /cm2 (100MPa). If it is assumed that the mechanism is made of steel S235 (blood flow c = 235Mpa = 2350daN /cm2), static requirements are much lower blood flow. It compares the 3 cases and found that the rod-platform is the most disadvantaged in case 1, the multiplier having the maximum value, the maximum voltage exceeds the yield, so it is necessary to increase the time of descent from T = 30s to 50 or 60 seconds (blood decreases maximum 1890 daN /cm2). In case 2 and 3 locking platform is less dynamic query request while remaining high (blood flow is reached).In case 2 the request is slightly larger than in case 3. Comparing requests oscillating bars are found in the worst case case 3, which recommends preferable drive lock joint articulation A. B rupture respective oscillating bar; if it doubles during descent (T = 60s)


42

swing over the yield required, so it must necessarily bar to rotate freely. It is possible that in other specific conditions (size, strength, speed, material, etc) comparative analysis of requests and cases lead to other conclusions. The mathematical model being defined parameters, numerical determinations allow complete and easy.

Acknowledgement This work was supported by the strategic grant POSDRU/159/1.5/S/133255, Project

ID 133255 (2014), co-financed by the European Social Fund within the Sectorial Operational Program Human Resources Development 2007-2013.

References [1].Ilincioiu, D, Miritoiu, C, Strength of materials,vol.II,Ed. SITECH, Craiova 2012. [2].Tripa, P, Strength of materials, Ed. MIRTON, Timisoara 2001. [3].Muscalagiu, C, Ilincioiu, D, Report 3 doctorate, Craiova 2015. [4].Galaftion, S, Pascu, A, Strength of materials, Ed Univ. ,,Lucian Blaga”, Sibiu 2007. [5].Dumitru, I, Prticularities of calculation to requests by shock, Ed. Timisoara 2007.


43

THE STUDY ON INCREASING THE RELIABILITY OF WHEELS AT RAIL VEHICLES WITH DIESEL-ELECTRIC TRACTION

drd. eng. Bianca Silvia ORHEI, Facultatea de Constructii de Masini si Management Industrial Iasi, [email protected],

dr. eng. Daniel APOSTOL [email protected]

Abstract: We have presented a study regarding reliability growth of driving wheels at railway vehicles with diesel-electric traction. The study has been done on 060 DA diesel-electric locomotive and it was possible thanks to existing modernization, made in repairing plants at mechanic and electric equipments of the locomotive. Keywords: diesel-electric locomotive, defects, profile, reliability, modernization

1.Introduction

At present, Romania has a railroad system with a total length of 10818 km, of which 6816 km is unelectrified and 4002 km is electrified. The most important railway operators which activate on this railway have in their own rolling-stock depot railway vehicles as locomotives or railcars. Locomotives are in a greater number than railcars, such as electric and diesel-electric locomotives.

In Romania, the tow of passenger trains and goods trains is made, mostly, using two types of locomotives : 060 DA diesel-electric-used on unelectrified railway and 060 EA electric locomotive, used exclusively on electrified railway. These locomotives are equipped with continuous power electric engines with serial excitation, which are electrically connected series-parallel or parallel only.

The study intends to point out the influence of the type of electric connection at traction engines on the wear of driving wheels of railway vehicles with continuous power electric traction. Thus, we have followed and analysed the appearance of profile defects of the driving wheels in operation of 060 DA diesel-electric locomotives partially modernized and totally modernized.

These types of diesel-electric locomotives were designed by Electroputere Craiova Plant between 1965-1985, after 2000 were modernized, in partial version (MP-partial modernization), without the replacement of group diesel engine-generator and in total version (MT-total modernization) with thw replacement of group diesel engine-generator.

General technical features for the diesel-electric locomotives partially and totally modernized :

- axle formula : Co-Co - gauge : 1435 mm - length between buffers sides : 17 000 mm - maximum width : 3000 mm - wheelbase bogie : 4100 mm - total wheelbase : 12400 mm


44

- the distance between bogie pivots : 9000 mm - wheels diameter with new bandages : 1100 mm - wheels diameter with half-worn bandages : 1060 mm - top speed : - in current line : 100 km/h for 060 DA

- in current line : 120 km/h for 060 DA1 - diesel engine : type Sulzer for 060 DA MP type GM for 060 DA MT 2.Generalities regarding profile defects appearing at the wheels of traction vehicles

The profile defect could be defined as a deviation from the parametersof wheel profile (wear) admitted in operation of traction railway vehicles. The deviations appear due to wear of driving wheels and it is influenced by many factors, but the most important is slip.

Profile defects can appear due to normal wear of the wheel (bandage) at the wheel-rail contact or accidental due to external factors as : improper braking, material defects, slip,etc.

The repair of the driving wheel profile consists in reshaping (readjustment), this operation being made on underground lathes. Any repair of the profile implies the removal of metallic material and thus-reducing the wheel diameter. The number of the reshapings is limited, so is intended as these should be as few.

The profiles after that the wheels of traction vehicles in Romania are turned respect STAS 112/3/1990.(fig.1).

Fig. 1 - Wheelset and external profile of the wheel rims(bandages) processed for locomotives, in accordance to STAS 112/3/1990

The main odds which are measured at a bandage profile and which decide if we have a

profile defect due to a normal wear are : - radial tire wear on rolling circle; - thick bandage measured on rolling circle; - flange thickness; - the distance between internal sides of driving wheel bandages; - the distance between external sides measured at 10 mm above the rolling circle. Profile defects which are determined by abnormal wear of the tire (accidental) due to the

following : flat appearances (improper brakings, slips), various traces on tire surface (gouges, channels, deformities after impact with hard objects), material peels (material defect).


45

3.The analysis of profile defects at a traction railway vehicle with diesel-electric traction-060 DA diesel-electric locomotive

The study regarding growth of reliability of driving wheels by optimal electrical connection of continuous current of electric engines with serial excitation was made on traction railway vehicles -060 DA diesel-electric locomotives, partially modernized(MP) and totally modernized (MT), between 2005-2009.

We have chosen this locomotive of the following reasons : - after locomotive modernization, the whole bogie remained the same; - wiring diagram connection of traction motors differs at the two versions of

locomotives; - the locomotives have modern equipments from wheel –slide protection - the locomotives towed trains on the same sectors tow.

We kept in mind the following aspects :

a) At the 060 DA partially modernized locomotive : - connection of traction motors is series-parallel type (fig.2) - the wheel tread profile –in accordance to STAS 112/3/190 - the reshapings were made on underground lathe-Hegenscheidt type 106 - we noticed 060 DA MP-060 DA partially modernized locomotive-with top speed 100 km/h and with 060 DA1 MP-060 DA1-partially modernized locomotive with top speed 120 km/h

Fig. 2 - Connection of traction motors at 060 DA partially modernized locomotive

We can write : U = E + I ∑R = ke n Φ + I ∑R [ V ] ( 1 )

Considering the transmission ratio i = zmotor / zroata = 69 / 15 (used at 060 DA locomotive with top speed 100 km/h)

nmotor = 24,41 vlocomotive / dwheel [ rot / min ] ( 2 )

Neglecting internal resistances of electric motors

U = E1 + E4 = E2 + E5 = E3 + E6 [ V ] ( 3 )


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U = 20,91 vlocomotive [( ke1 Φmotor1 / dwhee1 ) +[( ke1 Φmotor4 / dwheel4 ) [ V ] ( 4 )

b) At the 060 DA totally modernized locomotive : - connection of traction motors is parallel type (fig.3) - the wheel tread profile- in accordance to STAS 112/3/190 - the reshapings were made on underground lathe-Hegenscheidt type 106 - we noticed 060 DA MT-060 DA totally modernized locomotive with top speed 100

km/h and 060 DA1 totally modernized locomotive with top speed 120 km/h -

Fig. 3 - Connection of traction motors at totally modernized locomotive with

GM-USA equipment Now, we can write:

U = E1 +I1 ∑R1= E2+I2 ∑R2 = E3 + I3 ∑R3= E4 +I4 ∑R4= E5+I5 ∑R5 = E6 + I6 ∑R6 [V] (5) Neglecting internal resistances of electric motors

U = E1 = E2 = E3 = E4= E5 = E6 = ke nmotor Φmotor [ V ] ( 6 )

U = 20,91 ke Φmotor vlocomotive / dwheel [ V ] ( 7 )

The 060 DA partially modernized locomotive is using a connection series-parallel type,

and 060 DA modernized locomotive with General Motors equipment is using a connection parallel type for the traction motors. Thus, we have observed :

- in the case of 060 DA partially modernized locomotive, due to series parallel connection, electrical current that circulates by two traction motors will be always the same, so any increase of revolution of one motor (slip) will lead to the decrease of revolution of the other motor;

- at the 060 DA modernized with General Motors equipment, traction motors work independently one another, the increase of revolution of one motor doesn’t influence the others motors;

- from the equations (4) and (7) we observed that the wheel diameter (dwheel) intervenes in tension equation, thus influencing motor operation. In vehicle operation, the diameter difference between wheels is limited ;


47

- if we consider that traction electric cars aren’t equally perfect as parameters and that the driving wheels diameters are not equal, than all traction engines will have different revolutions, so wiring diagram connection in parallel should be better;

- in tabel 1 and 2, and fig. 3 and 4 are represented the number of kilometers between two profile defects at 060 DA partially and total modernized locomotive.

Tabel 1 - Mileage between two profile defects at 060 DA modernized locomotive with increased speed (120

km/h) Locomotive Mileage between two profile defects

060 DA1 MP 112884 156918 332000 405300 510000 658000 726000 853000 060 DA1 MT 49099 403800 541000 610000 896000 - - -

Fig. 4 - Wiring diagram connection of traction motors at totally modernized locomotive with GM-USA

equipment

Tabel 2 - Mileage between two wheel profile defects at 060 DA modernized locomotive with top speed 100 km/h

Locomotive Mileage between two profile defects 060 DA MP 98000 230420 447009 519000 708000 770000 787000 - 060 DA MT 48980 393400 588000 688000 770000 879000 930000 1160000


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Fig. 5 - Mileage between two wheel profile defects at 060 DA modernized locomotive with top speed 100

km/h 5.Conclusions ▪ At present the 060 DA diesel-electric locomotive suffered profound changes, becoming a modern locomotive, equipped with the newest technologies. ▪ Profile defects of the driving wheel are more numerous at 060 DA locomotive and 060 DA1 than at 060 EGM diesel-electric locomotive and 060 EGM1, thus meaning repair of wheel profile more often at the same mileage. This fact makes us believe that also electric scheme of connection of traction engines has an influence in increase/decrease of driving wheel wear. ▪ The modernization of a locomotive is more efficient considering that at the electric equipment design is important their influence on mechanical equipments. 6.References

[1] Al. Popa, Locomotives and railcars with heat engines, R.A. Bucuresti – 1978 [2] Dan Bonta, Diesel-electric locomotive 060 – DA – 2100 CP, ASAM, Bucuresti –

2003 [3] Dumitru Mihailescu, Locomotives and electric trains asynchronous traction motors,

R.A. Bucuresti – 1997 [4] Lorin Cantemir, Mircea Oprisor, Electric traction, Bucuresti – 1971 [5] V. Hoanca, L. S. Bocii, Railway vehicles with thermal engines, ASAB, Bucuresti –

2002 [6] Benn Coifman, The Evolution of the Diesel Locomotive in the United States, 1994; [7] STAS 112/3/1990 [8] www.cfr.ro [9] www.remarul.eu [10] www.electroputere.ro

http://www.cfr.ro/

http://www.remaru/

http://www.electroputere.ro/


49

OPTIMAL KINEMATIC DESIGN OF A CAR AXLE GUIDING MECHANISM IN MBS SOFTWARE ENVIRONMENT

Dr. eng. Cătălin ALEXANDRU

Transilvania University of Braşov, [email protected]

Abstract: This work deals with the optimal kinematic design of the guiding mechanism used for the rear axle of a motor vehicle. A guiding mechanism of type 2S1C (S - sphere, C - circle) is taken into consideration, the optimization study being conducted by using the multi-body systems (MBS) software environment ADAMS of MSC. The global cartesian coordinates of some design points (the locations of the joints between the guiding arms and car body) are used as independent design variables for the optimization study. The main design objectives refer to the longitudinal displacement of the axle and its proper rotation (around the transversal axis), the kinematic optimization goal being to minimize these variations. Keywords: axle guiding mechanism, kinematics, multi-body system, optimal design.

1. Introduction

For the guidance of the rear axle of the motor vehicles, two solutions are used: independent guidance of the wheels, case in which each wheel is guided by its own linkage, and dependent guidance of the wheels, case in which the rear axle is guided relative to car body [1]. The dependent guidance of the rear axle, which is frequently used for the off-road and commercial vehicles, is assured by spatial linkage mechanisms, on which between axle and car body a number of binary links or kinematic chains are interposed. The links connections to axle and car body are made through compliant joints (bushings) with 6 elastic restricted degrees of freedom. Usually, for the kinematics of the axle guiding linkages, the bushings are modeled as spherical joints, neglecting the linear deformations. At the same time, the car body is attached to ground. In this way, the axle guiding linkages have low degree of mobility, m=1 or m=2.

The guidance of the axle is made by driving a number of its points on suitably chosen surfaces and curves (fig. 1): sphere (s), circle (c), and coupler curve (cc). The guidance on sphere (fig. 1,a), respectively circle (fig. 1,b), is achieved by a binary link, interposed between axle and car body, with spherical joints in both ends, respectively a rotational joint to car body and a spherical joint to axle. The guidance on coupler curve (fig. 1,c) is performed by a spherical joint between axle and coupler; in this case, watt mechanism configuration is frequently used, but roberts, chebyshev or evans straight-line linkages can also be used.

Joining in parallel the basic chains, various axle guiding linkages with m=1 and m=2 can be obtained. The structural systematization of the guiding linkages was presented in [3], taking into account the type of joints (spherical and/or revolute), the type of kinematic chains, and the number of kinematic chains connected in parallel. The axle guiding linkage by three points (so called 2s1c) is obtained by joining three binary links of type 2[a] + 1[b], in compliance with figure 1. The mechanism contains two lower rods (1s, 1d) and one upper triangular arm (3), which are longitudinal disposed (fig. 2). The connections of the upper arm



50

to car body can be disposed in front (fig. 2,a) or in rear (fig. 2,b) relative to axle. The two spherical joints between the upper arm and car body determine, in fact, a revolute joint.

Fig. 1. The basic types of guidance of the axle points.

a. b.

Fig. 2. The axle guiding mechanism by three points (2s1c).

Relative to car body, the rear axle must have the possibility of vertical motion and rotation around the longitudinal axis of the car. The modification of the vertical position of the wheels (when the vehicle passes over bumps) determines, besides the above-described necessary motions of the rear axle, secondary undesirable motions, as follows: displacements of the axle center P along the longitudinal (XP) and transversal (YP)

directions; rotations of the axle around the vertical (Z) and transversal axes (proper rotation, Y).

The minimization of the undesirable motions can be transposed into kinematic optimization criteria, as follows: XP 0, YP 0, Z 0, Y 0. These criteria cannot be equally satisfied, and for this reason a certain criterion has priority, or, usually, a compromise is accepted, such as: XP [XP min, XP max], YP [YP min, YP max], Z [Z min, Z max], Y [Y min, Y max]. The boundaries can be established depending on the top speed of the car, type of tires, type of car, and other criteria.


51

2. The kinematic model of the axle guiding linkage ADAMS of MSC.Software is a powerful modeling and simulation environment that

provide build, simulate and refine models of mechanical system. In ADAMS, the kinematic model of a mechanical system is characterized as a constrained, multi-body system, in which the parts are connected through geometric and kinematic constraints (joints and motion generators) [4, 5].

The mechanism is modeled and analyzed in a global coordinate system (GCS), which it is an inertial system. The ground body acts as the global coordinate system that defines the global origin and axes about which the model is created. For any mobile body (guiding arms, axle), a local coordinate system (LCS) is assigned, which moves with the body and its original position defaults to that of the global coordinate system. It must be mentioned that - for the kinematic model - the car body is attached to ground.

The 2S1C guiding mechanism contains 4 mobile parts (lower guiding arms - 1s/d, upper arm - 3, axle - 2, see the notations in figure 2) that are connected through 6 geometric constraints, and 2 kinematic constraints that control the vertical displacements of the wheels. For modeling the vertical motion of the left/right wheels, there are used motion generators, which dictate the part (wheel/axle) motion as a function of time. Having in view that the independent kinematic parameters are the vertical positions of the centers of the left/right wheels, the motion generators have been modeled as single point motions (fig. 3), which prescribe the motion of the mobile parts (left/right wheels) relative to the ground body along the vertical axis (Z). Each motion generator removes one degree of freedom (DOF).

Fig. 3. The motion generators that drive the kinematic model.

The total number of degrees of freedom, which represents the number of

undetermined motions (generalized coordinates), is equal to the difference between the number of allowed part motions and the number of constraints (Gruebler count), DOF = 6n - r, as follows:

generalized coordinates for 4 mobile bodies: 4 6 = 24; degrees of freedom restricted by the geometric constraints (joints):

- spherical joints between axle and lower/upper arms (Ms, Md, N): 3 (-3) = -9,


52

- spherical joints between car body (ground) and lower arms (M0s, M0d): 2 (-3) = -6, - rotational joint between car body and upper control arm (N0): 1 (-5) = -5,

degrees of freedom controlled by the kinematic constraints: - motion generators (point motions): 2 (-1) = -2,

resulting DOF = 24 - 22 = 2, meaning the two passive proper rotations of the lower guiding arms (namely 1s and 1d - see fig. 2). These passive degrees of freedom can be eliminated by applying motion generators in one of the corresponding spherical joints that are used to connect the guiding arms (for example, in Ms and Md - see fig. 2), obtaining in this way DOF = 0; therefore the model is kinematically determined.

Given the aim of the paper (the kinematic optimization of the axle guiding linkage), the mechanism has been modeled by using parameterization tools. Parameterizing the mechanism simplifies changes to model because it helps to automatically size, relocate and orient the design objects (e.g. bodies, joints). In this way, relationships into the model are created, so that when a modeling object is changed, ADAMS updates any other objects that depend on it.

For the guiding linkage in study, the design points are the easiest way to parameterize the model. Using design points, important locations can be specified in the model, and then other objects (e.g. joints) can be attached to these points. According to figure 4, in the 2S1C guiding mechanism there are the following design points: M0s, M0d - the locations of the joints between the lower guiding arms and car body; N0s, N0d - the points that define the axis of the revolute joint between the upper control arm and car body; Ms, Md, N - the locations of the joints between the lower & upper guiding arms and axle; Gs, Gd - the centers of the wheels.

Fig. 4. The virtual model of the 2S1C guiding linkage.

The geometries of the bodies have been attached to these points, as follows: the axle -

between Gs and Gd; the left lower arm - M0s and MS; the right lower arm - M0d and Md; the upper control arm - N0s and N, respectively N0d and N. The left/right wheels are fixed connected / attached to axle.


53

3. Results and conclusions To optimize a mechanical system, the following steps are necessary [2]: defining the

design variables, defining the design objectives for optimization, performing design studies, optimizing the model on the basis of the main design variables. Design variables allow creating independent parameters and tie modeling objects to them. Design study describes the ability to select a design variable, sweep that variable through a range of values and then simulate the motion behavior of the various designs in order to understand the sensitivity of the overall system to these design variations. Before running the design study, the range (list) of values for each design variable must be specified. Design optimization represents the capability to define design objectives, constraints and variables, and then have the software iterate automatically to the optimally - performing configuration.

In the first step, the coordinates of the points that parameterize the model (shown in fig. 4) have been transformed in design variables (DV): XMs DV_1, YMs DV_2, ZMs DV_3 and so on. Having in view the symmetry of the mechanism relative to the longitudinal axis of the car, between the points that define the topological scheme of the mechanism the following expressions have been modeled: XM0d = XM0s, YM0d = -YM0s, ZM0d = ZM0s, XN0d = XN0s, YN0d = -YN0s, ZN0d = ZN0s, XMd = XMs, YMd = -YMs, ZMd = ZMs, YN = 0. When an expression is created, ADAMS stores the expression and updates the value whenever a value in the expression changes. In these terms, the design variables that control the axle guiding linkage during the optimization are presented in figure 5.

It must be mentioned that for this paper only the coordinates of the joints on car body (M0s/d, N0s/d) have been chosen as design variables for the optimization (in other words, DV_6 DV_11 are the selected design variables) the locations of the joints on axle (Ms/d, N) remaining established by constructive criteria.

Fig. 5. The design variables in the optimization process.


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The design functions for optimization are represented by measures that define the kinematic behavior of the guiding mechanism (see section 1): XP, YP, Z and Y. The objective is to minimize these motion variations of the axle. With this end in view, for each design variable and objective function, design studies were performed to determine the design variables that have great influence on the kinematic optimization criteria. Thus, a list of values was specified for each design variable, the simulation of the mechanism being performed for all values of the design variable. Finally, the design study report offers information about the sensibility of the objective function on the variations of a certain design variable. The report allows selecting the main variables that will be used for optimization.

On the basis of the design studies done for the above-described guiding linkage, the kinematic optimization is performed taking into account the criteria XP [XP min, XP max], and Y [Y min, Y max] (the other two variations, YP and Z, are minimal for the initial configuration of the mechanism), for ZGS = ZGd (the values of the generator motions simulate the vertical displacement of the axle, without roll motion). In the case of opposite displacements of the left and right wheels (ZGS = -ZGd), the influence of the design variables on the objective functions are insignificant, considering rational-constructive modifications of the design variables.

The optimization study was conducted in ADAMS/View by using the GRG (Generalized Reduced Gradient) algorithm from the OPTDES code of Design Synthesis [6]. In these terms, the diagrams shown in figure 7 (in which the independent kinematic parameter is the vertical position of the wheels/axle) present the variation of the longitudinal displacement of the axle (XP) and its proper rotation around the transversal axis (Y), for the initial (before optimization) and optimal guiding mechanism. The corresponding initial and optimal values of the selected design variables are presented in table 1.

Fig. 7. Results of the optimization study.

Table 1. The initial and optimum values of the design variables (in mm).

DV XM0s(d) YM0s(d) ZM0s(d) XN0s(d) ZN0s(d) initial value 2014.5 536.0 40.0 2362.0 168.0

optimal value 2026.0 536.0 41.5 2130.0 167.5


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As can see, the optimal design study leads to significant improvements in the kinematic behavior of the axle guiding linkage, with minimal changes in the geometric configuration of the mechanism, and this proves the viability (usefulness) of the adopted optimization strategy. References [1] Alexandru C., Pozna C. Dynamics of mechanical systems using virtual prototyping tools

(in Romanian). Transilvania University of Braşov Press, 2003. [2] Alexandru, C. Software platform for analyzing and optimizing the mechanical systems.

Proceedings of the 10th IFToMM International Symposium on Science of Mechanisms and Machines - SYROM, Springer, 2009, p. 665-677.

[3] Alexandru, C. The kinematic optimization of the multi-link suspension mechanisms used for the rear axle of the motor vehicles. Proceedings of the Romanian Academy, Series A: Mathematics, Physics, Technical Sciences, Information Science, 2009, vol. 10 (3), p. 244-253.

[4] Haug E. J. Computer aided kinematics and dynamics of mechanical systems. Allyn and Bacon, 1989.

[5] Haug E. J., Choi K. K., Kuhl J. G., Vargo J. D. Virtual prototyping simulation for design of mechanical systems. Journal of Mechanical Design, 1995, vol. 117(63), p. 63-70.

[6] *** Getting started using ADAMS/View. MSC Software, 2005.


56

LOSS FACTOR AND DYNAMIC YOUNG MODULUS DETERMINATION FOR COMPOSITE SANDWICH BARS

REINFORCED WITH STEEL FABRIC

1Assistant Phd. Eng., Cosmin-Mihai MIRIŢOIU Postdoctoral researcher at University of Craiova, Faculty of Mechanics, Department of

Vehicles, Transports and Industrial Engineering, Calea Bucuresti Street, no. 107, Craiova, Code 200512, Romania, [email protected]

Abstract. In this paper I have build some composite sandwich bars. For these bars I have determined the dynamic response by recording their free vibrations. These bars have the core made of polypropylene honeycomb with upper and lower layers reinforced with steel wire mesh. For these bars I have determined the the eigenfrequency of the first eigenmode in this way: the bar was embedded at one end and free at the other where there was placed an accelerometer at 10 mm distance from the edge and I applied an initial force at the free end. I have determined the eigenfrequency because I will use its values for the loss factor and dynamic Young modulus determination. Keywords: steel fabric, sandwich bar, loss factor, dynamic Young modulus

1. Introduction The vibration of complex structures has been a subject of interest for many engineers

during the time. In this sense, studied regarding the modal identification in complex mechanical structures such as buildings, ships or aircrafts are presented in Hodges (1986) [1]. In Nakra (2001) [2] there are presented the damping mechanisms in materials and their characterization for viscous, hysteretic, coulomb and viscoelastic types of damping. There are outlined the techniques for damping characterization and the influence of parameters like frequency or temperature. There are presented also the ways for damping measurement, like: logarithmic decrement, viscous damping ratio, loss factor, Q factor and energy ratio. The elaboration and the characterization (regarding the mechanical characteristics) of some epoxy textile fibers composites were presented in Tărâţă (2000) [3].

The material loss factor for technically orthotropic plates was measured by using the half-power bandwidth method in Mandal (2004) [4]. There was used the concept of single degree of freedom system. The aim of the made tests were to highlight the effects of bending rigidity and mode orders over the material loss factor. It was observed that if the bending rigidity is higher, the loss factor is increased too. The values of loss factor in corrugated plates were higher than the ones from isotropic plate. In Cremer (1988) [5] there was shown that the half-power bandwidth method is useful for small loss factor and the thickness of the sample must be significantly smaller than the corresponding wavelength. This method can be applied in the case of beams and plates. In Vinson (2005) [6] there are presented techniques for vibration damping in sandwich structures.

In Jianxin (1999)[7], some calculus relations for the motions equations and boundary conditions, for the nonsymmetrical composite plates with active and passive damping layers being in vibration, were determined. There was investigated the influence of the inverse and direct piezoelectric effects on the frequencies and loss factors.

Kumar (2009) [8] has presented the vibration and damping characteristics of beams with active constrained layer treatments under parametric variations. The study aimed to


57

examine the effect of parametric variation of active constrained layer on the vibration control of the beams treated with optimally placed active or passive constrained layer damping patches. Other researches regarding the usage of composite materials for civil constructions are presented in Burada (2010) [9] and Burada (2010) [10].

2. Loss factor and dynamic Young modulus determination I have built some new composite sandwich bars with classical components combined

in an original way: the core is made with polypropylene honeycomb core reinforced with two layers of steel wire mesh. The thickness of the core is 10, 15 and 20 mm. I have chosen the bars width to be of 40 and 50 mm. The length of the bars will be 390 mm. In order to be easily identified, I have marked the bars like in table 1.

Table 1. The procedure to mark the bars

Sample set Specific Mass [kg/m]

Width Thickness

1 0,185 40 10 2 0,236 50 10 3 0,201 40 15 4 0,251 50 15 5 0,210 40 20 6 0,272 50 20

A general view with the sample from set 6 in presented in fig. 1. A general view with

the sample from the set 5 is presented in fig. 2.

Fig. 1. General view with a sample from set 6 In order to determine the damping factor we have used the next experimental montage

(its schematization is presented in fig. 3): the bars are clamped at one end in a massive vise and were left free at the other end. At a distance of 10 mm from the free edge, an accelerometer Bruel&Kjaer type was placed (with 0,04 pC/(m/s2). At the free edge was initially applied a force and the bar was let to vibrate freely. The accelerometer was connected to a signal conditioner NEXUS type. The dynamic response was recorded with a data acquisition system SPIDER 8 made by HBM connected through USB port with a notebook.


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The signal conditioner was connected with the data acquisition system. All the experimental recordings were made two times and the arithmetic mean of the obtained values was made in the end to determine the final half of the damping factor per unit mass and the eigenfrequency.

Fig. 2. A general view with the samples from set 2

Fig. 3. Experimental montage schematization

Fig. 4. The first experimental recording in the measuring point (set 6, L= 300 mm)


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I have considered two variants of the free length: variant 1 of 300 mm and variant 2 of

350 mm. The two experimental recordings in the measuring point are presented in fig. 4 and fig. 5. Also, the experimental processing for the half of the damping factor per unit mass calculus and the eigenfrequency of the first eigenmode, for the two experimental recordings, is presented in fig. 6 and fig. 7.

Fig. 6. The first experimental processing in the measuring point (set 6, L= 300 mm)

Fig. 5. The second experimental recording in the measuring point (set 6, L= 300 mm)


60

From the fig. 6 and fig. 7 we can see that, by using the logarithmic decrement, for 5

number of cycles, half of the damping factor and the eigenfrequency has been determined. All the obtained data, for all the samples, have been written in table 2 (processed from the experiments to be used for the loss factor calculus). To determine the bars loss factor, the formula (1) can be used. Formula (1) is obtained by writing the basic formula for a vibration in small damping domain.

13183098861,0 ; μ= 0,5∙c (1)

The results have been written in table 3. For the dynamic elasticity modulus, the formula (2) can be used (according to Jung

(2006) [8]). 2244438884,21 gLE ; μ= 0,5∙c (2)

The results have been written in table 4. In (2) I have marked with: ρ – the material

density of the bar; L – the bar free length; ち – the eigenfrequency; g- the bar thickness. Using the same methodology from Miriţoiu (2014) [9] and Miriţoiu (2014) [10] I have

determined a correlation between the loss factor and the bars thickness. I have used a second degree polynomial function because the correlation factor for the linear, logarithmic, power and exponential is bellow 0,9. The disadvantage is that, the written calculus formulas can only be used for the experimental results obtained in this research. For other researches, there must be searched for other functions. So, I propose for the samples with 40 mm width the formula (3) and for the samples with 50 mm the formula (4).

η(g)= -0,006∙g2 + 0,028∙g + 0,036 (3) η(g)= -0,005∙g2 + 0,02∙g + 0,044 (4)

Fig. 7. The second experimental processing in the measuring point (set 6, L= 300 mm)


61

Table 2. Experimental obtained data for the loss factor calculus

Set Variant I Variant II

c [(Ns/m)/( kg∙2-1)]

13183098,0 [s]

c [(Ns/m)/( kg∙2-1)]

13183098,0 [s]

1 16,80 0,00694393 13,74 0,00921301

2 16,94 0,00693032 12,86 0,00970752

3 26,72 0,00511587 18,94 0,00663836

4 24,84 0,0051724 17,46 0,00691827

5 32,86 0,00399234 23,74 0,00535334

6 29,90 0,00393217 22,68 0,00531934

Table 3. Bars loss factor

Sample no. 1 2 3 4 5 6 η (Variant 1) 0,058 0,059 0,068 0,064 0,066 0,059 η (Variant 2) 0,063 0,062 0,063 0,06 0,064 0,06

Table 4. Bars dynamic Young modulus [MPa]

Sample no. 1 2 3 4 5 6 E (Variant 1) 3017 3091 1789 1748 1295 1383 E (Variant 2) 3175 2918 1969 1811 1334 1400

Important remark: the presented research is a continuation to the researches from Miriţoiu (2014) [12] and Miriţoiu (2012) [13].

3. Conclusions In this paper I have built some new original composite sandwich bars with the core

made of polypropylene honeycomb and the exterior layers reinforced with steel wire mesh. Then, for each bar, I have experimentally determined the damping factor and the eigenfrequency. I have used their values to determine the bars loss factor and the dynamic Young modulus.

The added value of this paper is: - building some new original composite bars made of classical materials but combined

in an original way; - the experimental setup: clamping at one end and the other end is free, where there is

placed an accelerometer; - determining the damping factor per unit mass by using the logarithmic decrement;


62

- determining the eigenfrequency; - determining the bars properties like the specific mass; -determining the loss factor; - determining the elasticity modulus; - obtaining a calculus formula for the loss factor depending on the bars thickness, the

calculus formula is based on a second degree polynomial function. This type of bars can be used for: -ship floor building, -plane floor building, -the frames for concrete forming, -making parts of car and bus bodies and so on. 4. Acknowledgement This work was supported by the strategic grant POSDRU/159/1.5/S/133255, Project


References

Hodges, C.H., Woodhouse, J., (1986) Theories of noise and vibration transmission in complex structures, Rep. Prog. Phys., 49, 107-170. Nakra, B.C., Vibration damping (2001), PINSA, 67 (4&5), 461-478. Tărâţă, D., Stănescu, G., (2000) Contributions to the elaboration and characterization of certain epoxi-textile fibre composites, Conferinţa Internaţională de comunicări ştiinţifice Iaşi- Chişinău, mai 2000, Univ. Tehnică Gh. Asachi Iaşi, 157-160. Mandal, K.N., Rahman, R.A., Leong, M.S., (2004) Experimental study on loss factor for corrugated plates by bandwidth method, Ocean Engineering, 1313-1323. Cremer, L., Heckel, M., (1988) Structure-borne sound: Structural vibration and sound radiation at audio frequencies, 2nd Springer Verlag, Berlin. Vinson, J., R., (2005) Sandwich Structures: Past, Present and Future, Sandwich Structures 7: Advancing with Sandwich Structures and Materials, 3-12 Jianxin, G., Yapeng, S., (1999) Vibration and Damping Analysis of a Composite Plate With Active and Passive Damping Layer, Applied Mathematics and Mechanics, English Edition, 20(10), 1075-1086 Kumar, N., Singh, S.P., (2009) Materials and Design, 30, 4162-4174. Burada, C., (2010) Determinări practice în laboratoarele de analiză şi încercări pentru construcţii, Editura Universitaria Burada, C., (2010) Încercări de laborator pentru materiale de construcţii, Editura Universitaria, Craiova Jung, S.S., Kim Y.T., Lee, Y.B., (2006) Measurement of the Resonance Frequency, the Loss Factor, and the Dynamic Young’s Modulus in Structural Steel and Polycarbonate by Using an Acoustic Velocity Sensor, Journal of the Korean Physical Society, 49(5), 1961-1966. Miriţoiu, C.M., (2014) Correlations between some mechanical characteristics and the mass per unit length for some composite bars reinforced with steel fabric, Fiability and Durability, 1, 112-119. Miriţoiu, C.M., Bolcu, D., Stănescu, M., M., Ciucă, I., Cormos, R., (2012) Determination of Damping Coefficients for Sandwich Bars with Polypropylene Honeycomb Core and the Exterior Layers Reinforced with Metal Fabric, Materiale Plastice, 49(2), 118-123.


63

RESEARCH REGARDING THE BREAKING STRENGTH AND THE SECTION MACROSCOPIC SHAPE FOR COMPOSITE SANDWICH

BARS REINFORCED WITH STEEL WIRE MESH

1Assistant Phd. Eng., Cosmin-Mihai MIRIŢOIU Postdoctoral researcher at University of Craiova, Faculty of Mechanics, Department of

Vehicles, Transports and Industrial Engineering, Calea Bucuresti Street, no. 107, Craiova, Code 200512, Romania, [email protected]

Abstract. In this paper, starting from some samples built from polypropylene honeycomb core (with the thickness of 10, 15 and 20 mm) reinforced with steel wire mesh, I have determined the breaking strength by using the homogenization and Strength of Materials theories. Also, I have presented the macroscopic shape of the transversal breaking section. In the end I have inserted some new formulas, using the regression analysis, to determine the breaking strength depending on the sample thickness. Keywords: composite bar, sandwich bar, breaking strength, stress

1. Introduction In Krysan (1986) [1] there has been studied the behaviour of composite materials in

the vicinity of a stress raiser. Using Mellin’s transformation, it is proposed a solution for two connected wedges made of different anisotropic materials in the plain strain conditions or the generalized stress plane state. There is analyzed the special feature of the stress state at the boundary of the element from the composite-metal joint. The results obtained on the basis of theoretical investigations were verified for the boron-reinforced aluminum-titanium joint.

In Degtyarev (1972) [2] there is studied the deformation criteria under simple and composite stresses. There is solved the problem of the material limiting state in the composite stress state. Strain criteria of failure for structural steel with alloys have a number of advantages over stress criteria, since, in an experimental study of the material strain property subjected to various stress conditions, a number of fine distinctions were discovered associated with the in homogeneity of the strain process in specimens and with the effects of strain anisotropy.

In Baldan (2004) [3] are examined the factors that affect the mechanical and environmental durability and performance of the adhesively bonded joints in various adherents including metallic alloys, composites, polymers. There are presented two basic mathematical approaches for the analysis of adhesively bonded joints: a closed-form model (solved by analytically means) validated with finite element analysis. The two methods used present similar results. Some studied regarding the mechanical properties for composites reinforced with textile fibers were made in Tărâţă (2000) [4].

Berbinau (2001) [5] has investigated the failure of orthotropic laminates with a filled hole subjected to biaxial compression-tension loading. The considered case simulates a situation where the impact damaged laminate has been repaired by drilling a hole and then plugging the hole with a perfect-fit core made of dissimilar material. The failure strength predictions are compared to the open hole results and experimental data.


64

Van de Meer (2012) [6] has presented a review for the mesolevel modeling of failure in composites laminates. There are discussed, in the end, the limitations and challenges for mesolevel analysis of composite failure. Other composite researches were made in Burada (2010) [7] and Burada (2011) [8].

2. Experimental determination of the maximum force and displacement Some composite materials samples with polypropylene honeycomb core have been

built. These have been marked in this way: - set 1: core = 10 mm, width= 40 mm; - set 2: core = 10 mm, width= 50 mm; - set 3: core = 15 mm, width= 40 mm; - set 4: core = 15 mm, width= 50 mm; - set 5: core = 20 mm, width= 40 mm; - set 6: core = 20 mm, width= 50 mm. General views with the samples from the set 1 and 2 are presented in fig. 1 and 2.

Fig. 1. A general view with the sample 1

Fig. 2. A general view with the sample 2


65

The samples have been subjected to three points bending on the Walter Bai testing machine (maximum force is 30 kN). The bending scheme is presented in fig. 3.

The force versus the displacement, for all the samples 1, 2 and 3, is presented in fig. 4, 5, and 6. The maximum forces (with approximations), where there is produced the samples breakage, are:

a.

b.

Fig. 3 Bending loading; a. loading scheme; b. the fracture process of the studied samples


66

Fig. 4. Force versus displacement (sample 1)

F [N]

Displacement

F [N]

Displacement Fig. 5. Force versus displacement (sample 2)

Fig. 6. Force versus displacement (sample 3)

F [N]

Displacement


67

- sample 1: 333 N; - sample 2: 376 N; - sample 3: 416 N; - sample 4: 422 N; - sample 5: 430 N; - sample 6: 448 N.

3. Breaking strength calculus In order to rapidly obtain the breaking strength, by using the homogenization theory, I

will use the formula (1). σr= Mmax∙Wy

-1= 0,25∙Fmax∙a∙Wy-1 (1)

In (1) I have marked with Wy the axial strength modulus which is determined with the classical formula from Strength of Materials, corresponding to a rectangular area. The results are:

- sample 1: 29,97 MPa; - sample 2: 27,072 MPa; - sample 3: 16,64 MPa; - sample 4: 13,504 MPa; - sample 5: 9,675 MPa; - sample 6: 8,064 MPa. Using the regression analysis, I have determined direct calculus formulas for the

breaking strength, which depend on the bars thickness. So, if the bars have a width of 40 mm I recommend the formula (2), and for the width equal to 50 mm, formula (3). The correlation factor R2 for the formula 2 is 0,999 and for the formula (3) is 0,9927.

σr (g)= 52,3439045626∙e-0,5653258212∙g (2) σr (g)= 48,1372103971∙e-0,6055451360∙g (3) 4. Macroscopic aspect of the section failure The macroscopic aspects of the section failure are presented in fig. 7 (samples 1, 2, 3,

4) and fig. 8 (samples 5 and 6).

a. b.

c. d.

Fig. 7. Macroscopic aspect of the section failure; a. sample 1; b. sample 2; c. sample 3; d. sample 4


68

We can see that the fracture has been produced starting from the lower side of the composite bar, the face which is opposed to the acting roller.

5. Conclusions In this paper I have built some new original composite sandwich bars with the core

made of polypropylene honeycomb and the exterior layers reinforced with steel wire mesh. Then, for each bar, I have experimentally determined the maximum force, displacement, the characteristic curve of the material which was used to build the bars, the breaking stress.

The added value of this paper is: - building some new original composite bars made of classical materials but combined

in an original way; - determining the dependence between the force and the displacement; - determining the force-displacement curve of the material used to make the bars; - determining the breaking strength; - determining some calculus relation for the breaking strength depending on the bar

thickness. This type of bars can be used for: ship floor building, plane floor building, the frames

for concrete forming, making parts of car and bus bodies and so on. 4. Acknowledgement This work was supported by the strategic grant POSDRU/159/1.5/S/133255, Project


References

1. Krysan, V.A., Nikitin, L.V., (1986) Failure of the element of the composite-metal joint in stress concentration conditions, Mekhanika Kompozitnykh Materialov, 2, 269-275. 2. Degtyarev, V.P., (1972) Deformation criteria of failure under simple and composite stresses, Problemy Prochnosti, 7, 22-25. 3. Baldan, A., (2004) Review: Adhesively-bonded joints in metallic alloys, polymers and composite materials: Mechanical and environmental durability performance, Journal of Materials Science, 39, 4729-4797.

a. b. Fig. 8. Macroscopic aspect of the section failure; a. sample 5; b. sample 6.


69

4. Tărâţă, D., Stănescu, G., (2000) Contributions to the elaboration and characterization of certain epoxi-textile fibre composites, Conferinţa Internaţională de comunicări ştiinţifice Iaşi- Chişinău, mai 2000, Univ. Tehnică Gh. Asachi Iaşi, 157-160. 5. Berbinau, P., Filiou, C., Soutis, C., (2001) Stress and Failure Analysis of CompositeLaminates with an Inclusion under Multiaxial Compression-Tension Loading, Applied Composite Materials, 8, 307-326. 6. van der Meer, F.P., (2012) Mesolevel Modeling of Failure in Composite Laminates: Contitutive, Kinematic and Algorithmic Aspects, Arch Comput Methods Eng, 19, 381-425. 7. Burada, C.O., (2010) Laboratory tests for asphalt mixtures, Metalurgia International, 9, 233-237. 8. Burada, C.O., (2011) Experiments regarding the effect of reinforced concrete and reinforced bars processing on mechanical-physical characteristics, Metalurgia Inernational, 1, 94-97. 9. Stănescu, M., M., Bolcu, D., Ciucă, I., Miriţoiu, C., Baciu, F., (2012) Experimental Determinations of Stiffness and Damping Coefficient of the Sandwich Bars with Core of Polypropylene Honeycomb with Steel Fabric, Materiale Plastice, 50 (3), 179-183.


70

THE GEOMETRY OF THE SPATIAL FOUR-BAR MECHANISM AND OF ITS PARTICULAR FORMS

Păun ANTONESCU, PH.D. „POLITEHNICA” UNIVERSITY OF BUCHAREST, [email protected]

Ovidiu ANTONESCU, PH. D. „POLITEHNICA” UNIVERSITY OF BUCHAREST, [email protected]

Constantin BREZEANU, SILCOTUB S. A. (TENARIS GROUP) CĂLĂRAȘI, [email protected]

ABSTRACT: starting from the rssr spatial four-bar mechanism, this paper analyses the most significant particular cases for which the specific transfer functions are written in comparison to the respective function deduced in the general case. From among the particular cases of the spatial mechanisms, the authors mention the oscillating washer mechanism and the cardan mechanism, demonstrating that spherical mechanisms can be obtained as the simplified variants in which the spherical joints may be replaced by simple rotation joints. For the spherical mechanism, the respective transfer functions have been determined as the simplest particular forms of the transfer function achieved by the rssr spatial four-bar mechanism.

KEYWORDS: spatial four-bar mechanism, spherical joint, transfer function, cardan mechanism 1. GENERAL CONSIDERATIONS

The bar spatial mechanisms rssr (fig. 1a) and rsst (fig. 1b) are used in numerous fields of manufacturing engineering such as: sewing machines in the textile industry, agricultural equipment, hydraulic pumps and engines, wind screen wiper mechanisms.

These mechanisms consist of a minimum number of kinematic elements, are some of the simplest spatial mechanisms [1, 2] and have the advantage that they can achieve, for a small gauge, the transmission and transformation of the rotation motion between the in and out elements whose axes are displaced randomply in space.

x

1l

ll

A

23

B

0A

1y

y

1x

0

B0

s

s1

3

01

z

l0

l1

A1s

0

0A

B

1 s

0

l0

l2

03

z

B

y1

y

x

s

a. b.

x

y

z

A 1

A 1

A 1

= s

= l sin

= l cos

A

x = s

y =

z = l

B 3

0

cos +

-

l3 sin sin

BB

s3 sin l3 sin cos

B + l3 cos

A

x s= 1A

y A= sinl1

cosz l=A 1

-

+3=x s cos

y B

=z lB 0

sin= sB 3

B

-

sin

cos

s

s

Fig. 1. Kinematic diagrams of the rssr and rsst spatial mechanisms


71

As compared to other types of spatial mechanisms with upper kinematic couplings (with

spatial cams or with conical / auger gears), bar spatial mechanisms are made of lower kinematic couplings (cylindrical and spherical), which offers simpler technical solutions (requiring less expensive technologies), and has higher reliability in difficult working conditions.

Starting from the general case of the rssr spatial quadruple mechanism, the paper deals with the classical particular cases [2] as well as with the new ones [1, 7]. These mechanisms have apparently no geometrical or construction characteristics in common.

Thus, the transmission functions for each type of spatial mechanism can be easily obtained from the transmission function deduced in the general case of the rssr spatial mechanism [1, 2, 6]. If we take the general case of the rssr spatial mechanism (fig. 1a), as written out in the diagram, the transmission function can be written as [2, 6]: 0)sinsincoscos(coscossincossin 543210 aaaaaa (1)

The relation (1) is a function of the type 0),( F , where the angles and place the crank 1 (drive element) and the equalizing bar 3 (driven element) as to the axis oz (fig. 1a). If we use the notations in the kinematic diagram (fig. 1a), the coefficients in the equation (1) can be expressed as follows:

cos2 3123

21

23

22

21

200 sssslllla ;

102311 2;sin2 llasla ; (2)

315304313 2;2;sin2 llallalsa .

2. The rssr mechanism with orthogonal axes ( 090 ) In the particular case of 090 (fig. 2) the o1b0 axis is parallel to the oy axis, so that the

crank 1 and the equalizing bar 3 rotate in perpendicular planes.

x

l1

A 1s

0

0A

B x1

1s

0l0

l20 3

l3

z B

y1

y

= 90o

x

1l

0A

l

l

A

2

3

B

1y

z = z1

y

1x

B0

0

l0

s1

x

A

1l A0

1l

l2

3

B

y

z = z1

y

1x

l0

B0

os = s = 01 3= 0

Fig. 2. Kinematic diagram of the rssr spatial mechanism when 0,0,90 31

00 ss

The coefficients 10, aa and 3a which depend on the angle therefore become: 23

21

23

22

21

200 sslllla ;

311 2 sla ; 313 2 lsa . (3)


72

This mechanism where the in and out axes are perpendicular (fig. 2) is called an orthogonal spatial quadrangle [1, 5]. For 090 the transmission function (1) is written:

0coscoscossincossin 543210 aaaaaa (4)

3. THE RSSR MECHANISM WITH PARALLEL AXES ( 00 ) When 00 the rotation axes of the crank 1 and the equalizing bar 3 become parallel

(fig. 2) when we can consider 03 s , so that the point b0 is on the oz axis and the equalizing

bar 3 rotates in the yoz plane. For 00 180,0 the 1a and 3a coefficients are null, and the 0a coefficient is

231

23

22

21

200 )( sslllla (5)

This particular case (fig. 2) corresponds to the plane-parallel of the elements 1 and 3 whose rotation axes are parallel. The transmission function (1) coincides in this case (fig. 2) to the one of the plane mechanism:

0)cos(coscos 5420 aaaa (6)

4. THE RSSR MECHANISM WITH 031 ss The two rotation axes of the crank 1 and of the equalizing bar 3 are displaced randomly in

space (fig. 2), but the points a0 and b0 are situated on the oz axis. Since 031 ss , the 1a and

3a coefficients are null ( 01 a , 03 a ), and the 0a coefficient can be inferred from (5)

23

22

21

200 lllla (7)

The transmission function (1) becomes in this case 0)sinsincoscos(coscoscos 5420 aaaa (8)

For the orthogonal spatial mechanism ( 090 ), the transmission function (8) becomes 0coscoscoscos 5420 aaaa (9)

The formula (9) can be also inferred from (4) if we introduce the conditions 031 ss .

5. THE RSSR MECHANISM WITH 00 l The condition 00 l imposes the concurrency of the rotation axes of the in and out

elements (fig. 3), which corresponds to the geometrical condition of spherical mechanisms. From the relations (2) null values for the 2a and 4a coefficients are obtained ( 042 aa ),

and for the coefficient 0a we infer the expression

cos2 3123

21

23

22

210 ssssllla (10)


73

B

l

A

l

x

y

z

A

B

y

s

l

s

x

0

2

1

3

3

1

0

0

1

1

0

z

B

y

x1

A

x

1l

45o

45o

a. b.

l = 0 ; s =1 ; s = l A0B = 90o

l = 0 0 1 3 3 ;0

Fig. 3. Kinematic diagram of the rssr mechanism with concurrent axes

For the transmission function (1) we obtain the expression 0)sinsincoscos(cossinsin 5310 aaaa (11)

In the particular case when the oa and ob lines are perpendicular (fig. 3b), we have the following relation between the specific liniar lengths of the mechanism

22

23

21

23

21 lssll (12)

This condition determines a much simpler expression for the 0a coefficient

cos2 310 ssa (13)

Considering the formulas of the 31, aa and 5a coefficients of (2) versus the formula of the

0a coefficient in (10) or (13), it is inferred that, when 3311 ; lsls , the transmission function

(11) becomes: 0)sinsin1(cos)sin(sinsincoscos (14) Which shows that the function 0),( F no longer depends on the liniar parameters of

the mechanism, this being specific to the spherical mechanism. Indeed, in this particular case, the s type spherical joints in points a and b (fig. 3a) can be

replaced by r type cylindrical joints (fig. 3b). For the orthogonal spherical mechanism ( 090 ), the transmission function (14) becomes 0sinsincoscos (15)

6. THE RSSR MECHANISM WITH 0,0 310 ssl This particular case of the spatial mechanism corresponds to the simple spherical

mechanism of the crank-equalizing bar type (fig. 4). If we take into account the above mentioned geometrical conditions (fig. 4), the

4321 ,,, aaaa coefficients are null, and the equation of the transmission function is expressed in a simplified form:

0)sinsincoscos(cos50 aa (16)


74

l = 0 ; s l= l2

2

2

1

2

3

2

1+ + sl0 3 0 ; =0l0= ; s =03s=1

z

x

y

B0

x1

01

y

2l

1l

l3

B

A0B

1l0

x

2ll

A

3

B

y

1x

z

s

A0

1

Fig. 4. Spherical rssr mechanism Fig. 5. Cardan rssr mechanism

Where the 0a and 5a coefficients are expressed:

23

22

210 llla ; 315 2 lla (17)

When the x and x1 axes (fig. 4) are perpendicular, we can infer the transmission function from (16) for 0 as below

0coscos50 aa (18)

7. THE RSSR MECHANISM WITH 0,0,0 030 asl

For these particular values, between specific geometrical parameters we can infer from (2) the relation

23

21

21

22 lsll (19)

Which corresponds to the kinematic scheme of the rssr mechanism (fig. 5) when the ab0b triangle is a right triangle, that is 0

0 90)( BAB .

If on the kinematic scheme (fig. 5) we write )( 00 ABA we can write the relation

tgsl 11 , so that the coefficient 1a of (2) becomes

tglla

sin2 311 (20)

The other coefficients of the transmission function (1) are 0432 aaa and 315 2 lla ,

which determines for the implicite function 0),( F the following particular form 0)sinsincoscos(cossinsin tg (21) In which we have only constant angular parameters ),( together with variable angular

parameters ),( . The equation (20) shows that the rssr spatial mechanism (fig. 5) can now operate as a

spherical mechanism, where the spherical couplings in a and b are replaced by rotation couplings (fig. 3b).


75

If we consider the case of perpendicular axes with this mechanism (fig. 5) )90( 0 , this corresponds to the spherical mechanism with eqalizing bar of the oscillating washer type [1, 7]. The transmission function is obtained from (21) in its simple form:

cos tgtg (22)

8. THE RSSR MECHANISM WITH 0,0,0 0310 assl For the new geometrical conditions imposed to the spatial rssr mechanism (fig. 6) the

relation (19) is written 23

21

22 lll .

which corresponds to the constant 900 angle between segments a0a and a0b (fig. 6).

l1

x

l2

A

B

l3

y

x1

y1

A0

1z = z

l0l 0= =3s; ;03

2l=2

2l+1

2s1=

Fig. 6. Kinematic diagram of the cardan rssr mechanism

The spherical mechanism obtained (fig. 6) is of the cardan type, where the transmission function is inferred from the equation (16) for 00 a as follows:

0sinsincoscoscos (23) From equation (23) we obtain the expression

cos

1 tgtg (24)

In this case we can replace the spherical couplings in a and b (fig. 6) by rotation cylindrical couplings, just like with the classical cardan spherical mechanism.

we should mention that formula (23) represents the transmission function of the mechanism known as simple cardan coupling, its expression being different from the one known and given in the papers [3, 4, 7].

This is due to the way in which the and angles are measured, so that in this case the two angles are measured against the same direction parallel to the a0z axis that is common to both cartesian reference points a0xyz and a0x1y1z1 (fig. 6).

We should also notice that, if the and angles are measured against the a0y axis and the a0y1 axis (fig. 6), the transmission function is obtained as follows [3, 8]: cos tgtg

If the angle is measured against the a0y axis, and the angle is measured against the a0z1 axis (fig. 6), the transmission function becomes:

l0 =0, s1=s3=0, 23

21

22 lll ,

A

B

A0 A’

B’ x

x1

y

y1

z, z1

l2

l1

l3

l1

l3

.

.

l3

l3

l1

B1

A1

ψ


76

cos

1ctgtg (25)

With cardan transmissions, the angle(between the axes a0x and a0x1 ) is an obtuse ( 090 ), therefore 0cos . If in this case we consider the acute angle 0180 (fig. 6), the transmission function (25) can be written in its well-known explicite form [1, 4, 7, 9]:

cos

tgtg (26)

9. Conclusions Starting from the representation of the kinematic diagram of the rssr spatial mechanism, in

an axonometric projection (fig. 1a), in the paper we infer the expression of the transmission function 0),( F in its implicite form (1), of which specific forms are obtained for different cases.

Of the specific variants of the rssr spatial quadruple mechanism, the authors mention in particular spherical mechanisms, known as crank – equalizing bar spherical mechanism (fig. 7), called sherical mechanism with oscillating washer and the crank – crank spherical mechanism (fig. 8) called cardan (coupling) mechanism.

It is proven that these two spherical mechanisms, which have multiple applications in manufacturing engineering, can be obtained as simplified variants of the rssr quadruple mechanism (fig. 1a), in which spherical couplings are replaced by rotation cylindrical couplings. The cardan type spherical rssr quadruple mechanism can be achieved as two parallel chains (fig. 6), with equal length ab and a’b’ reciprocating rods.

REFERENCES

1. Antonescu, P., Mechanisms - Structural and Kinematic Calculation, Polyt. Inst. Press, Bucharest, 1979;

2. Luck, K., Modler, K-H., Einfache raumgetribe für getribetechnishe grundaufaben, Wiss. ZDTU. Dresden, 1978;

3. Manafu, V., Theory of Mechanisms and Machines. Structure and kinematics. The Technical Publishing House, Bucharest, 1959;

4. Autorenkollektiv, Getriebetechnik Lehrbuch, VEB Verlag Technik, Berlin, 1969; 5. Antonescu, P., Contributions to the Graphic Synthesis of Spatial Mechanisms, Ph.D.

Thesis, The Polytechnics Institute Press, Bucharest, 1969; 6. Alexandru, P. and others., Mechanisms Vol. II, Synthesis, Braşov University Press,

1984; 7. Dudiţă, F., Cardan Transmissions, The Technical Publishing House, Bucharest, 1966; 8. Pelecudi, Chr. and others, Mechanisms, E. D. P. Bucharest, 1985; 9. Antonescu, P., On the Specific Cases of the Spatial Quadruple Mech., SYROM’89, vol.

II.1, p. 1-10; 10. Antonescu, P., Mechanisms, Printech Publishing House, Bucharest, 2003. 11. Vişa, I. and others, Functional Design of Mechanisms, Classical and Modern Methods,

Braşov, 2004.


77

STRUCTURAL AND GEOMETRICAL ANALYSIS OF THE LIFTING MANIPULATORS FOR A GREEN ENVIRONMENT

Ioana POPESCU, Iuliu Maniu Highschool, Bucharest, e-mail: [email protected]

Dr. Ovidiu ANTONESCU, Univ. Politehnica of Bucharest, e-mail: [email protected] Dr. Păun ANTONESCU, Univ. Politehnica of Bucharest, e-mail: [email protected]

Abstract: The lifting and getting off the bins, to and from the body of special waste trucks, by some planar linkage – manipulators are studied. These lifting manipulators are equipped with gripper systems in order to load and unload the bins. Several kinematical schemas of type mono– and bi-mobile manipulators are analyzed, these being driven by one or two linear actuators. The kinematical geometry of these planar manipulators by means of scale drawing of the kinematical schema is displayed. Two solutions for a better efficiency and a green environment have been proposed. Finally, a modeling and simulation case of the lifting manipulator is presented.

Keywords: lifting manipulator, mobilitie, simulation 1. INTRODUCTION

The development of lifting manipulators for loading and unloading the waste freight into and from specialized trucks has not been treated so much in literature [2].

One of the best reference titles on bin lifting automotive history is “The photographic archive of waste trucks” by John B. Montville [4] that presents the development of garbage gathering vehicles since First World War to nowadays.

The first waste vehicles had an open top part of the body to collect the garbage though they were not specially designed to perform this task. By 1920 about twenty garbage trucks with closed carriage were accomplished in Great Britain. The advantage of this type was a bigger quantity of garbage that can be loaded in cleaner conditions into a greater carriage.

First truck carriage with outer bunker had been made in 1929 and the rear loading body with waste compactor in 1938. This principle of rear loading compactor carriage is the most used in present, even if at that time the waste bins were manually lifted and unloaded. Other models of waste trucks are with side (1947) or front (1955) loading.

2. WASTE BINS AND LOADING MECHANISMS The waste is collected in special containers or bins [3], [4] of diverse sizes being made of

steel (for larger dimensions) or plastic.


78

Fig. 1. Waste loading process Fig. 2. Mechanical bin lifting

After the mechanical loading of the waste (fig. 2) from bin into the receiving bunker of

truck, the garbage is pushed into the main compartment of carriage (fig. 1). These bin lifting mechanisms (fig. 1) represent mono-mobile or bi-mobile manipulators

[1] that grasp the bin, lift it until the receiving bunker level and lean it until the waste begin to fall into bunker. Ones the bin is rotated by over 90 degrees from initial position, the bin opens by itself maintaining the lid in vertical position (fig. 1 or 2) and the waste is unloaded.

3. MONO-MOBILE LIFTING MANIPULATORS Let’s consider a mono-mobile lifting mechanism for heavy containers (fig. 3). This

manipulator is a planar mechanism [1] that consists of one closed kinematical contour with hydraulic cylinder and another kinematical contour which is alternatively open in lifting phase or closed in waste unloading phase.

Fig. 3. Mono-mobile lifting manipulator for heavy containers

Initially, the rotate cylinder 1 of manipulator is in vertical position (thick line - fig. 3) having the piston 2 at top of it so that the rocker 3 has segment BB0 in horizontal position. Bar 4 is linked to the container 5 by a hook which allows an easy hanging of it and, also, a rotation of it in unloading phase (thin line - fig. 3).


79

In final position (thin line – fig. 3) the piston 2 is at bottom of cylinder 1 and the container 5 leans with point M (now M’) on a fixed point of carriage, the open kinematical chain formed of 3, 4 and 5 element becoming closed (B0 C’ D’ M’).

In lifting phase the mechanism mobility results by using the following formula [1]:

5

1

6

2m rrm rNmCM (1)

In this phase all six kinematical joints are mono-mobile (m = 1) and there is only one closed loop of rank 3 (r = 3), so that N3 = 1.

Therefore, by formula (1) results: 31361 M Among these three mobilities (DOF) only one is active (controllable) – actuator (1+2).

The others two mobilities are passive – rotation of bar 4 related to joint C and rotation of container 5 about joint D.

In the second phase there are two closed loops of rank 3, six joints of class 1 and one joint of class 2, resulting: 2231261 M , but only one is active.

Fig. 4. Mono-mobile lifting manipulator Fig. 5. Bi-mobile lifting manipulator with for light bins parallelogram

1

2 3 5

4

6

7


80

To manipulate light bins of plastic (dashed line – fig. 4) it is used a mono-mobile planar mechanism with two closed kinematical contours, the first one having the actuator as rotate cylinder and the second one being a rotate jointed quadrilateral (fig. 4).

The kinematical scheme of manipulator has been represented in two limit positions, the bottom one displayed with continuous thick line (initial position) and the top one with continuous thin line (final position). There are three fixed axes on the truck body A0, B0 and D0. In the initial position (bin grasping) the cylinder 1 and piston 2 are in vertical position with the minimum length A0B. Bar 4 has two mobile rotate joints C and D, and also the rotate joint E by which the bin 6 is positioned.

In order to obtain the final position (waste unloading) the piston 2 slides into cylinder 1 to the end of stroke, the length A0B’ being maximum (fig. 4).

It can be observed that in initial position the rotate jointed quadrilateral B0CDD0 is convex and in the final position it becomes concave (B0C’D’D0), the two rockers 3 and 5 being crossed. As it was mentioned in previous chapter, the bin 6 opens itself by maintaining the lid 7 in vertical position, this being linked to 6 by a rotate joint F’.

The mobility is checked by formula (1): 12371 M This type of mechanism (fig. 4) allows 130-145 degrees rotation of the bin, being the

most used lifting manipulator in street salubrity. Of course, these mono-mobile manipulators (fig. 3 and 4) are achieved as double mobile

structures operating in parallel planes (on both carriage sides). Therefore, the two hydraulic actuators must be synchronized in order to lift the bin(s) properly. In the case of light bins, the manipulator lifts two or three plastic bins in the same time by using a horizontal bar (joint E in fig. 4) which links the two parallel mechanisms. On this connecting bar there are catching systems mounted, they being equipped with safety devices on manipulated bins.

4. BI-MOBILE LIFTING MANIPULATORS The bi-mobile lifting manipulators operate in the same conditions as mono-mobile ones,

being double systems mounted in parallel planes, between them the waste bins being lifted and unloaded. The actuation of these mechanisms is provided by four hydraulic cylinders, each two of them in one working plane.

Let’s now consider a bi-mobile lifting manipulator with rotate jointed parallelogram (fig. 5). The two cylinders work as following: one actuator drives in the first manipulating phase (bin lifting – bottom and middle positions) and the other actuator drives in the second phase (bin rotation – top position). This has the advantage of maintaining the bin vertically in the first phase, the movement being a circular sliding (in the same plane).

This bi-mobile mechanism (fig. 5) has the first mobility obtained by actuator (1+2) (linked to element 6 by rotate joint A and to element 3 by rotate joint C) and the second mobility obtained by actuator (7+8) (linked to truck body by rotate joint I0 and to element 6 by rotate joint J). The mobility is checked by formula (1): 233111 M

In the first phase, when element 6 is static, the piston 2 slides into cylinder 1 until the stroke ''21 CCBBs is complete. Therefore, the bar 4 (by which the bin 9 is sustained) executes a circular sliding reaching the vertical E’F’ position.


81

In the second phase, with the actuator (1+2) blocked (AB’C’), the actuator (7+8) begins to drive by sliding the piston 8 into cylinder 7 until the stroke is '' 0087 JIJIIIs . During

this phase the element 6 rotates as a rigid body, together with elements 1, 2, 3, 4 (9) and 5, around the rotate joint A0 by 130 degrees.

As it was explained in the previous case the bin lid 10 remains in vertical position, allowing the waste to fall into the receiving bunker of truck.

After the bin is completely unloaded, the actuator (7+8) slides at maximum extended position I0J and then the actuator (1+2), once it reaches the vertical position, extends to maximum stroke AC = s21max . Further, we consider a bi-mobile lifting manipulator with rotate jointed quadrilateral (fig. 6). This may be called as general case comparing with the last one (fig. 5) because in the first phase the bin is lifted and rotated to about 45 degrees, and in the second phase it continues to rotate to extra 90 degrees.

The position of fixed rotate joint I0 of the second actuator lies under the fixed rotate joint A0 on truck body. This is an advantage regarding the length of hydraulic actuator-supplying pipes: closer actuators–shorter supplying tubes. In initial position of the first phase the kinematical schema (fig. 6) is drawn by continuous thick line (including the bin).

The actuator (1+2) by cylinder 1 is rotate jointed to the bar 6 (which is rotate jointed to carriage by fixed point A0) and by piston 2 to the rocker 3 which is also rotate jointed to the bars 6 and 4. The bars 6, 3, 4 and 5 form a rotate jointed quadrilateral having “the base” 6 and 3 as driving element which is actuated by (1+2).

The kinematical schema (fig. 6) in the final position of the first phase is drawn by continuous thin line and the bin 9 by dashed thin line.

During this phase from DEFG position the quadrilateral gets to D’E’F’G’ position, where the bin 9 is rotated to about 45 degrees.

In the second phase the actuator (1+2) is blocked in retreated position (A’B’C’) so that the quadrilateral becomes as one unitary element, with no movements among its bars, which, together with bar 6, is rotate jointed in the fixed point A0.

Now, the actuator (7+8) takeovers the command, having the cylinder 7 rotate jointed in the fixed point I0 on truck body, and the piston 8 linked to element 6 in point J. The mobility is checked by formula (1): 233111 M . In the final position of second phase the manipulator is represented by continuous thin line (fig. 6) using the following notations: A”, C”, D”, E”, F”, G” and K”. In this position the bin 9 is rotated to extra 90 degrees so that the bin lid 10 takes a vertical position allowing the waste to fall into receiving bunker.


82

Fig. 6. Bi-mobile lifting manipulator with

quadrilateral Fig. 7. CAD modeling and kinematical simulation

5. MODELING AND SIMULATION OF THE LIFTING MANIPULATOR Nowadays, by means of advanced CAD software systems, any component or mechanical

assembly having a complex structure can be modeled and simulated. Designing a virtual entity, with precisely 3D dimensions (using three-dimensional space), requires a spatial vision and a detail-orientated eye. In order to simulate a process, we need that part to be as much as possible like the real one. So, a virtual material with real properties can be applied to it. Thus, we can obtain a component which has its own characteristics that can be updated anytime.

The parametrical design has the advantage that by modifying a geometrical parameter of the component, all the other dimensions will be automatically changed in correlation with the given constrains.

Further, certain virtual assemblies, obtained by connecting their mobile mechanical elements, can be created, accomplishing mechanisms with one or more mobilities that can be animated and simulated on their operation as if they would have a real behavior.

Let’s consider the mono-mobile lifting manipulator with improved efficiency depicted in figure 7. The first step is the modeling of the main mechanical components such as the base 0 (truck body), the hydraulic cylinders 1 (including the pistons 2) as actuators, the T-rockers 3 linked to cylinder’s pistons, the rods 4 fixed to the transversal shaft (that connects the two parallel lifting mechanisms and sustains the two waste bins by special supports), the rockers 5, and the waste bins 6 (including their lids 7).

7

8


83

All these components of the lifting manipulator have been modeled by using a 2D sketcher. Then, the third dimension for each element, by using the 3D module, was created. Many other commands were used in order to shape the right configuration of them. The total number of parts (subcomponents) is 24.

Afterwards, these parts were assembled using a specific virtual workbench by adding the connection elements (such as shafts, pins and bearings) between them that, in fact, represent the rotate joints (fig. 7). The only sliding joints are between pistons and cylinders. It is necessary to be mentioned that before assembling the components, the accuracy of each part dimension has to be checked, so that the outcome should be a perfect combined product.

The 3D rendering of the lifting manipulator assembly can be achieved very easy just by computer mouse. The system can be rotated, moved or zoomed in/out by user in order to see the final product from any point of view.

For the kinematical joint applying between the mechanism’s components, the manipulator mobility has to be considered. This lifting mechanism having 1 DOF, the sliding range of the driving cylinders can be imposed so that the manipulator working space to be modeled, related to the necessity of a completely waste unloading from the bins.

The kinematical operating simulation of the lifting manipulator can be achieved by using the command panel by which the virtual displacement can be controlled.

6. CONCLUSIONS

Two simply solutions on lifting manipulators have been proposed. One of them is meant to reduce the waste truck consumption by implementing a high efficiency mechanism. The other is to reduce the city pollution by using an electric powered driving transmission. The first one has been CAD modeled and simulated in order to test its kinematical performances. REFERENCES 1. Antonescu P., Antonescu O. Mechanism and Machine Dynamics (in Romanian).

Printech Publishing House, Bucharest, 2005. 2. Voicu G., Paunescu I. Processes and Machines for City Cleaning, Matrix Press,

Bucharest, 2002. 3. Coltofeanu R. Structural-Topological Analysis of the Lifting Mechanisms on Urban

Salubrity Equipments. PhD Paperwork, Politehnica Univ. of Bucharest, 2005. 4. Antonescu, O., Coltofeanu, R., Antonescu, P., Geometria manipulatoarelor pentru

descărcarea recipientelor cu reziduuri gunoiere, Rev.Mecanisme și Manipulatoare, Vol. 5, Nr. 2, 2006, pag. 25-30.

5. Geonea, I.,Coltofeanu, R., Motofeanu, S., Kinematics and dynamic modelling of a plane manipulator, Journal Mechanisms and Manipulators, Vol. 8, No 2, 2009, p. 41-46.

6. *** www.tigerdude.com/garbage ; www.mechlift.com .


84

PLANAR MECHANISMS USED FOR GENERATING CURVE LINE TRANSLATION MOTION

Dr. Ovidiu ANTONESCU, Politehnica University of Bucharest, [email protected]

Viorica VELISCU, Transport CF High School, Craiova, [email protected] Daniela ANTONESCU, „Iuliu Maniu” High School, Bucharest,

[email protected]

Abstract: The curve line translation motion can be generated in the particular form of the circular translation, through mono-mobile mechanisms with articulated links of simple parallelogram type (with a fixed side) or through transmission with toothed belt with a fixed wheel. Also, the circular translation can be generated through planar mechanisms with two cylindrical gears with a fixed central wheel. It is mentioned that the two cylindrical gearings of the Fergusson mechanisms are both exterior and interior.

Keywords: planar mechanism, circular translation motion, cylindrical gear, kinematic scheme 1. INTRODUCTION

Curve line translation motion, in particular circular translation, can be generated by means of planar mechanisms with articulated links [1, 3, 4] such as the articulated parallelogram (fig. 1), or by means of planar mechanisms with cylindrical gears [1, 2, 3] such as the Fergusson mechanism (fig. 2).

With the planar mechanism of the articulated parallelogram type (fig. 1), the reciprocating rod 2 does a circular translation motion as the velocities of points A and B are equal, which can be noticed by maintaining the segment AB in a parallel position to itself.

Fig. 1. The parallelogram mechanism Fig. 2. Fergusson planar mechanism

With the multi-level planar mechanism (fig. 2), where the gear wheels 1 and 3 are equal, if the central gear wheel 1 is fixed ( 01 ), the satellite wheel 3 does a circular translation

motion ( 03 ).

Indeed, by actuating the planet wheel carrier p ( 0p ), by means of the distribution of

linear velocities (fig. 2), it results that the velocities of points B and D are equal: ;0)( Cc )(2)()( AaDdBb

A B

A0 B0

a b

1

2

3

0

A

B

D

C(c) a

b

d

O

p

3

1(0)

2




85

2. USING THE CIRCULAR TRANSLATION MOTION IN AUXILIARY MECHANISMS OF MOTOR VEHICLES

2.1. Doors closing / opening mechanism without hinges The doors of some modern buses no longer contain hinges, and they are actuated by

means of the articulated parallelogram mechanism (fig. 3) in which the door side MN is one with the reciprocating rod AB of this planar mechanism.

Fig. 3. Kinematic scheme of the mechanism Fig. 4. Kinematic scheme of the used for actuating the bus door screen wiper

The kinematic scheme of the parallelogram mechanism (fig. 3) is presented by means of a continuous line in an intermediate position (when the door side MN is outside the bus), and by means of a dotted line in the extreme closing (right) and opening (left) positions.

The mechanism is pneumatically driven by means of an actuator situated below, under the bus platform, at the same level with the stairs used by passengers to go on and off the bus.

2.2. The screen wiper mechanism

The parallelogram mechanism is used as screen wiper (fig. 4) for urban buses, where the wiping shim is fixed in the MN position perpendicular to the reciprocating rod AB in a point situated to the right of the AB segment.

The kinematic scheme of the parallelogram mechanism (fig. 4) is shown in three distinct positions, of which the extreme position to the right is represented by a continuous line, and the other two positions (middle and left) are represented by means of a dotted line.

This mechanism is usually electrically driven by means of an equalizing rod – crank mechanism.

This type of parallelogram mechanism with a fixed small edge is used for the doors of motor vehicles in terms of a crane to lift / lower the glass.

The reciprocating rod of such a parallelogram does a circular translation motion, and by means of some rollers, gets into contact with the glass frame, which is vertically guided.

A

A0 B0

B M N

M' N''

A' B' N'

M''

B A

B0 A0

N

M M'

M''

N' N''

B' A' A'' B''


86

This kind of mechanism is driven manually or electrically by means of a cylindrical gear having a large multiplication ration, placed inside or outside, which ensures a certain self-locking degree with regard to the intention of actuating from the led element to the leading one.

3. GEOMETRY AND KINEMATICS OF THE FERGUSSON PLANAR MECHANISM The Fergusson planar mechanism (fig. 2, 5) consists of three geared wheels to the

outside (1, 2, 3) placed in a series, articulated in points O, A and B on the planet wheel carrier p, which does a translation motion around the fixed point O. Providing that the distances between the axes of the wheels 1 and 2, respectively 2 and 3 are equal (OA = AB), the gear wheels 1 and 3 have equal diameters, and the same number of teeth (z1 = z3). If gear wheel 1 is fixed by means of blocking (fig. 2, 5), then point C becomes the

instantaneous rotation centre of gear wheel 2, with a null velocity (Cc=0).

a) b)

Fig. 5. Kinematic scheme (a) and general kinematic scheme (b) of Fergusson planar mechanism in a double orthogonal projection

For one rotation of the planet wheel carrier p at an angular velocity , points A and B have velocities in the ratio OA/OB = 1/ 2. The distribution of velocities (fig. 2) on the gear wheel 2 is obtained by uniting the point a (the peak of the A point velocity) with point C whose velocity is null.

Point d, peak of the velocity of point D, is located on the extension of the ac segment, so that the ratio of the velocities of points A and D is CA/CD = 1/ 2. For the velocities of points A, B, C and D we can define the ratios (fig. 2, 5):

2

1OB

OA

Bb

Aa;

2

1CD

CA

Dd

Aa (3.1,2)

Of the ratios (2.1) and (2.2) it results DdBb , that is the velocities of points B and D on the gear wheel 3 are equal, which determines 03 , so gear wheel 3 does a circular

translation motion.

A

C

A

B

O

D

C

1(0)

3

2 p

B

O

D

1

3

2

p

1(0) O

A A

B B

C C

D D

b

a

O

1

1(0)

2’ 2’

3 3 p

(2)

(3)

(1)

p

2

2 c

1(0) 0


87

The same conclusion may be obtained by means of the analytic method, writing the relative transmittal ratios provided the planet wheel carrier p does not move:

)1( 2

1

221

PP

Ppi

;

1

1

2

3

2

332

P

P

P

Ppi

(3.3,4)

Multiplying relations (3.3) and (3.4) we obtain

3

13221

3 11z

zii pp

p

(3.5)

For 31 zz and 0p it results from (3.5) that 03 , which corresponds to the

circular translation motion. 4. NEW CYLINDRICAL PLANAR MECHANISMS FOR GENERATING CIRCULAR TRANSLATION MOTION

A new kinematic scheme (fig. 5b) is obtained starting from the Fergusson planar mechanism (fig. 5a), where the interim gear wheel 2 has been replaced by two joint and several gear wheels 2 and 2’ with different radii.

Let us consider the formula (3.5) which is now written as

32

'21'3221

3 11zz

zzii pp

p

(4.1)

Provided that gear wheel 3 does a circular translation motion, that is 03 , we obtain

from formulae (4.1):

132

'21 zz

zz (4.2)

Thus, if we impose the transmittal ratio 0i of the exterior cylindrical gear (1,2), from

(4.2) we obtain:

10'2

3

1

2 iz

z

z

z (4.3)

Relation (4.2) can be checked by means of the graphical method of the linear velocities distribution (fig. 6), where, if the velocities of two points on gear wheel 3 are equal, for example DB VV , the instantaneous rotation centre is to infinity, that is the angular velocity

3 is null. Considering the distribution of linear velocities (fig. 6) we notice that the

equivalence DdBb implies the relations

AB

OA

AD

AC or 3'2

21

'2

2

rr

rr

r

r

(4.4)

Considering the second relation of (4.4), we obtain 32'21 rrrr , equivalent to (4.2), as

the radii of the dividing circles are proportional to the numbers of teeth.


88

A smaller gauge alternative of the kinematic scheme presented above (fig. 5b) is obtained if the gear wheel 3 gears with the gear wheel 2’ on the same side of the 2 axis with the exterior gear mechanism (1,2) (fig. 6).

Fig. 6. Kinematic scheme of the Fergusson mechanism – compact alternative

The position of the 3 axis, correspondingly of point B , is obtained by means of the

distribution of velocities Oa and Ca , depending on the position of point D . Thus, knowing the vector Dd with Cad , the position of B is given by the peak of its velocity where the parallel line from d to OA meets Oa .

In this case (fig. 7), following the distribution of linear velocities in the hypothesis DdBb , that is 03 , the geometrical construction implies

AD

CD

AB

OB or '2

'22

3'2

3'221

r

rr

rr

rrrr

(4.5)

which leads to the relation 32'21 rrrr , the same as with the previous alternative, namely (4.2),

32'21 zzzz . For the particular case when 0OD it results (fig. 7a) ACOC , and points A

and B are antipodal [1, 3], which allows for better balancing.

Fig. 7a. Kinematic scheme of the planar Fig. 7b. Kinematic scheme of the planar mechanism mechanism, improved version with interior gears

O

O

C

A

B

D D

A

C,c

B

a

d

b

2

1

2 2’

3

p

3

1(0)

2

2’

3

p

1(0)

A A

B B

D O,D

C,c

a

d

b

C

1

2

2’

3

p

1(0)

O

d

b

A A

B B

C,c

D D C

O O

a

1

2 2’

3

p

1(0)


89

Fig. 8. Images foto of mechanism model with external cylindrical gearings

Coming back to formula (4.1), for 03 we can write the general expression

1'3221 po ii (4.6)

Analysing the relation (4.6) it results that the transmittal ratios pi21 and pi '32 must have

the same sign, that is the cylindrical gears (1, 2) and (2’, 3) are either both exterior (fig. 7a) or both interior (fig. 7b).

If both gearings are interior (fig. 7b), we shall start by conveniently choosing the characteristic points CAO ,, and D , respectively the numbers of teeth of the gear wheels 1, 2 and 2’.

We build the velocity of point A through the segment Aa oriented to the left (fig. 7b). Knowing that the velocity of point C is zero, the velocity of point D is obtained by means of the linear distribution of velocities, points ca, and d are collinear.

From point we draw a parallel line to the reference line OA till it crosses the extension of the Oa segment in point b . Thus we obtain on the drawing (fig. 7b) the OBsegment, positioning the mobile axis of the satellite wheel 3 with interior dents, which does a circular translation motion. 5. CONCLUSIONS

The curve line translation motion can be generated in the particular form of circular translation, by means of mono-mobile mechanisms with articulated links of the simple parallelogram type (with a fixed side) or by means of transmissions with a toothed belt with a fixed wheel.

Also, circular translation can be generated through planar mechanisms with two cylindrical gears with a fixed central wheel. It is mentioned that the two cylindrical gearings of the Fergusson mechanisms are both exterior and interior.


90

The general form of the curve line translation is obtained by means of planar bi-mobile manipulating mechanisms with articulated links, auxiliary kinematic chains mounted in parallel with the main kinematic chain.

REFERENCES 1. Antonescu, P., Mechanisms –Structural and Kinematic Calculation, U.P.B. Printing Press, Bucharest, 1979; 2. Maros, D., Theory of Mechanisms and Machines – Gear Kinematics, Tech. Publ. House, Bucharest, 1958; 3. Antonescu, P., Antonescu, E., Synthesis of Cylindrical Planetary Mechanisms in view of the Circular Translation, SYROM’81 Bucharest, Vol. III, pp 9-14, 1981; 4. Antonescu, P., Antonescu, O., Mechanisms and the Dynamics of Machines, Printech Publishing House, Bucharest, 2005.


91

MAIN TYPES OF MECHANISMS USED AS WINDSHIELD WIPER Ph.D. Ovidiu ANTONESCU, „Politehnica” University of Bucharest: [email protected] Viorica VELIȘCU, Railroad Transportation Highschool, Craiova: [email protected] Constantin BREZEANU, Silcotub S. A. (Tenaris Group) Călărași: [email protected]

Abstract: This paper deals with the main types of windshield wiper mechanisms, consisting of bars and gear links. These mechanisms are classified into two groups: those with constant length arm and those with variable length arm. For each type of mechanism used as a windshield wiper, the paper highlights the main advantages and disadvantages regarding the wiped windshield surface and the constructive structural complexity.

Keywords: windshield wiper mechanism, bar, gear, degree of mobility, kinematic diagram 1. GENERAL CONSIDERATIONS

Windshield wipers are plane or spatial mechanisms, having one, two or three driven elements, in terms of wiping arms, with elastic blades on which the rubber blades that wipe the windshield glass or the rear window glass are attached [2 - 7].

Electric motors of continuous current are used to drive the windshield wipers, which are power by the car battery. This electric drive ensures the oscillation of the wiping arm within certain imposed frequency limits. [2]. With windshield wipers with a single constant length arm, the wiped surface is smaller than with those provided with two arms that are more often used [5]. Modern cars have telescopic arm windshield wipers, which ensures maximum coverage of the windshield surface [4, 8, 9, 10].

Windshield wipers are designed and built as mechanisms with articulated bars, cams and gears. [7]. The paper analyses the main types of mechanisms used as windshield wipers (ws. w.) from a structural and topological point of view.

2. WS. W. MECHANISMS WITH ARTICULATED BARS

This type of ws. w. mechanism with articulated bars (fig. 1) is used for the majority of road vehicles. It has the advantage of being a simple mechanical structure, which is safe during operation (at several operating speeds), as well as very reliable.

Each of the two arms b1 and b2 are provided at the upper edge with an elastic blade (L1, L2). These arms are connected to the equalizing bars 3 and 5, which rotate around the fixed points B0 and C0.

The two equalizing bars (3, 5) are driven by the same crank 1 (articulated in A0 at the

bottom), by means of the reciprocating rods 2 and 4, articulated in the same point A at the crank. The degree of mobility of the ws. w. mechanism is determined by formula [11]

6

2

5

1 rr

mm rNmCM (1)


92

a) b)

Fig. 1. Kinematic (a) and structural (b) diagrams of the ws. w. mech. provided with two parallel arms

In formula (1) Cm stands for the number of m operational class kinematic couplings, and Nr stands for the number of r rank independent closed contours. These structural-topological parameters are highlighted in the matrix

00020

00007

65432

54321

NNNNN

CCCCC (2)

Having these numerical values, formula (1) is written: 12371 M (3) Which shows that the analysed ws. w. mechanism (fig. 1) has a single degree of mobility; this consists in the movement of the leading element, that is crank 1.

In fact, the 3 fixed articulations in A0, B0, C0 do not have rigurously parallel axes. That is the reason why the kinematic couplings in points A, B, C are spherical articulations.

The structural-topological matrix of the actual mechanism is written according to (2) as

20000

00403

65432

54321

NNNNN

CCCCC (4)

And the degree of mobility is inferred from (1): 326)4331( M (5) The two additional mobilities are represented by the rotations of bars 2 and 4 as to the axes AB and CD. If we consider the mechanism in terms of a plane mechanism (fig. 1a), the structural-

topological diagram (fig. 1b) shows that it results from the Watt kinematic chain. With this plane mechanism we identify two open kinematic chains of the dyad type (2, 3) and (4, 5), paralleled to the fundamental mechanism MF (0, 1), which corresponds to the aggregation formula of the motor mechanism (MM):

)5,4(

)3,2()1,0(

LcD

LcDMFMM (6)

Even though this kinematic diagram (fig. 1) is simple, the solution has the disadvantage that the surface wiped on the windshield (by the rotating oscillating arm) is nevertheless limited. For ws. w. with a single rotating oscillating arm (usually used for the rear window), longer wiping blades are mounted.

0

B 2 A

4 C

b1 b2

L1 L2

1 5 3

B0 C0

A0

B

B

3

2 A A’ C 4

C0

5 A0

1

0


93

Another kinematic diagram of ws. w. mechanisms [1, 7] uses a complex topological structure (fig. 2), where the kinematic chains are connected serially.

a) b)

Fig. 2. Kinematic (a) and structural (b) diagrams of ws. s. M. with two serial arms

We can notice a complex topological structure, where 3 and 4 bars (criss-crossing each other) form together with the kinematic elements 2 and 5, a closed kinematic contour of the quadrangle type (BCEDB). The wiping arms (connected to bars 5 and 7) are serially linked by means of bar 6, and the quadrangle F0FGG0F0 is an articulated parallelogram. In order to determine the degree of mobility of this ws. w. M. (fig. 2a), we write the structural matrix:

00030

000010

65432

54321

NNNNN

CCCCC (7)

The degree of mobility is obtained from formula (1): 133101 M (8) The structural-topological diagram of this type of ws. w. m. (fig. 2b) corresponds to a complex kinematic chain, where we can distinguish two open kinematic chains with zero mobility, which are identified in the order of a possible disaggregation: LcD(6, 7) and Lc tetrade type LcTt(2, 3, 4, 5). The structural-topological formula of the complex ws. w. m. (fig. 4) can be written after stating the driving element, that is crank 1: )7,6()5,4,3,2()1,0( LcDLcTtMFMM (9) 3. WS. W. MECHANISMS WITH BARS AND GEARS WITH A CONSTANT LENGTH ARM

Ws. w. m. with a constant length arm [3, 10] can be grouped into two variants, different from the construction point of view: those with a translational gear rack T (fig. 3) and and with a roto-translational gear rack R+T.

A0

G0

G

F0

F

E

C

D B

A 4

3 2

7 6

5 0

1

0 7 6

5 4 3

2 1

G0

G F

F0 D

E C

B

A A0


94

a) b)

Fig. 3. Kinematic diagram of the ws.w.m. with a T gear rack The Ws.W.M. with translational gear rack (fig. 3) is provided with a single wiping arm,

connected to bar 5; thus it has a plane roto-translational motion. From crank 1, as driving element, motion is transmitted by means of the reciprocating rod

2 to the B axis of the gear 3 (pinion type). Pinion 3 engages at the same time to the upper side with the fixed gear rack 0 and to the

lower side with a mobile translational gear rack 4. To the translational gear rack 4 the kinematic dyadic chain LcD (5, 6) is articulated in

point C, with the fixed articulation D0. The translational motion of point B along the fixed guide 0 is provided by the double

engagement of pinion 3 with the two gear racks. Thus, skid 7, ensuring the translational motion of point B (fig. 3b), together with the two

couplings (rotational and translantional) is not represented in the kinematic diagram (fig. 3a). The structural matrix of the bar and gears plane mechanism (fig. 3a) is completed taking

into account the two kinematic couplings of skid 7:

00040

00029

65432

54321

NNNNN

CCCCC (10)

If these numerical data (10) are replaced in (1), we determine the real mobility degree 1432291 M (11)

This result (M=1) verifies the existence of a single element 1 with an independent motion. Ws. W. M. with a roto-translational gear rack (fig. 4) has a single wiping arm connected to bar 5 with a plane-parallel motion.

0

0

6 5 4

3

2 D

D

C B

A 1

A0

3

0

4

0

B

7


95

a) b)

Fig. 4. Kinematic (a) and structural (b) diagrams of the ws.w. M. with a roto-translational gear rack The rotation motion of crank 1 (fig. 4) is transmitted to the gear sector 4 by means of the

bar – gear rack 2. This is guided in the oscillating box 3, which keeps it in contact with the gear sector 4:

The oscillating rotation motion of gear sector 4 is sent to an antiparallelogram mechanism made up of bars 4, 5 and 6, with fixed articulations in B0 and D0. The wiping blade MN is mounted and fixed on bar 5, perpendicular to CD, whose motion is plane roto-translation.

The antiparallelogram B0CDD0 mechanism (fig. 4a) is represented by a dotted line in the extreme left position, in which we notice the position of the wiping blade M’N’ rotated at 1800 as to the initial position MN. The structural matrix of the ws.w.M. includes the number of the kinematic couplings of functional classes Cm with m =[1,5] and the number of closed independent contours Nr of rank r =[2,6]:

00030

00018

65432

54321

NNNNN

CCCCC (12)

The mobility of the analysed mechanism is determined by means of the formula (1) in which we replace the data in (12): 1331281 M (13)

The structural-topological diagram of this mechanism (fig. 4b) is obtained after equating the upper coupling (represented by the engagement of gear rack 2 with the gear sector 4). This engagement equates with a binary element, with two lower kinematic couplings [3]. The structural-topological formula is: MM = MF(0,1) +LcD(2,3) +LcD(e24,4)+LcD(5,6) (14)

4. WS. W. MECHANISM WITH BARS AND GEARS OF A VARIABLE LENGTH ARM With these variable length arm ws.w. M. [4, 7, 10], the wiping blade does a plane roto-translational motion consisting in a rotation of an oscillating bar and a translational along this one on a radial direction (fig. 5a).

D

D

C

B’0

B’B’

B

B A

A

3 4 5

6

e24

2

1 0

4

6

5

1

D’C’

M’

M N

N’

D0

C

D

B

4

3 A

2

A0


96

a) b)

Fig. 5. Kinematic diagram of the ws. w. M. with a variable length arm

The wiping blade MN is fixed on the rod 5, and the latter does the translational motion along the oscillating guide 1’ connected to the gear 1, which has inside teeth.

Together with the oscillating rotation of the box-gear 1, motion is sent to crank 3 (by means of the gears 2 and 3), and from here, by means of bar 4, the rotation motion is changed into translational motion at rod 5 (fig. 5a).

The surface wiped by the blade MN corresponds to the stroke of rod 5, a relative translational motion along the oscillating guide 1’ .

The structural matrix of this ws. w. M. (fig. 5a) is

00030

00026

65432

54321

NNNNN

CCCCC (15)

From this numerical data the mobility of the ws. w. M. is inferred with the formula (1):

1332261 M (16) The structural-topological diagram (fig. 5b) is done after equating the two upper

kinematic couplings (1, 2) and (2, 3), which are inside and outside engagements. The structural-topological formula is written (fig. 5b): MM = MF(0,1) +LcD(2,e12) +LcD(e23,3)+LcD(4,5) (17)

6. WS. W. M. WITH THE WIPING BLADE IN A CIRCULAR TRANSLATIONAL MOTION

This kind of ws. w. M. is used for buses where the surface of the windshield screen is flat and much larger than it is for cars. The wiping blade MN is placed vertically, fixed to the reciprocating rod 4 of a plane mechanism of an articulated parallelogram type (fig. 6), in which the sides -bars 3 and 5 are much longer than horizontal sides. In the practical solution, bar 5 is thin as compared to bar 3, which receives the oscillating rotation motion from crank 1, by means of bar 2 (fig. 6a). The structural matrix of that mechanism contains:

N

M

1’

2 1

5

3

4 A B

5 4

e2e1

2

1 3 0

A

B


97

00020

00007

65432

54321

NNNNN

CCCCC (18)

Thus, the mobility degree of the Ws.W.M. (fig. 6a) is determined with formula (1) in which we introduce the data in (18): 12371 M (19)

The structural-topological diagram of the Ws.W.M. analysed above corresponds to the Watt kinematic chain (fig. 6b). The structural-topological formula is (fig. 12):

MM = MF (0, 1) + LcD (2, 3) + LcD (4, 5) (20) a) b)

Fig. 6. Kinematic (a) and structural diagrams (b) of the Ws. W. M. with the blade doing a R+T motion

7. CONCLUSIONS Ws. W. Mechanisms have a diverse topological structure, being made both as spatial

mechanisms and as plane mechanisms with bars and gears. The paper aimed at a classification of the Ws.W.M. according to: the kinematic

elements used (bars, gear racks, cylindrical gears with outside and inside teeth), the number of wiping blades, their motion and their serial or parallel connection. The Ws.W.M. is driven by means of continuous current electric motors, powered by the car battery.

The wiping arm is moved together with the elastic blade. It is a rotational motion in the case of bar ws.w.m., or a roto-translational motion for those with bars and gear elements.

Comparing the 6 kinematic diagrams analysed, we pointed out the complexity of the variable length arm Ws.W.M. from the structural and the construction point of view.

From the structural-topological analysis of the main types of Ws.W.M. used for cars, we find out that most of them consist of dyadic kinematic chains, created either of articulated bars or of rack and pinion or gear sectors, or outside / inside cylindrical gears.

REFERENCES 1. Autorenkolektif – Getriebetechnik Lehrbuch, VEB, Verlag Technik Berlin, 1969. 2. *** - Passengercar windshield wiper systems – SAE 5903 C, 1973.

3

D’

C’ C

D

D B

4

5

3

2 1

B’

A’

N’

A

M’

N

M

B A

D C

A

D0 B0

A0

5

4

2 1

0

3 B


98

3. *** - Patent U.S.A. No. 3721878, 1978. 4. Schüche, S. et al., - Windshield wiper arrangement, U.S. Patent No. 4447928, 1984. 5. Antonescu, P., Cocosilă, M., Tempea, I. – A Comparative Study of the Drive Mechanisms of Windshield Wipers used for Cars, MERO’87, Vol. 4, pp. 11-19, Bucharest, 1987. 6. Antonescu, P., Mitrache, M., Cocosila, M. – Contributions to the Synthesis of Mechanisms used as Windshield Wipers, SYROM´89, Vol. IV, pp. 23-32, Bucharest, 1989. 7. Antonescu, P., Tempea, I., Adîr, G. – Windshield Wipers Mechanisms for Passenger Cars., SYROM’89, Vol. IV, pp. 41-50, Bucharest, 1989. 8. Antonescu, P., Crişan, M., Antonescu, D. – A Synthesis of the Plane Mechanisms used for Driving Windshield Wipers, ESFA’ 92, Vol. II, pp. 367-374, Bucharest, 1992. 9. Antonescu, P., Cocosila, M., Antonescu, D. – The Geometrical and Building Structure of Windshield Wipers Mechanisms, ESFA’92, Vol. II, pp. 375-386, Bucharest, 1992. 10. Antonescu, P., Cocosila, M., Antonescu, O. – The Geometrical Structure of Windshield Wipers Mechanisms, SYROM’93, Vol. IV, pp.125-132, Bucharest, 1993. 11. Antonescu, P., Antonescu, O. – Mecanisme și dinamica mașinilor, Editura Printech, 2005.


99

TOPOLOGICAL STRUCTURE OF CONNECTING MECHANISMS IN THE ELECTRIC GRID

Constantin BREZEANU, Silcotub S. A. (grup Tenaris) Călărași, [email protected] Ioana POPESCU, Iuliu Maniu Highschool of Bucharest, [email protected]

Dr. Păun ANTONESCU, Politehnica University of Bucharest, [email protected]

Abstract: The paper presents the main types of mechanisms used within the connecting systems from the high- and low-voltage electric grid. The purpose is the accurate construction of the kinematic diagrams of the mechanisms of electrical connection. For the low voltage connecting systems the topological structure of three kinematic schemes of articulated plane mechanisms is analysed. The structural-topological analysis is extended to other three kinematic schemes of simple plane mechanisms used as high voltage connecting systems. The structural-topological study is then applied to the complex plane mechanisms used as high voltage separators.

Keywords: topological structure, kinematic scheme, mobility, connecting mechanism, electric grid. 1. ARTICULATED PLANAR MECHANISMS USED FOR LOW VOLTAGE CONNECTING SYSTEMS Connecting systems normally use plane mechanisms with articulated bars [1,3,6,7], or in the simplest form of a single equalizing bar articulated frame (fig. 1.1) or in the shape of an articulated quadrangle (fig. 1.2, 1.3).

Fig. 1.1. Balanced mech. B; Fig. 1.2. Quadrilateral mechanism; Fig. 1.3. Quadrilateral mech. B-B

The equalizing bar mechanism (fig. 1.1) shows the a1 arc as the resistance force, opposing

to the electromagnetic force Fm of EM, and it is one of the oldest solutions of electromagnetic relay switch [20] for small nominal currents. The quadrangle mechanism (fig. 1.2) is mechanically driven by the coupling Mm, by means of the crank 1, till the bars 1 and 2 are placed one continuing the other, which corresponds to the extreme position of the equalizing bar 3, when connection C is made.

The quadrangle type mechanism (fig. 1.3) is driven by the electromagnet EM till the bars 1 and 2 are one continuing the other, and the connection in point C is obtained in the extreme position of the equalizing bar 3. The arc a1 acts as the resistance force and can open

1

A0

A

B0

B

0

EM

a1

Fm

1

A0

A

B0

B

0

C

2 3

Mm

A0

B0

A

B

C

D D0

1 2

3

a1

0

EM

Fm


100

connection C if electric power is no longer supplied into the EM. The mobility of the equalizing bar mechanism (fig. 1.1) shall be determined with the

formula for plane mechanisms [5]: 453 23 CCnM (1)

From the kinematic scheme the following values stand out: 0,1,1 45 CCn .

Replacing these numbers in the formula (1) we obtain: 1012133 M

The mobility corresponds to the rotation motion of the bar 1 (the equalizing bar) around the axis of the fixed articulation A0. The connection in point A results from the rotation of bar 1, which can be obtained with the attraction driving force Fm. Interrupting the power supply into the EM results in breaking the connection in A assisted by the arc in tension a1.

The mobility of the crank – equalizing bar mechanism (fig. 1.2) or of the equalizing bar - equalizing bar mechanism (fig. 1.3) is determined with the formula (1), where the numerical values are introduced: 0,4,3 45 CCn .

The following mobility results from replacement: 1042333 M

The only independent motion is the rotation of bar 1 by means of the driving torque Mm (fig. 1.2) or by means of the driving force Fm (fig. 1.3), which leads to the connection in C. 2. SIMPLE PLANAR MECHANISMS USED FOR HIGH VOLTAGE CONNECTING SYSTEMS

We consider the kinematic scheme (fig. 2.1) related to the mechanism of a low oil power switch, for really high voltage with breaking arcs [1,3].

Fig. 2.1. Quadrangle mechanism. Fig. 2.2. Slipper mechanism. Fig. 2.3. Roller mechanism.

The mobile connection 3 has a rotating motion in an upper horizontal plane, and a part of it, DE, gets into the fixed connection. The articulated quadrangle B0BCC0 (fig. 2.1) receives the motion in the A0A arm that is rotating in a lower horizontal plane of force Fm. Disconnection is obtained by means of the arc a1 that is tensioned in the D’E’ position.

0

A0

A

B0

C

B

1

2

3

3

Mm Fm

1

D’

A0

B0

B

a1

C0

D0

A

D

1

1 2 3

3

E

E’

0

0

A’

A0’

C

A

B

B0

C’ 1 2

3

4 5 0

A0

C0

Fm

C

B’


101

The mobility of the mechanism is determined by means of the formula (1) in which we introduce the structural numerical parameters: 0,4,3 45 CCn . 104233 M .

This value verifies the unambiguous determined motion of the mechanism with a single leading element 1. With regard to the slipper mechanism (fig. 2.2), guiding in the oscillating crank lever 3, the topological structure is the same as for the previous mechanism (fig. 2.1).

The difference consists in the existence of a translational coupling between the slipper 2 and the crank lever 3. The mechanism is actuated by means of the Mm coupling, and the connection position C must be reached in the extreme position of the bar 3, which corresponds to the angle (A0AB0) of 900.

The roller mechanism (fig. 2.3) is driven by a translational piston 1 actuated by the force Fm of the compressed air. We should notice that the rollers 2 and 4 (articulated by the equalizing bar 3) are connected to the upper part with the corresponding guide on the rod 1 respectively on the oscillating crank lever 5.

The mobility of the mechanism is determined by means of the formula (1) where we introduce the numerical values identified on the kinematic diagram (fig. 2.3):

2,5,5 45 CCn . Thus, we obtain for the mobility the value 325253 M .

One of the three mobilities corresponds to the translational motion of the leading piston 1. The other two mobilities are represented by the independent rotary motions of the rollers 2 and 4. The connection in point C is obtaine din the left position of the piston 1, where the angle B0BC0 is 900, or in the right position of piston 1, where the angle B0B’C0 is equal to 900. The position of the connecting point C changes, and it can be placed in point C’ in the right part of the figure (fig. 2.3).

3. COMPLEX PLANAR MECHANISMS USED AS AUTOMATED PNEUMATIC SWITCHES

These planar mechanisms with a complex structure are used as separators (switches) for three-phase high voltage power lines [1,2,3]. We consider the kinematic scheme (fig. 3.1) of an automated switch of 35 KW [3], which is pneumatically driven by means of a double piston with a rack bar, or by a roller guided in a crank lever.

From piston 1, driven by force Fm, by means of the rack bar gear (1) – geared sector (2), motion is transmitted to the articulated quadrangle (2, 3, 4). Thus, by means of the reciprocating rod 3, the rotation of bar 2 is transmitted to the equalizing bar 4, through the articulation D.

Following the kinematic chain, from the equalizing bar 4, through the reciprocating rod 5 with the articulations E and F, motion is transmitted to the translating rod 6 at the end of which there is the mobile connecting point K1 of the first phase of electric power.


102

Fig. 3.1. Kinematic scheme of the three-phase connecting complex mechanism

Together with reaching the final position of rod 6, we obtain the synchronous displacement of rods 10 and 14, corresponding to the mobile connecting points K2 and K3 of the other two phases of the high voltage electric power.

From the equalizing bar 4, through the double articulation G, motion is transmitted to the articulated bars 7 and 11, and then to the equalizing bars 8 and 12.

On the way to the mobile connecting point K2 we identify the articulated quadrangle D0GG’D’0 or through the component elements (0,4,7,8). Also, the kinematic way to the mobile connecting point K3 contains the anti-quadrangle D0GG’’D0’’ (0, 4, 11, 12). The equalizing bar 12 is linked to a buffer made up of the kinematic elements 15 (piston rod) and 16 (cylinder). On the kinematic scheme (fig. 3.1) we identify the following numerical values of the parameters in the formula (1): 1,23,16 45 CCn ; 11232163 M .

Mobility shows that the mechanism can be driven by a single leading element, piston 1. The motion flow can be traced by means of the structural – topological formula of the

drive mechanism motor MM for each contact K1, K2 and K3, starting from the actuator mechanism MA(0,1). Thus, the structural – topological formula for the contact K1 is

)6,5()4,3()2,()1,0( 12 LDLDeLDMAMM (2) In the second phase of the contact K2, the structural – topological formula becomes

)10.9()8,7()4,3()2,()1,0( 12 LDLDLDeLDMAMM (3) For the third contact K3, the structural – topological formula is written

)16,15()14.13()12,11()4,3()2,()1,0( 12 LDLDLDLDeLDMAMM (4) In formulas (2, 3, 4) we noted as e12 the imaginary kinematic element equivalent to the

superior kinematic joint made by the gear formed of the rack 1 and the geared sector 2. We should mention that the dyadic chain LD(15,16) of the RTR type is a sort of hydraulic buffer.

A Fm

B0

C

D

D0

E

F

G

D’0

G’ E’

F’ F’’

E’’

G’’

D0’’ H

I

I0

1 2

3

4

5

6

0

7

8

9

10

11

12

13

14

0 0

15

16

0

B

2

K1 K2 K3


103

4. MECHANISM OF THE HIGH VOLTAGE THREE-POLE SEPARATOR We take into consideration the kinematic scheme (fig. 4.1) of the mechanism of a

separator for voltages higher than 60 kV. Actuation of the contact bars in points K1, K2 and K3 is carried out by means of the

mechanism of a quadrangle of the equalizing bar – lever type (A0ABB0) through lever 1. The three contacts are obtained in the extreme position of the equalizing bar 3, when lever

1 and the reciprocating rod 2 are one following the other. From the equalizing bar 3, through the articulation C, the rotary motion of the former

(clockwise) is transmitted through the reciprocating rod 4 to the equalizing bar 5 that is rotating, trigonometrically, until the bars b3 and b5 reach a vertical position (in contact K2).

From the equalizing bar 3 the motion is transmitted, to the left and to the right, by means of articulated quadrangles to the equalizing bars 3’ and 3’’ with the fixed articulations B0’ and B0’’. Between the upper and lower axes of the fixed articulation B0 and D0 respectively B0’, D0’ and B0’’, D0’’ we mount insulators. Following the kinematic scheme of the separator mechanism (fig. 4.1) we infer the numerical values 0,23,15 45 CCn that we introduce in

the formula (1), resulting in 10232153 M This result corresponds to a rigid structure, so that the mechanism should be an

undetermined static system. In reality, the mechanism operates on the basis of only one leading element 1, and the result above is due to the double link between the equalizing bars 3, 3’ respectively 3 and 3’’. Thus, the reciprocating rods 7 and 7’ are mounted parallel to the reciprocating rods 6 and 6’, which does not introduce additional geometric conditions.

Fig. 4.1. Kinematic scheme of the three-pole separator mechanism From a geometrical point of view, the mechanism can operate without bars 7 and 7’, case

in which the structural parameters are: 0,19,13 45 CCn . Introducing the numerical

values in the formula (1) we obtain 10192133 M

2

1

3

3

4’

4

5

3’ 3’’

A0

A

E

B0

C

D

D0

4’ 4’’

4’’ 4’

C’ C’’ 6 6’

B0’’ B0

’

E’ E’’

7 7’

5’ 5’’ D’ D’’

D0’ D0’’

K1 K2 K3

F F’ F’’

0 0 0

0 0 0

B

b3

b5


104

Indeed the mechanism transmits the unambiguous determined motion from the leading element 1 to the led elements 3, 5 respectively 3’, 5’ and 3’’, 5’’. The structural-topological formula of the drive mechanism is

)''5,''4()''3,'6()'5,'4()'3,6()5,4()3,2()1,0( LDLDLDLDLDLDMAMM (5) 5. THE MECHANISM OF THE SINGLE-COLUMN SEPARATOR (WITH ROLLER AND CRANK LEVERS)

The kinematic scheme of the mechanism (fig. 5.1) shows that the drive uses a pneumatic actuator p with a double piston 1 [3].

a) b) c)

Fig. 5.1. Kinematic scheme of the single-column mechanism

We notice that the mechanism with a symmetrical structure has two rollers 2 and 4 guided in the corresponding crank levers 3 and 5. Is identified 5n mobile kinematic elements; 55 C class kinematic couplings

(mono-mobile) out of which a translational coupling A(0,1) and 4 rotary couplings B(1,2), C(1,4), B0(3,0) and C0(5,0); 24 C are 4 class kinematic couplings (bi-mobile) of plane rotary translation. Introducing these numerical values in the formula (1) we obtain

21325253 M We should notice that two of the three independent motions of the mechanism are

passive mobilities, represented by the rotation of each of the rollers 2 and 4 around their centre.The available independent motion is the translational motion of the double piston 1 in the fixed pneumatic cylinder, vertically mounted.

The structural – topological formula of the drive mechanism analysed above is )5,4()3,2()1,0( LDLDMAMM (6) The kinematic scheme in the position of closed contact K (fig. 5.1a) corresponds to the

up-and-down displacement of piston 1, and the separating position of the two bars b3 and b5 (fig. 5.1b) corresponds to the up-and-down displacement of piston 1.

2 3 5

4 1

0

0

A

B

B0 C0

C

K

3 5

2

3 5

4 1

0

0

A

B

B0 C0

C

3 5

2 3 5

4 1

0

0

A

B

B0 C0

C

K

3 5


105

By replacing the plane upper kinematic couplings (2,3) and (4,5) with one equivalent element e23(2) respectively e45(4), we obtain the equivalent kinematic diagram (fig. 5.1c).

In this equivalent kinematic diagram all the kinematic couplings are class 5, or translational (0,1), (2,3), (4,5), or rotary (1,2), (3,0) and (1,4), (5,0).

We should mention that, for the constructive diagram of the pneumatic separating mechanism, we shall provide locking bolts for the two equalizing bars 3 and 5 especially due to the weight of the double piston 1. 6. CONCLUSIONS

The mechanisms related to electric switch systems are plane mechanisms with articulated bars, having a simple topological structure in the case of low electric voltages. The mobility of these mechanisms is usually carried out by means of electromagnets.

For high electric voltages, the plane mechanisms used are based on kinematic diagrams with single contour lines articulated bars, driven by spiral arcs.

For three-phase power lines, the mechanisms used are carried out as complex plane structures with parallel kinematic chains. These mechanisms are pneumatically driven, and are provided with a pneumatic buffer for one of the three phases.

We carried out an equivalent kinematic scheme, both for the pneumatic actuator and for the parallel final kinematic chains. The structural topological analysis of the mechanism of a three-phase separator shows that the kinematic diagram uses serial quadrangle mechanisms.

REFERENCES 1. Hortopan, G., Electric Devices, The Didactic and Ped. Publ. House, Bucharest, 1972; 2. Hortopan, G., Low Voltage Electric Devices, The Technical Publ. House, Bucharest, 1969; 3. Macsymiuk, J., Mechanisms of Connecting Electric Devices, The Technical Publ. House, Bucharest, 1970. 4. Antonescu, P., Mechanisms, Printech Publishing House, Bucharest, 2003; 5. Antonescu, O., Antonescu, P., Mechanisms and Manipulators, Printech Publishing House,

Bucharest, 2007; 6. Nedela, N., Antonescu, O., Mechanisms used for medium voltage power switches, Journal

Mechanisms and Manipulators, Vol. 8, No 1, 2009, p. 43-50; 7. Nedela, N., Geonea, I., Mechanisms used for high voltage switching devices, Journal Mechanisms

and Manipulators, Vol. 9, No 1, 2010, p. 51-58.


106

CABLE MECHANISMS USED FOR ACTUATING CAR ELEVATORS WITH 2 AND 4 POLES

Dan MESARICI, Railroad Transportation High School Craiova, [email protected]. Viorica VELIȘCU, Railroad Transportation H. S. Craiova, [email protected]

Daniela ANTONESCU, "Iuliu Maniu" H. S. Bucharest, [email protected]

Abstract: This paper presents the kinematic scheme and the operating mode of the cable mechanism used for car elevators with 2 and 4 poles. In order to increase the efficiency of the 4-pole elevator, we suggest using a new hoists type cable mechanism, which can multiply the piston travel inside the actuating cylinder twice at the platform level, when lifting and lowering.

Keywords: cable mechanism, car elevator, actuating cylinder, platform

1. GENERAL CONSIDERATIONS A 2-pole elevator (fig. 1) has on each pole two safety bracing plates [1-5], oscillating in a

horizontal plane (fig. 2a). For the upper position (fig. 1a) a safety level to the ceiling is provided [6-9]. In the lower position the minimum necessary distances are pointed out, both to the right and to the left of the elevator (fig. 1b).

a) Fig. 1. Building scheme of the elevator with two cable-actuated poles [8]

a) b)

Fig. 2. Building scheme of the elevator (a) and kinematic scheme (b) of the cable mechanism

b)




107

The building scheme of the car elevator with 4 poles is presented, as a photo image, perspective view (fig. 3, 4) and then the kinematic scheme is presented in a horizontal plane view (fig. 5a) as well as in side plane view (fig. 5b).

Fig. 3. Photo image of a car elevator with 4 poles – upper position [10]

Fig. 4. Photo image of a car elevator with 4 poles (simultaneously actuated by means of a cable),

in the lower position (on the ground)


108

The 4-pole elevator uses a cable mechanism for rising / lowering and it is only suitable for indoor operation (fig. 3). Such an elevator shall not be mounted on asphalt or on other such unstable surfaces. Each elevator pole is supported only by means of anchorage / fixed installation on the floor.

The space left to the ceiling must be checked first in order to determine the height of the elevator that can be placed inside. Mounting such a 4-pole car elevator is relatively easy and can be done in a few hours.

2. THE CABLE MECHANISM WITH A PNEUMATIC ACTUATING CYLINDER

We shall consider the kinematic scheme of the cable mechanism in two orthogonal projections, a horizontal one (fig. 3a) and another on right side (fig. 3b).

a) b)

Fig. 5. Kinematic scheme of the cable mechanism (option 1)

The cross distance between two poles (left and right) is about 2,800 mm, while the longitudinal distance between two poles (front and back) is 5,000 mm.

The width of the two tracks (rolling tracks) is about 500 mm, and the distance between these tracks is 1,100 mm. The power unit is mounted on one of the front poles (electric motor + pump + compressed air tank).

Of the two tracks (rolling tracks), only one is provided with a cable spatial mechanism, actuated by a pneumatic actuator (cylinder) (fig. 5a).

The mechanism actuates four different lengths cables, one for each pole, each cable passing over a horizontal roll and over a roll situated in a vertical plane (fig. 5b).

It should be noticed that, for option 1 (fig. 5), the piston travel inside the mobile cylinder is not used efficiently, as the platform moves (rises and goes down) at half the actuator travel.

a) b)


1 2

1 1

2

2 3 3 4

4

1,3 2,4

3

4

2

1 3 4

1 2 3 4


109

In a different option (fig. 6a), in which the hydraulic cylinder is fixed onto the cover of the rolling track, two cables pass over only a horizontal roll each, while the other two cables pass over two horizontal rolls each (one on the right safety bar and one on the left safety bar). 3. THE CABLE MECHANISM WITH A FIXED ACTUATING CYLINDER

In this option 3 (fig. 7), four cables have been connected to the piston rod, two symmetrical cables as to the rod axis (1 with 3 and 2 with 4). Corresponding to the four cables, there are four superimposed cable wheels in a horizontal plane (fig. 7b).

The cables connected to the right side poles (fig. 7a) are placed in a lower plane (fig. 5b), and they are shorter than the cables going to the left side poles.

There is also option 3 (fig. 7a) for the cable mechanism, in which we use only two superimposed cable rolls, placed in the area of the cross-bar of the right side poles (fig. 7b).

a) b)


For a better view, cables 1 and 3 respectively 2 and 4 have been represented as having different directions (fig. 7a). We should mention that the length of cable 1 is shorter than the length of cable 2.

In fact, these cables can have the same direction, which provides the rotation of rolls 1 and 3 respectively 2 and 4 in the same direction (fig. 7b).

In this case, cable 1 passes over (in a horizontal projection) cable 3, while cable 2 passes over cable 4 (fig. 7a).

Therefore, cables 1 and 2 arrive, after they pass the corresponding rolls, in the positions given by the figures between brackets (fig. 7a).

Cables 3 and 4 remain unchanged, which shows that the length of cable 3 is shorter than the length of cable 4.

The rolls of cables 3 and 4 are mounted to the left in parallel planes, just like rolls 3 and 4 are mounted to the right side (fig. 7b).

The size of the cable varies according to the maximum weight of the vehicles lifted to a certain level for inspection.

It should also be noticed that for options 2 and 3, the piston travel inside the fixed cylinder is equal to the movement of the platform when rising.

2

4

4

2(1)

1(2) 3

3 4 3 1

1 2 3 4


110

When going down, the platform is moved due to the vehicle gravity force, till the actuating piston goes back to the initial position (zero travel).

4. A NEW KINEMATIC SCHEME FOR THE CABLE MECHANISM

To increase the efficiency of the 4-pole elevator, we suggest using a new cable mechanism of the hoists type (fig. 8), which can multiply the piston travel in the fixed cylinder twice (fig. 8a) or three times (fig. 8b) at the platform level, both when lifting and when going down.

a) b)

Fig. 8. Hoists type cable mechanism: simple (a) and double (b)

Thus, in the first option (fig. 8a), when point A moves (together with the mobile roll), point B (on the cable passing over the fixed axis roll) moves along a double distance:

2'

' AA

BB

In the second option of the hoists (fig. 8b) the ratio between the two rectilinear motions (out / in) equals 4:

4'

' AA

BB

Therefore, by means of using this hoists type cable mechanism, for the same movement of the 4-pole platform, the piston travel in the fixed cylinder is much shorter (half or a quarter), than for the current option of the installation (fig. 5).

With the new simple hoists type cable mechanism (fig. 8), for cables 1 and 3 respectively 2 and 4, two such cable systems overlap (fig. 9).

a) b)

Fig. 9. Kinematic scheme of the new simple hoists type cable mechanism [1, 2]

A

B

A

B'

A'

B

1 2 3 4

2

4

4

2

1

3 3

3 1

4 A1,3 A2,4 B1,3

B2,4


111

Considering the new kinematic scheme (fig. 9a), four simple hoists type cable mechanisms are identified, one for each car elevator pole.

According to the previous notations (fig. 8a), points A1,3 respectively A2,4 connect the piston rod to the mobile hoist roll, where cables 1 and 2 are shown by means of a continuous line, and cables 3 and 4 by means of an interrupted line.

It should be mentioned that, in the horizontal projection of the mechanism (fig. 9a), cables 1 and 3 respectively 2 and 4 overlap for a certain sector, between the mobile axis roll and the fixed axis roll (to the right side).

5. CONCLUSIONS

The cable mechanism of the 2-pole elevator consists of two vertical actuating cylinders, which are pneumatically driven at the same time (fig. 2b).

The cable mechanism used for the 4-pole elevator consists of a single actuating cylinder mounted on one of the two horizontal platforms (fig. 5a, 6a).

With the 4-pole elevator, the actuating cylinder is mobile, and the cable rolls are mounted two on each vertical axis.

If the actuating cylinder is fixed (options 2 and 3), building solutions are different as the cable rolls are mounted four (fig. 6b) and two (fig. 7b) on the same vertical axis. The new cable mechanism presented in this paper (fig. 9a) uses a fixed actuating cylinder in a horizontal position and has the cable rolls mounted two on the same vertical axis (fig. 9b). REFERENCES 1. Antonescu Păun, Mechanisms, Printech Publishing House, Bucharest, 2003 2. Dumitru Nicolae, Machine Parts. Gears. Design Elements. University of Craiova Press,

1996. 3. Antonescu Păun, Antonescu Ovidiu, Standardization of Terminology of MMS and

Graphical Symbols, Works of the Third National Workshop on Mechanisms, Craiova, 2008.

4. Antonescu Păun, Antonescu Ovidiu, Mihalache Daniela, Lifting Manipulators for a Green Environment. Proceedings of the Twelfth World Congress on Mechanism and Machine Science, Besanson, 2007.

5. Antonescu Ovidiu, Antonescu Păun, Mechanisms and Manipulators, Printech Publishing House, Bucharest -2006.

6. Jula Alexandru, Chişiu, Emil, Lateş, M., Mechanical Transmissions, Transylvania University Press, Braşov, 2006.

7. "Vehicle lifting points for frame engaging lifts”, ali/lp-guide. 8. www. H.T.C. Technology. 9. www .iscir. ro 10. www.heshbon.com; HL-25a instruction manual.

http://www.dmg-lib.org/dmglib/main/portal.jsp?mainNaviState=browsen.docum.meta&id=19826009




112

EXPERIMENTAL INVESTIGATION REGARDING THE STRESS VALUES FROM LATHE CUTTERS DURING THE MANUFACTURING

OF STEEL SHAFTS

1 Phd.Eng. Catălin ROŞU, S.C. DICO Romania S.R.L., Pieleşti, Cîrcea Street, No. 2, Romania, [email protected]

Abstract. In this paper these is studied the stress distribution from lathe cutters during the manufacturing of shafts made by normal steel (S355JR). The main purpose of this study is to determine, by experimental procedures, the stresses from lathe cutters during the turning process of steel shafts. The stresses were determined from the equivalent strain obtained from the strain gauges glued on the lathe cutters. The experimental setup and results are presented from an original point of view. Keywords: lathe cutter, turning, strain gauge, stress.

1. INTRODUCTION In this paper I will present the stress recordings from lathe cutters during the turning

process of steel shafts. The shafts are made from S355JOH according to SR EN 10219-1 (equivalent to OL 52.3K STAS 500). I have used the next chipping parameters: the feeding f= 0,2 mm/rot; the cutting depth ap= 0,5; 0,7; 0,9; 1,1; 1,3; 1,5 mm; the speciffic pressure pm= 310 MPa (according to Cârstea (2007) [1]). The speciffic pressure was chosen depending on the manufactured material: S355JOH from the tables in Cârstea (2007) [1]. The work presented in this paper is a continuation of the investigations made in Roşu (2014) [2]. Similar experimental montage was used in Roşu (2014) [2] for studying bronze materials.

As presented in Roşu (2014) [2], there will be used two different lathe cutters with the

next transversal sections: 25x25 mm (first lathe cutter) and 25x10 mm (second late cutter). The lathe cutters are presented in fig. 2 (the lathe cutter from fig. 2.a. will be named in the following part of the paper as variant 1 and the lathe cutter as variant 2). With 1 the active strain gauge was marked and with 2 the passive strain gauge was marked.

n

Steel product

Ф 20

30

f

Lathe cutter

Strain gauge

Fig. 1. Experimental montage used to determine the stress values from lathe cutters


113

The difference between the two lathe cutters is that, the one from variant 1 has the

cutting part made from hard-cutting alloy and the one from variant 2 has the cutting part made from high steel. In order to verify if the strain gauges were properly glued on the lathe cutters the UT-50 B meter (fig. 3). The continuity of the strain gauge resistance, for variant 1, is presented in fig. 4.

The continuity of the strain gauge resistance, for the variant 2, is presented in fig. 5.

The signal validation with the mass is presented in fig. 6. From fig. 4 and 5 we can see that the signal of the active strain gauges is around 120 Ω (which is according to the signal given by a proper glue of the strain gauges). Also, in order to check the connection, there must be made a validation with the mass (in order to see if one of the wires are in contact with the lathe cutter body). So, in fig. 6 we can see that there is a lack of signal (or infinite signal) with the mass which shows that the wires are correctly bonded to the strain gauge.

a. b.

Fig. 2. The used lathe cutters; a. varianta 1; b. variant 2

Fig. 3. UT-50 B meter Fig. 4. The continuity checking for the active strain gauge (variant 1)


114

Most of the experiments made in manufacturing investigate vibrations or residual

stresses (after drilling, especially). Some of these are presented in Rubio (2012) [3], Schajer (2013) [4], Schmitt (1996) [5], Mahmoudi (2009) [6] or Mahmoudi (2013) [7].

The strain gauges were glued to the lathe cutters according to the procedure presented in Miriţoiu (2012) [8], to obtain a half-bridge connection.

2. EXPERIMENTAL MONTAGE AND RECORDINGS The montage of the steel products in the lathe machine is made like in Roşu (2014)

[2]. Also, a detail with the lathe cutter (variant 1) montage in the slide rest is presented in fig. 8.

Fig. 5. The continuity checking for the active strain gauge (variant 2)

Fig. 6. The signal validation with the mass (variant 1)

Fig. 7. Details with the product montage in the lathe machine self-centring chuck

Fig. 8. Tool cutter montage in the slide rest


115

For the experimental recordings there was used the same apparatus from Roşu (2014)

[2]. Because of the experimental montage presented in fig. 2, from the equivalent strain the equivalent stress was directly obtained (by using a program made in the Test Point software). All the equivalent stress values (marked in an oval shape), for the variant 1, are presented in the figures bellow.

Fig. 9. Equivalent stress (variant 1, ap= 700μm)




116





117

All the stress values, obtained from the two variants, are written in table 1.

Table 1. Stress values versus the cutting depth

ap [mm] 0,5 0,7 0,9 1,1 1,3 1,5 σechexp [MPa] (varianta 1) 6,6895 9,3591 12,118 14,826 16,43 19,677

σechexp [MPa] (varianta 2) 38,785 56,143 73,365 90,629 108,37 123,67

3. CONCLUSIONS In this paper I have made an experimental investigation regarding the stress values

from lathe cutters during the turning process. The cutting depth was varied and the other chipping parameters were kept constant.

The added value of the current work is: - the equivalent stress values from different lathe cutters in the turning process of steel

products; - the used experimental montage (including the way for gluing the strain gauges to the

lathe cutters). The results obtained will be used in a future work to determine, by using the

regression analysis (according to the researches from Miriţoiu (2014) [9], Miriţoiu (2014) [10], Miriţoiu (2011) [11] or Zamfirache (2007) [12], direct calculus formulas for the equivalent stress.

REFERENCES

1. Cârstea, V., et. al., (2007) Course Support for Turning Processes, Sitech Publishing House, Craiova.

2. Roşu, C., Miriţoiu, C.M., Ilincioiu, D., Văduvoiu, Gh., Stanimir, Al., (2014) Experimental setup used to determine the stresses from lathe cutters, International congress Science and Management of Automotive and Transportation Engineering SMAT 2014, Craiova, 23-25 October, Romania, 219-225, Tome II

3. Rubio, E., (2012) Experimental analysis of the tool-to-workpiece vibrations transmission during chatter development in milling processes, Scientific Research and Essays, 7(39), 3286-3291.

4. Schajer, G., S., et. al., (2013) Hole-Drilling Residual Stress Measurement with Artifact Correction Using Full-Field DIC, Experimental Mechanics, 53, 255–265.

5. Schmitt, D., R., Li, Y., (1996) Three-dimensional Stress Relief Displacement Resulting from Drilling a Blind Hole in Acrylic, Experimental Mechanics, 36(4), 412-420.

6. Mahmoudi, A., H., s.a., (2009) A New Procedure to Measure Near Yield Residual Stresses Using the Deep Hole Drilling Technique, Experimental Mechanics, 49, 595–604, DOI 10.1007/s11340-008-9164-y.

7. Mahmoudi, A., H., s.a, (2013) A Procedure to Measure Biaxial Near Yield Residual Stresses Using the Deep Hole Drilling Technique, Experimental Mechanics, 1-9, DOI 10.1007/s11340-013-9729-2.

8. Miriţoiu, C.M., (2012) A Simple but Accurate Device and Method Used for Bending and Stress Measurement of Metallic Structures, IOSR Journal of Engineering, 2, 6, pp. 1334-1339.


118

9. Miriţoiu, C.M., (2014) Correlations between some mechanical characteristics and the

mass per unit length for some composite bars reinforced with steel fabric, Fiability and Durability, ISSN 1844-640X, 1, Editura Academica Brâncuşi, 112-119.

10. Miriţoiu, C.M., (2014) Some calculus relations for determining some of the mechanical properties of composite sandwich bars reinforced with metal fabric, Fiability and Durability, ISSN 1844-640X, 1, Editura Academica Brâncuşi, 119-126.

11. Miriţoiu, C., Ilincioiu, D., (2011) A Stress Analysis of a Metallic Structure using Finite Element Method, World Academy of Science, Engineering and Technology, International Conference of Automotive and Mechanical Engineering, Print ISSN 2010-376X, Electronic ISSN 2010-3778, Venice, Italy, 27-29 April.

12. Zamfirache, M., (2007) Research regarding the rougness of the surfaces titanium alloys grinding, Annals of the University of Craiova, Mechanical Series, 1, 63-68.


119

DIRECT CALCULUS FORMULAS FOR THE LATHE TOOL EQUIVALENT STRESS VALUES DURING THE MANUFACTURING

OF STEEL SHAFTS

1 Phd.Eng. Catălin ROŞU, S.C. DICO Romania S.R.L., Pieleşti, Cîrcea Street, No. 2, Romania, [email protected]

Abstract. In this paper, starting from the stress values presented in Roşu (2015) [1], a method for determining direct calculus formulas for the lathe tool equivalent stress values during the manufacturing of steel shafts is established. There is used the regression analysis for the formulas determination. The equivalent stress will be dependant on the cutting depth. The obtained formulas will be verified by using an analytical model from Strength of Materials. The results and the calculus formulas are presented from an original point of view. Keywords: regression analysis, equivalent stress, bending, turning.

1. INTRODUCTION In this paper, starting from the researches presented in Roşu (2015) [1] regarding the

stress determination from the lathe cutters that manufacture steel products through turning, I will establish a method to determine direct calculus formulas for the equivalent stress by using the regression analysis. I will also take into account the researches from Miriţoiu (2014) [2] and Miriţoiu (2014) [3]: there will be searched for linear, logarithmic, power and exponential functions to approximate the experimental results in order to obtain a correlation factor close to 1 value. The polynomial functions were avoided because their obtained factors insert errors if the experimental results are repeated (see Miriţoiu (2014) [2] and Miriţoiu (2014) [3] for these aspects). Also, some researches regarding the regression analysis applied in the manufacturing process can be found in Zamfirache (2009) [4] (titan alloys manufactured by turning) and Zamfirache (2009) [5] (stainless steels manufactured by grinding).

The stress results from Roşu (2015) [1] are presented in table 1.

Table 1. Stress values versus the cutting depth (Roşu (2015) [1]) ap [mm] 0,5 0,7 0,9 1,1 1,3 1,5



I have used the next chipping parameters: the feeding f= 0,2 mm/rot; the cutting depth

ap= 0,5; 0,7; 0,9; 1,1; 1,3; 1,5 mm; the speciffic pressure pm= 310 MPa (according to Cârstea (2007) [6]). There will be used two different lathe cutters with the next transversal sections: 25x25 mm (first lathe cutter) and 25x10 mm (second late cutter) (see Roşu (2015) [1] for details). The dependence of the stress and cutting depth, for both considered variants, are presented in fig. 1 and 2.


120

2. REGRESSION ANALYSIS USED TO DETERMINE DIRECT CALCULUS

FORMULAS In this paprt of the paper I will use the regression analysis to determine a direct

calculus formula for the equivalent stress experimentally obtained. I have searched for the next correlation functions:

- linear: 21 vv - logarithmic: 21 ln vv

- power: 21

vv

- exponential: vev 21

Tensiune-adancime de aschiere

0

5

10

15

20

25

0.5 0.7 0.9 1.1 1.3 1.5

ap[mm]

si[

MP

a]

siexp

Fig. 1. Stress vs. cutting depth (variant 1)

Tensiune-adancime de aschiere

0

20

40

60

80

100

120

140

0.5 0.7 0.9 1.1 1.3 1.5

ap[mm]

si[

MP

a]

siexp

Fig. 2. Stress vs. cutting depth (variant 2)


121

By using the regression analysis and the experimental data, I have obtained the values for α, β, η, λ, ξ, こ parameters. For every set, I have written the results in tables 2 (variant 1) and 3 (variant 2).

Important remark: the free length of the lathe cutter is 50 mm.

Table 2. The parameters α, β, η, λ, ξ, こ values – variant 1

Function type

Linear Logarithmic Power Exponential

Parameters α1= 2,5388

α2= 4,2974

β1= 6,9766 β2= 5,5332

η1= 6,4547

η2= 0,5955

λ1= 5,9853 λ2= 0,2081

Correlation factor R2

[%] 99,52 94,29 99,18 96,58

Table 3. The parameters α, β, η, λ, ξ, こ values – variant 2

Function type

Linear Logarithmic Power Exponential

Parameters α1= 17,096

α2= 21,99

β1= 46,821 β2= 30,486

η1= 37,228

η2= 0,6519

λ1= 34,249 λ2= 0,2281

Correlation factor R2

[%] 99,97 94,406 99,32 96,89

From the tables 2 and 3 it is seen that the linear function approximate the best the

experimental results. So, for the variant 1 I propose the formula (1) and for the variant (2) I propose the formula (2).

Important remark: because there was no distributed force to act upon the lathe cutter, it was expected to obtain, for the stress, a first degree function.

3. THE RESULTS VALIDATION In the next part of the paper I will validate the experimental results obtained in Roşu

(2015) [1], used in this study to obtain the direct calculus formula. If I consider that the lathe cutter is loaded only to bending in the vertical plane (according to the theory from Cârstea (2007) [6]), the relation (3) can be used for stress calculus.

22max

60

6

hb

afp

hb

lF

W

M pmz

yss

(3)

2974,45388,2exp ppech aa (1) 99,21096,17exp ppech aa (2)


122

If I take into account all the loadings (from axial force, shear force and bending moments in vertical and horizontal planes), then I will use the direct calculus formulas from Roşu (2013) [7].

2 8203

2 hb

hlblhbafp

W

lF

W

lF

A

Fpm

z

y

y

zxechB

(4)

222

10020196

1lhlh

hb

afp pmechB

(5)

222 641613

33

lblbbh

afp pmechB

(6)

The Bi points (i=1,2,3) are determined accordhing to the scheme from fig. 3 (presented in Roşu (2013) [7]).

In (3), (4), (5) and (6) I have marked with: - f the feeding; - ap the cutting depth; - pm the specific pressure; - b,h the base and the height of the lathe tool; - l the free length of the lathe tool. The obtained results for the variant 1 and 2 are written in tables 4 and 5.

f

F

Fz

Fy

Fx

x z

y

B1 B2

B3

1

2

Fz

Fy

Fx

Fig. 3. The loading of a lathe tool Roşu (2013) [7]


123

Table 4. Stress values (variant 1)

ap [mm]

0,5 0,7 0,9 1,1 1,3 1,5

σechB1 [MPa]

6,235 8,73 11,224 13,718 16,212 18,706

σechB2

[MPa] 8,482 11,874 15,267 18,66 22,052 25,445

σechB3

[MPa] 2,582 3,614 4,647 5,679 6,712 7,745

σss [MPa] 5,952 8,333 10,714 13,094 15,475 17,856

Table 5. Stress values (variant 2)

ap [mm]

0,5 0,7 0,9 1,1 1,3 1,5

σechB1 [MPa]

37,71 52,794 67,878 82,962 98,046 113,13

σechB2

[MPa] 43,524 60,934 78,343 95,753 113,162 130,572

σechB3

[MPa] 6,454 9,036 11,617 14,199 16,78 19,362

σss [MPa] 37,2 52,08 66,96 81,84 96,72 111,6

The stress values versus the cutting depth, for both variants are presented in fig. 4 and fig. 5

Valori tensiuni (varianta 1)

0

5

10

15

20

25

30

0.5 0.7 0.9 1.1 1.3 1.5

ap[mm]

si[

MP

a]

siechB1

siechB2

siechB3

siss

Fig. 4. Stress versus the cutting depth (variant 1) – analytical values


124

Valori tensuni (varianta 2)

0

20

40

60

80

100

120

140

0.5 0.7 0.9 1.1 1.3 1.5

ap[mm]

si[

MP

a]

siechB1

siechB2

siechB3

siss

Fig. 5. Stress values versus the cutting depth (variant 2) – analytical values

4. CONCLUSIONS From the fig. 1,2,4 and 5 and tables 1,4 and 5 we can see the next tendencies: - the stress increases with the cutting depth; - the stresses are higher if the second lathe tool is used; - the graphics variation is almost linear; - the maximum stress is in point B2 and minimum in point B3; - the stress determined from the simple bending in the vertical plane is almost equal

with the equivalent stress in point B1 (this shows that the shear stress given by shear force Fz has small influence over the equivalent stress);

- the σss values and the experimental stress values have small errors, under 12% (the errors may appear because of the vibration from the manufacturing process);

- the σechB1 values and the experimental stress values have small errors, under 8% (the errors may appear because of the vibration from the manufacturing process);

- the σechB2 values and the experimental stress values have high errors (maximum is around 35%) because the influence of the horizontal plane moment is high, and its value has not been determined in the experimental part;

One of the added values of this paper is also the determination of some direct calculus formulas for the equivalent stress, depending on the lathe tool type and on the cutting depth value.

REFERENCES 1. Roşu, C., (2015) Experimental investigation regarding the stress values from lathe

cutters during the manufacturing of steel shafts, Fiability and Durability, ISSN 1844-640X, 1, Editura Academica Brâncuşi.

2. Miriţoiu, C.M., (2014) Correlations between some mechanical characteristics and the mass per unit length for some composite bars reinforced with steel fabric, Fiability and Durability, ISSN 1844-640X, 1, Editura Academica Brâncuşi, 112-119.


125

3. Miriţoiu, C.M., (2014) Some calculus relations for determining some of the mechanical properties of composite sandwich bars reinforced with metal fabric, Fiability and Durability, ISSN 1844-640X, 1, Editura Academica Brâncuşi, 119-126.

4. Zamfirache., M., Stanimir, A., (2009) Research regarding the roughness of the surfaces titanium alloys (TiAl5 Fe2.5) turning, Annals of the University of Craiova, Mechanical Series, 107-112.

5. Zamfirache., M., (2009) Research regarding the roughness of the surfaces at stainless steel grinding, Annals of the University of Craiova, Mechanical Series, 161-164.

6. Cârstea, V., et. al., (2007) Course Support for Turning Processes, Sitech Publishing House, Craiova.

7. Roşu, C., (2013) Some considerations regarding the strength calculus of a lathe tool, Fiability and Durability, ISSN 1844-640X, 1, Editura Academica Brâncuşi, 153-158.

8. Alexandru, I., R., (2011) Studiul comportamentului dinamic al sistemului port-scula –scula aşchietoare –Piesa în cazul operaţiei de frezare cu viteze înalte, Teză de Doctorat, Universitatea « Transilvania » din Braşov, Facultatea de Inginerie Tehnologică, Catedra «Tehnologia construcţiilor de maşini, Brasov.

9. Cernăianu, A., (2005) Maşini-unelte. Elemente de proiectare, organologie şi Cinematică, Editura Universitaria, Craiova.

10. Cernăianu, A., (2007) Researches Regarding the Linear and Angular Deformations of the Shafts of the Centerless Grinding Machines, Annals of the University of Craiova, Faculty of Mechanics, 123-135.

11. Freudenberger, J., s.a., (2010) Springer Handbook of Mechanical Engineering. Volume 10. Materials Science and Engineering, pp. 75-218, Springer.

12. Inţă, M., E., E., (2010) Contribuţii privind monitorizarea proceselor şi a echipamentelor pentru aşchierea materialelor, Teză de Doctorat, Universitatea “Lucian Blaga” din Sibiu Facultatea de Inginerie “Hermann Oberth ”Catedra Tehnologia Construcţiilor de Maşini, Sibiu.


126

MECHANISMS USED FOR DRIVING WINDOWS OF CAR SIDE DOORS

Viorica VELIȘCU, Transport C.F. High School of Craiova, [email protected] Daniela ANTONESCU, "Iuliu Maniu" High School of Bucharest,

[email protected] Dan MESARICI, Transport C.F. High School of Craiova, [email protected]

Abstract: The paper presents the main types of the mechanisms used for opening and closing of windows on the car doors. It is identified the three types of the mechanisms used as “crane” for the door windows: the mechanism with bars and gears, the mechanism with cable and the mechanism with elastic rack.

Keywords: “crane” mechanism, door window, bars, gears, cable, elastic rack.

1. GENERAL ASPECTS For sliding of the windows on the car side doors there are used the so-called “crane”

mechanism of side windows [2]. This mechanism allows the raising (closing) and lowering (opening) respectively of car side door window.

Usually, the window of the car side door has got a rectilinear translation (fig. 1a) or, in the case of some modern cars [1,4], the displacement of window is a curved translation (fig. 1b). a) b) a) b)

Fig. 1. Displacement of door window: Fig. 2. Position types of the window guides a) Rectilinear translation on line (A); b) Curved translation on curve (A).

On classical cars, these mechanisms are driven only manually with the hand crank, while on the modern cars the driving is both manual and electrical, through distance control.

The sliding of the window in the direction of the displacement is achieved in a rubber guide, which is fixed in the metal frame of the car side door.

The window is guided by a rubber frame, which is fixed in a metal support (fig. 2).

(A)

D

C B

E

A

D

(A)

B C

E

A

D

C B

E

A

D

C B

E

A




127

This guide has at its lower side one or two horizontal (fig. 2a) or inclined (fig. 2b) openings.

The support in which the window „crane” mechanism is mounted lies in the internal panel of the door, this panel being achieved in a frame shape [2].

The base plate of the window mechanism is fixed; to this frame has got an aperture through which out the drive shaft goes. This shaft has a crank fixed to it.

The driving force of the „crane” mechanism, for raising the window, must not exceed the value of 15 kN. In the case of the electrical action, this is achieved with an electromotor with a small gauge, which is supplied with 12V d. c. from the accumulator battery.

In the following, there are presented some types of construction schemas of „crane” of the mobile windows, from among which mention is made of: the bar and cylindrical gearing mechanism, the mechanism with elastic rack and the mechanism with cable and guide rollers. 2. PLANAR MECHANISM WITH BARS AND GEARS

From among the types of planar mechanisms with articulated bars, the most widely used is the four bar linkage, the crank - rocker type (fig. 3) or parallelogram linkage type [3], such as the mechanism used on the Dacia car [2].

The crank 2 has a geared segment, attached to it, where the latter has z2 teeth. The geared segment is in gearing with the pinion 1, with z1 teeth. On the pinion shaft 1 there is fixed the hand crank for driving the mechanism.

A

Sec?iu

nea A

-A

Sec?iu

nea B

-B

B

B A

1

2

3

4

5

A

A0

B0

B

C

Fig. 3. Constructive schema of planar mechanism for window displacement, achieved with spur cylindrical gearing and four bar linkage crank-rocker type.

The bar 3 represents the connecting rod in the linkage quadrilateral A0ABB0 and is prolonged with the segment BC.

A portion from the path of point C is used for the displacement of the window frame 5 in the direction of opening / closing the window. The connection between the bar 3 and the window frame 5 is achieved with a high joint of rotation - translation [3].


128

The rocker 4 is articulated at base in B0 and swings between two limit positions, which correspond to the extreme positions of (open / close) window.

The constructive schema (fig. 3) presents the lower (open) position of the side door window. The disadvantage of this solution is that then the frame 5 starts hoist together with the window, the former tends to occupy an oblique position against the vertical guide, and the mechanism is blocked. The length of the groove inside frame 5 corresponds to raising (closing) the window, such that in the top position, the point C of applying the raising force is placed at the middle point of the window width. The spur cylindrical gearing (made up of pinion 1 and gear sector 2) is placed behind bar 4 and the pinion 1 axis is on the left of the fixed joint A0 and above the other fixed joint B0. In this way, the support plate of the “crane” mechanism has a smaller gauge. The number of evolving teeth of pinion 1 is z1 = 6...12, and for the gear sector 2 the number of teeth can be z*2 = 35...90, in proportion to angle at the circular sector, whose values are = 900 ...1800. In the case of the above mechanism (fig. 3), the gear sector 2 has the angle = 1800 and the number of teeth is z*2= 90. The cylindrical gearing achieves a reduction, where the transmission ratio from pinion 1 to the gear sector 2 being i12 = - 2z*

2/z1 . A variant of the bar mechanism is the one using the crank-slider mechanism (fig. 4), where the slider is replaced by a high rotation-translation pair which guides the connecting rod 3 through a bolt inside a horizontal fixed guide.

35°

Secţiunea A-A

A

A

1

2

2

1

Secţiunea B-B

Secţiunea C-C

B

B

C

C 36

3

zz

'21

z'2'

Fig. 4. Constructive schema of crank-slider mechanism type

2

3

3

A0

A

B

2

C D

2

4

3


129

Typical of this bar mechanism of a crank-slider type (fig. 4) is the fact of the crank length A0A being equal to the AB length of connecting rod 3.

Both the connecting rod 3 and the crank 2 are extended with the same length, that is, segment AC is equal to AD, which makes the straight line defined by the points C and D be parallel to A0B, that is the horizontal position.

The bar 4 has a translation movement along a vertical line, that is perpendicular to the fixed guide of point B, center of the bolt or of the roller at the end of the connecting rod AB.

The spur cylindrical gear is a reducer, being formed by the pinion 1 (cu z1 = 6...8 teeth) and a gear sector (with the angle of 900 ...1000 at the center of the gear crown). The sector is solid with the crank 2 and has z2 teeth, where z2 = z’2 + z’’2 = 40...45 teeth. 3. MECHANISM WITH PINION AND HELICAL SPRING - ELASTIC RACK

These mechanisms are the most simple, having only two mobile elements (fig. 5): the drive pinion 1 and the elastic rack 2 as a driven element.

This special rack is under the form of a helical spring (fig. 5a), having inside a synthetic core. The helical spring has a proper stiffness so that it doesn’t modify its pitch, operating as a rack. The synthetic core is impregnated with mineral oil, through which the lubrication of the gear-rack gearing is ensured. Fig. 5a

460

30

125 12

140

70

R120

300

R60

735

440

R65

Fig. 5. Constructive schema of pinion and helical spring – elastic rack mechanism

2

0

3

C

D

Helical spring

1(pinion)

2 Synthetic core

1


130

The elastic rack is guided by a fixed metal tube 0, in the shape of letter S, provided with a longitudinal slit and makes the opening ends.

The rod 3 is horizontally connected on these open parts in the points C and D (fig. 5), the rod being provided with a supporting frame for the door mobile window.

By choosing the S-shaped path of the elastic rack 2, the two points C and D of the (helical spring) rack move in the same direction on a vertical line, which allows bar 3 to be maintained in a horizontal position and also achieves a translation movement of the window.

The overall length of the metal tube 0 is longer than the length elastic rack 2 with approximately the value of the maximal opening stroke of the side door window.

Besides the casing of the pinion – rack - gearing, the metal tube 0 is also fixed to the side door girder in at least other four points (fig. 5).

A variant of the elastic rack mechanism uses the metal tube with the opened loop in the shape of letter U (fig. 6), so that the window is supported on one single point C, on the right side of the tube.

cauciuc

430

200

R 77

10

12

20

155

Fig. 6. Constructive schema of the windows displacement mechanism (rotated with 900);

this is achieved with the pinion and the elastic rack in U shape, variant 1.

The line of the tube is fixed in two points D and E (fig. 6) on the door girder. In order to reduce the gauge of this mechanism, the metal tube extends to the other side of the pinion-rack gearing, with a rubber hose, having its free end on the support in D.

The casing 0 of the pinion 1 – elastic rack 2 gearing (fig. 6) allows for a gearing along a longer length, which ensures the transmission of a higher force through the helical spring to the mobile window.

A second variant of the window drive mechanism which has one single support (fig. 7) uses the pinion 1 – rack 2 gearing with a smaller continuity degree.

C D

E

1

2

0

Tube of rubber


131

Fig. 7. Constructive schema of the mechanism for the window displacement (rotated with 900), achieved with a pinion and an elastic rack in U shape, variant 2

In the part of maximum curving F, the tube is replaced by a special casting between

the curved and linear parts. Beyond the gear-rack gearing (1, 2), the tube extends through a rubber segment for the helical spring to penetrate into this part when the window descends to its lower limit given by point E. 4. CABLE AND ROLLER MECHANISM

As a topological structure, this mechanism is the most simple, having a shaft provided with a cylinder 1 round which the metal cable 2 winds (fig. 8).

Under normal conditions (steel on steel) the cable 2 has 4 - 6 windings round the drive shaft 1 (fig. 8). In order to protect the metal cable, its two branches pass through plastic tubes. When coming out of these tubes, the cable passes over two rollers, after which the cable becomes rectilinear between the 2 rollers, between points C and D (fig. 8).

Fig. 8. Constructive schema (rotated with 900) of

the window mechanism, with crank-shaft and cable.

A

A

Secţiunea A-A

5înfăşurări

clemă

C D E

1

2

F

2

1

2’ 2

C D

hoist

F


132

Cable 2 winds round the shaft 1 (fig. 8), through with the displacement of the window supported on the support 2’ (solid with cable 2) is performed manually or through the control of a d. c. electric motor). The number of windings of cable 2 on the shaft 1 must be sufficiently to stop the two surfaces slide on each other (one cylindrical and the other ring-shaped. REFERENCES

1. Dudiţă, F., Diaconescu, D., Gogu, G., Mecanisme articulate, Editura Tehnică Bucharest, 1989; 2. Mondiru, C., Autoturisme Dacia, Editura Tehnică Bucharest, 1990; 3. Antonescu, P., Mecanisme, Editura Printech Bucharest, 2003; 4. Antonescu, P., Dugăeşescu, I., Antonescu, O., Mecanisme pentru acţionarea geamurilor uşilor laterale de automobile, Rev. Mecanisme & Manipulatoare, Vol. 2, Nr. 1, 2003, pp. 21-26.


133

TOPOLOGICAL STRUCTURE AND MOBILITY OF THE MECHANISMS USED IN CAR MECHANICAL JACKS

Viorica VELIȘCU, Railroad Transportation H. S. of Craiova, [email protected] Dan MESARICI, Railroad Transportation High School of Craiova, [email protected] Dr. Păun ANTONESCU, „Politehnica” University of Bucharest, [email protected]

Abstract: This paper presents a structural analysis of the mechanism of high topological type jack - screw and translator rectilinear- patina and mobility mechanism analysis using various generally applicable formulas.

Keywords: lifting mechanism, jackscrew, mechanical jack, topological structure, mobility

1. INTRODUCTION Jacks are devices used for lifting loads at a low height without any flexible lifting part.

Jack building forms depend both on the load lifting way and on their driving type. Furthermore, we shall present several kinematic schemes for different types of lifting mechanisms used for car mechanical jacks [1, 2].

Fig. 1.1. Screw and patina lifting mechanism Fig. 1.2. Screw and parallelogram lifting mechanism

Fig. 1.3. Lifting mechanism with pneumatic actuator Fig. 1.4. Screw and parallelogram lifting mechanism and parallelogram mechanism

1 2

3

4

5

0

0

4

5

6

0

3

2 1

0

6

5

4 0

3 2 1

0

6 5

4

0

0 3 2 1

0

1




134

Besides these relatively simple kinematic schemes, there are also complex ones (fig. 1.5 ... 1.11).

Fig. 1.5. Lifting mechanism with screw Fig. 1.6. Lifting mechanism with screw, hexagonal chain and cylindrical gears pentagonal chain and cylindrical gear Fig. 1.7. Lifting mechanism with screw, articulated bars and a cylindrical gear Fig. 1.9. Lifting mechanism with screw and a flexible support bar

5(0)

3 2 1

4

0

5

4

0

3 2 1

0

5 4

3 2 1

6 7

0

8

1

8

7 6

5

3 2

1

0

4

Fig. 1.10. Lifting mechanism with screw and articulated bars

Fig. 1.8. Lifting mechanism with screw and articulated bars

7 6

0

5

3 2 1

4

9

8 1 7

6

0

5

3 2 1

4

9 8

10

1


135

Fig. 1.11. Lifting mechanism with two hydraulic cylinders and parallelogram, manual drive

2. STRUCTURE AND MOBILITY OF THE VARIABLE DIRECTION JACK-SCREW MECHANISM.

The variable direction jack-screw (fig. 1.9, 2.1) is fixed to the ground by means of a flexible support leg frame bar 5. Among mechanical jacks, the most common ones are those provided with a motion screw, where the jack is actuated by rotating the lever attached to the screw.

Such a jack type is presented in the figures below with: - the kinematic scheme of the lifting mechanism (fig.2.1.a). - photo image of the jack-screw with a flexible support arm under load (fig.2.1.b).

a) b)

Fig. 2.1. Lifting mechanism of a jack-screw with a flexible support arm

The kinematic scheme of the analyzed mechanism (fig. 2.1) shows that it includes a closed kinematic chain, formed of the elements 1, 2, 3, 4 and 5. Among these kinematic elements, bar 5 acts as a fixed element 0, once it is supported on the horizontal plane.

This closed kinematic chain is a spatial 3D chain as the rotation of element 1 (the driving screw) goes beyond the motion plane of the other parts.

In order to use this jack (fig. 2.1b) we mount the support bar 5 (fig. 2.1a) on the horizontal surface close to the car body that is about to be lifted.

1 2 3

4

5 0

0

6

6 7

8 1

5(0)

3 2 1

4

0

A B

C

Mm

Fr


136

The space associated to the closed kinematic chain has the rank r = 4, and thus the mobility of this screw mechanism can be determined with formula [3]

6

2

5

1 rr

mm rNmCM (2.1)

Visually inspecting the jack mechanism (fig. 2.1a), we notice that all kinematic couplings have a single mobility ( 1m ) and therefore 5mC . Kinematic couplings are the overlapping

articulations A(1,2) (2,5) and the articulations B(3,4) as well as C(4,5). It should be mentioned that the screw coupling (1,3) overlaps in B, whose mobility 1m ,

just like with the articulation, as there is an interdependence between the two relative rotation and translation motions and s, through the p lead of the motion screw: sp 2 .

In this particular case 1rN (number of rank r contours), formula (2.1) is written

145 rCM m (2.2)

The result obtained is 1M , which proves that this mechanical jack works by means of rotating the joint lever with screw 1 (fig. 2.1a).

Thus, by rotating lever 1 with momentum Mm object 3 moves vertically and surpasses the resistance force Fr represented by the gravity force of the car body.

3. TOPOLOGICAL STRUCTURE AND MOBILITY OF THE JACK-SCREW WITH BARS AND CYLINDRICAL GEARS.

Let us consider a mechanical jack-screw with bars and cylindrical gears (fig. 1.5, 3.1) in the rest position, next to a car (fig. 3.1a) and under load, in the working position (fig. 3.1b).

a) b)

Fig. 3.1. Jack-screw with bars and gears: rest position (a) and working position (b) The kinematic scheme of the mechanism used for this mechanical jack (fig. 3.2a)

symbolises a hexagonal contour of articulated bars, two circular gears (segments) and a motion screw 1 as the driving element.


137

a) b) c)

Fig. 3.2. Kinematic scheme of the mechanism - real (a), equivalent (b) and kinematic chain (c) To obtain the equivalent mechanism (fig. 3.2b) we equate each cylindrical gear: gear (6,7)

with a bar-type kinematic element e67 and two articulations C' and D', and the gear (4, 5) with fictive bar e45 and the articulations A' and B'.

Fictive bar e45 forms with bar 0 and bars 4 and 5 a second tetradic chain (4, 5, 0, e45) symmetrical to the other with respect to the screw axis 1. Thus, the lifting mechanism results from a hexagonal kinematic chain achieved by means of the symmetrical connection of the two tetradic chains (fig. 3.2c), when the potential articulations A and B become active.

Following the kinematic scheme of the mechanism (fig. 3.2a) we identify an articulated hexagonal contour BDCABA 00 (0, 4, 7, 6, 5), whose kinematic associated space has the family

31 f and the rank 36 11 fr . The driving element 1 (the screw), together with patinas 2 and 3, divide the hexagon into

two articulated pentagons, each having a space corresponding to 22 f respectively 42 r . Each cylindrical gear defines, together with bar 0 or 8, a closed independent contour: (0,

4, 5) and (6, 7, 8), and each has a corresponding space 33 f and 33 r . To determine the

mobility of the mechanism (fig. 3.2a) we use formula [3]

5

1)()6(

kkaafa CfknfM (3.1)

On the kinematic scheme of the mechanism we can identify ten couplings class 5k and two couplings class 4k . Subsequently 105 C , these are (0,4), (0,5), (1,2), (2,5), (2,6),

(1,3), (3,4), (3,7), (6,8), (7,8) and 24 C , being the gears (4,5) and (6,7). Inspecting the kinematic scheme of the analysed mechanism (fig. 3.1a) we identify the

existence of four independent closed contours, with the specific restrictions given by the families mentioned above: 3,3,2,3 4321 ffff . The apparent family af is determined

as the arithmetic mean of the four real families:

4/11)3323(41

4

141

i ia ff (3.2)

8

7 6

B

C D

D' C'

8'

A

0'

0

5 4

A0

B0

Mm

8

7 6

5

3 2

1

0

4

1

A B

A0 B

C D

Fr

5

3 2

0

4

A

AB

8 7 6

1 B

C D

D'

C'

e67

B

A'

e45


138

All numerical data shall be replaced in the formula (3.1) and it is obtained 12)4/114(10)4/115(8)4/116(4/11 M (3.3)

This result proves the fact that there is a single driving element – screw 1, which, by means of rotation, transmits the motion of all the elements including bar 8.

The mobility of the mechanism can be also determined by means of the formula [3]

5

1

6

2m rrm rNmCM (3.4)

For the analysed kinematic scheme we identify the following values ( km CC ): 101 C

class 1 kinematic couplings, of which nine couplings are plane articulations and one is the coupling screw-nut (1,3); 22 C class 2 kinematic couplings, these being the gears (4,5) and

(6,7); 33 N rank 3 closed contours ( 3r ); 14 N rank 4 closed contour ( 4r ).

After replacing these numerical values in the formula (3.4) we obtain 1)1433()22101( M (3.5) The result obtained by means of formula (3.4) coincides with the previously obtained one

1M , which proves the univocal transmission of the screw 1 motion to all the other kinematic elements.

4. THE JACK-SCREW MECHANISM WITH ARTICULATED BARS

4.1. The kinematic scheme of the jack-screw and the operation mode According to the kinematic scheme of the jack-screw with articulated bars (fig. 4.1), this

is manually driven with the help of a screw provided with a lever.

Fig. 4.1. Kinematic scheme of the jack-screw with bars The topological structure of the mechanism (fig. 4.1) is complex, highlighting a hexagonal

contour of articulated bars in a vertical plane with double concavity. The articulated plane kinematic chain (0,5,9,10,7,4) is a concave hexagon A0ADCBB0

consolidated by means of 6 bars (to the lower side) and 8 bars (to the upper side). Both lower bars (4, 5, 6) as well as upper bars (7, 8, 9) move in vertical parallel planes.

The geometrical configuration of the concave hexagonal mechanism (fig. 4.1) supposes the equality of bars 4, 5, 7 and 9. Bars 6 and 8 have an equal length, so that the quadrangle AFBE (fig. 4.1) forms an articulated plane parallelogram.

7

6

0

5

3 2 1

4

9 8

10

1

Fr10

Mm1

A0

A B

B0

C D

E

F


139

4.2. Topological structure and calculation of the mechanism mobility The mechanical jack mechanism (fig. 4.2a) consists of two kinematic chains, of which an

open kinematic chain (1,2,3), with potential articulations A and B (fig. 4.2c) and another closed chain (fig. 4.2b), made up of seven mobile kinematic elements (4,5, ... ,9,10) assembled in a complex topological structure with articulated bars, including the fixed element 0.

c) a) b)

Fig. 4.2. Kinematic scheme of the mechanism (a) and the component kinematic chains (b, c)

The closed kinematic chain (fig. 4.2b) has three independent plane contours with articulated bars: the hexagon (0,4,7,10,9,5), the quadrangle (0,4,6,5) and the parallelogram (5,6,7,8). The open kinematic chain (fig. 4.2c) consists of screw 1 connected to patina 2 through a rotation coupling and to the patina-nut 3 through the rotation-translation coupling.

The two kinematic chains are assembled by means of the two potential articulations A and B, and then they become triple articulations (fig. 4.2a).

It can be proved that, by the rotation of the motion screw 1, the hexagonal type closed kinematic chain moves all the component bars in a univocally determined way, of which the vertical motion of bar 10 is of interest.

5. CONCLUSIONS The structural topological analysis of mechanisms used in mechanical jacks has been concluded by

deducing the structural formula for each independent kinematic contour and for the entire mechanism. To determine the mobility of such mechanisms, we have used generally valid formulas, checking

that this allows univocal transmission of the driving element’s motion to the driven element. Considering all the things above, it results that the drive element, the one onto which the driving

mechanism is attached, does a rotation motion, while the driven element does a translation motion.

REFERENCES 1. Antonescu, P., Mechanisms and Manipulators, Applications – Projects, Printech H., Buch., 2000; 2. Antonescu, P., s.a. Mechanisms and Manipulators – a Laboratory Guide, Printech H., Buch., 2002; 3. Antonescu, P., Mechanisms, Printech Publishing House, Bucharest, 2003.

Mm1

3 2 1 1

B A

7

6

0

5

9 8

10

A0

4

A

B0

C D

E

F

B

7

6

0

5

3 2 1

4

9 8

10

1

Mm1

A0

A B

B0

C D

E

F


140

STUDY ON THERMAL DEFORMATIONS OF THE PRIMARY SEALING OF FRONT SEALING

PhD.Lecturer Mihaela ISTRATE,University of Pitești,Faculty of Mechanics and

Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected] PhD.Lecturer Monica BÂLDEA,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

PhD.Lecturer Ancuța BĂLTEANU,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

PhD.Lecturer Jan Cristian GRIGORE,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

Abstract: The thermal phenomena appear at the level of the film within the primary sealing interstitium. An important temperature gradient is performed in the stator and in the rotor, which produces thermoelastic deformations. These deformations are of the order of film width and affect essentially the interstitium geometry. According to the temperature increase direction the repartition in the friction ring is different. The farthest areas from the temperature drops or the nearest to the heat sources will have he highest temperature. These dilate more that the rest of the areas and modify the interstitium form. From the calculation relations it comes out that deformations depend also on certain operating conditions, which can be modified through time (pressure, temperature), the sealing efficiency being thus different in time. Keywords:Interstitium, temperature, deformations, sealing. 1.INTRODUCTION The space between the contact front surfaces represents the most important part of front sealing. An adequate lubrication is required in order to minimize the wear and to provide an accurate sealing (a very small loss debit). This lubrication is provided by the sealed fluid or by the cooling or locking fluid. Front sealing optimal functioning is ensured if the width of the film separating the two surfaces is of the order of 1-1,2µm. The control of the lubricating film width must be extremely precise, in order to avoid contact areas without lubricant which may lead to a premature wear for the primary sealing and to an inadequate functioning of the device. Nowadays, the industrial target is to be able to create reliable front sealing with a loss debit almost null and with a minimum wear. Also, the thermal phenomena appear at the level of the film within the primary sealing interstitium. An important temperature gradient is performed in the stator and in the rotor, which produces thermoelastic deformations. These deformations are of the order of film width and affect essentially the interstitium geometry. Consequently, thermoeasltic deformations play a preponderant role in front sealing stability. The hydraulic balance term used to face seal is used to specify the relationship between the ambient pressure and the contact pressure sealed to the sealing surface. From the viewpoint of the term hydraulic balance, front seals are classified in "balanced" and "unbalanced". For a balanced face seal contact pressure can be controlled so that it is possible to maintain a lower value hydraulic allowing the formation of a film of greater thickness. For this reason,a balanced face seal has the possibility of handling fluids with higher pressures and difficult operations,than a face


141

seal unbalanced. Normally a balanced face seal is designed to work with the pressure lower sealing surface which practically minimal losses between the surfaces. Types of outer face seal are illustrated in Figure 1.a, Figure 1.b and Figure 1.c,presents different arrangements of the sealing surface. In this figure, d is the effective diameter sealing, A is the surface on which the sealed fluid pressure acts, and B is the area of contact sealing surface. Force of the spring is ignored in all cases.

a)

b)

c)

Fig. 1.Different arrangements of the sealing surface.


142

In Figure 1.a, the entire contact surface B is arranged outside of the effective diameter of sealing "d" and the hydraulic surface area A is equal surface of contact B (A = B). This construction is a condition of 100% imbalance (out of balance), which also indicates that the mean contact pressure will be exactly 100% sealed hydraulic pressure

1KB

A (1)

In figure 1.b the entire surface of the contact B is disposed outside the effective diameter of sealing "d" and the area of action of the sealing pressure A is greater than a contact area B (A> B). In this case, the face seal is in imbalance, according to the report areas A and B

1KB

A (2)

Supported by the contact surface pressure is bigger with a report equivalent to the hydraulic pressure which is sealed. This is the condition in most unbalanced front seal. Figure 1.c shows the relation between the most balanced face seal. Here, a part of the contact surface B is indicated by B1, and is disposed outside the effective sealing diameter "d". The area B1 is therefore equal to the area A. Because the remaining area B2 hydraulic sealing surface is located inside the effective diameter of sealing "d", the total area of the seal B is equal to the sum of B1 and B2. Charging the sealing surface will be less than the pressure sealed, so:

121

1 KB

A

BB

B (3)

This value, expressed in percent indicates the degree of imbalance in front seals. 2.RINGS MOTION Each ring of primary sealing is connected to a support (shaft or case-frame) by joints that may be fixed or deformable. In this case, the ring rotating around the main shaft holds a maximum of 5 degrees of freedom, figure2

Fig. 2.Degrees of freedom


143

three elementary translations: an axial displacement and two radial displacements (or eccentricities) ; two elementary rotations: angular misalignment around the orthogonal direction ( ) on the main axis (z) and a rotation around the z axis. In numerous cases, one of the rings is rigidly connected to its support and moreover, the eccentricities of the rotating ring have a negligible effect in relation to the other degrees of freedom. This leads to a model of primary sealing with three degrees of freedom, figure 3.

Fig.3. Primary sealing with 3 degrees of freedom

3.SEALING PARAMETERS The operating parameters influencing an adequate sealing are defined first of all by the dimensions and assembly of the primary sealing. An important influence on non-sealing, lifetime, friction losses and operation safety are briefly presented hereunder: The hydraulic pressure ratio K and the proportion between the pressure exercised by the elastic element against the sealed medium pressure par/p1; The sliding speed between the rotor and the stator (friction); Surface roughness of and the parallelism of the friction surfaces ; The sealing medium temperature and the friction surfaces temperature; The form of the sealing interface depending on the mechanical and thermal deformations susceptible of occurring during operation ; The materials couple; The sealing fluid with its lubricating and cooling properties, its contamination degree, etc.; The friction type, the oscillations, the wear, the periodic idle time, the fluid circulation clockwise or counterclockwise, the eccentric operation, the assembly, the cooling etc. In most of the applications primary sealing with level sliding surfaces are used. These surfaces can be modified due to the influence of heat, tensions and wear. Level surfaces offer


144

the advantage of being able to be performed and controlled with simple means. Under the influence of axial and radial forces applied on the primary sealing rings, and through the temperature differences, the rings are deformed and the interface can become concave, convex or tilted with a contact on the exterior D or interior d diameter, or tilted without any contact. If the sealing functioning conditions remain constant, the friction surfaces remain parallel under the wear effect and subject to the adequate materials couple and to the sufficient time of applying a permanent contact pressure. In total three main factors influence the primary sealing deformations: axial, radial forces and temperature gradients. 4.THERMAL DEFORMATIONS INFLUENCE The operating temperature differences influence the interstitium geometry. Elastic deformations depend on the elasticity module and on the dimensions and thermal deformations depend on the thermal properties of the material with a heat conductibility ratio , a thermal dilatation ratio and on the thermal transmission ratio. The temperature gradients that can be in a radial or axial direction influence the interface geometry. 5.EFFECTS OF THE AXIAL TEMPERATURE GRADIENT Deformations, due to the axial temperature gradient, generate a conical increase in a radial direction for the temperature diminution towards D and a conical narrowing of the ring for a temperature diminution in d ( see fig. 4 a, b). According to [7] the relation for the deformation under temperature influence for a linear axial gradient is:

2

22Fiaa

rT

crrcS

(4)

where:

l

TTc A

a

(5)

For the case a, the deformation is considered negative and positive for the case b , figure 4.

Fig.4.Deformations due to the axial temperature gradient


145

6.EFFECTS OF THE RADIAL TEMPERATURE GRADIENT According to the temperature increase direction at the exterior D or interior d diameter, the repartition in the friction ring is different. The farthest areas from the temperature drops or the nearest to the heat sources will have he highest temperature. These dilate more that the rest of the areas and modify the interstitium form. Admitting a source of heat and constant operating conditions, supposing a linear temperature gradient in the radial direction,

l

TTc A

r

(6)

will determine the deformation of the ring in the axial direction with the approximate relation :

raT cblS (7)

When the temperature drop is in D, TaS will be negative and reverse.

7.CONCLUSIONS CONCERNING THE INFLUENCE OF INTERSTITIUM DEFORMATION The geometry of the primary sealing surfaces is affected in operation by the amount of individual mechanical and thermal deformations of the friction rings. The total deformation is: BA SSS (8) For an operation with parallel surfaces it is necessary that the amount of individual deformations according to the equations be null. Nevertheless, individual deformations depend on geometry, material and installation, and this ideal solution cannot be performed in practice. All the observations of individual deformations influence have been directed towards the interface configuration, supposing contacts of the two rings in D or d with the return to parallel surfaces after wear. The most important consequence of deformations is the leaking of the sealed fluid as a result of interface modification. If, for example, by the production of a configuration by interface deformation allowing the introduction of the fluid under pressure in the interface, a hydrostatic discharge occurs and the lost debit will increase. By rings contact return to D or d, the wear will restore a new interface with parallel or slightly conical surfaces. From the calculation relations it comes out that deformations depend also on certain operating conditions, which can be modified through time (pressure, temperature), the sealing efficiency being thus different in time. It can be appreciated that an inadequate contact in D or d, with moderate deformations, can be improved through time by means of running wear. REFERENCES [1]Crudu,I., Etanşări pentru organe de maşini in mişcare.Tribosisteme industriale, Tribotehnica 80, Bucureşti, [2] Istrate, M., Studiul etanşărilor primare la etanşările frontale, Editura Larisa, Câmpulung, 2013 [3]Istrate, M., Baldea,M., On predictive control using vibration detection method on condensation damage to themachine with single stage pump,Scientific Bulletin.Automotive,year XII,nr.16


146

[4]Lazăr, D.,ş.a. Influenţa mediului asupra alegerii cuplului de materiale pentru inelele etanşărilor frontale. Tribotehnica` 87, 24-26 sept., Bucureşti, vol. III, p. 79-84. [5]Popa, N.,ş.a. Asupra uzurii si tipurilor de uzură din etanşările axiale ale pompelor din industria petrochimică, Tribotehnica 87, Bucureşti, 24-26 sept. 1987, p. 139-143. [6]Popa, N.,The Reynolds equation solving for the constant central thickness hydrodynamic mechanical face seals, 3rd Vienna International Conference on Nano-Technology, march 18-20, 2009, Vienna, Austria [7]Cicone,T.,Apostolescu,A., Temperature distribution in the rings of liquid face seals-evaluation of simplified analytical models using 3D FEM analysis,BALKANTRIB’O5, 5th International Conference on Tribology,June 15-18,2005, Kragujevac, Serbia and Montenegro [8]Tournerie,B.,Brunetière,N., Danos,J.C.,2D Numerical modeling of the TEHD transient behaviour of mechanical face seals,Proc. of the 17th BHRG Conf. on Fluid Sealing, 2003,York, UK.


147

STUDY ON MECHANICAL DEFORMATIONS OF THE PRIMARY FRONTAL SEALING

PhD.Lecturer Monica BÂLDEA,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

PhD.Lecturer Mihaela ISTRATE,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

Prof.PhD.Alexandru BOROIU,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

Assoc.Eng.Andrei Alexandru BOROIU,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

Abstract: In many applications one of the primary sealing rings is mounted pressed into another ring, tightening materials to be made makes the required stretched tension in the outer ring and compression in the inner ring.It will be considered as the basis for calculating ultimate elongation for tenacious materials and breaking limit for brittle materials. Keywords: Tensions,mechanical,deformation, ring, primary sealing

1.INTRODUCTION In many applications one of the primary sealing rings is mounted pressed into another ring. Usually the graphite ring is mounted into a steel one and the pressed assembly is for example the stator or the rotor. In this way the necessary tightening applies over materials elongation tensions on the outer ring and compression on the inner ring. It will be considered as the basis for calculating ultimate elongation for tenacious materials and breaking limit for brittle materials. 2.THE CALCULATION OF THE MECHANICAL DEFORMATIONS In the figure 1 there are distinct presented those two rings from different materials, where the tightening is

S=D-d1 (1)

Fig.1.Rings from different materials

The elongation tension is [4]


148

22

111

1

12

1

2

11

11

11Ee

Eed

SE

d

D

E

E

e

ed

SE (2)

Compression tension in the 2 ring is [4]

11

22

2

1

1

2

1

2

22

11Ee

EeD

SE

DD

d

E

E

e

e

SE (3)

In these relations e1 and e2 represent the thickness of the rings (see figure 1) and E1 and E2 are the elasticity modules of the rings materials. These tensions will be compared with the tensions allowable at a safety factor c>3 [4]. In the same time, the two elements are receiving tangential tensions which influence on the thickness of the rings. For determining those tensions is considered the rings as tubes with thick walls and open ends and under inner and outer pressures (figure 2).

Fig.2.Rings as thick-walled tubes

The tangential tension is [4]

22

221

2 2

ia

aaait rr

rprrp

(4)

For the case when the outer pressure is pa =0:

22

22

ia

iaiz rr

rrp

(5)

And the rays variations (increases)

22

22

ia

iaiii rr

rr

E

rpr

..(6)


149

respectively:

22

22

ia

iaia rr

r

E

rpr

..(7)

For the case when only the outer pressure actions (pi =0),it’s resulting a maximum tension on the inner wall

2

2

1

2

a

i

aD

r

r

p

(8)

and the rays reductions

22

22

ia

aiai rr

r

E

rpr ( 9)

22

22

ia

iaaaa rr

rr

E

rpr (10)

All these relations are available for friction rings without section modifying. For this kind of situation the safety factors are taken into account. For reasons regarding the sealing security, should be excluded the deteriorations through breaking for brittle materials and the plastic deformation for metallic materials, and also should be taken into account the temperature variations, the pomp cavitation, the vibrations, etc. For the rings with section variations should be considered that from a smaller outer diameter to a bigger one should be done through connection rays. For the carbon, graphite or synthetic resins rings with section variations, when manufactured, the surfaces should be made with precision and from one section to another should be done with a appropriate big ray, and also the fraction between the degradation length a and the critical sealing thickness b should be less than 1 (a/b<1, see figure 3). According to figure 3, in those two parts 1 and 2 different tensions are made, so the arithmetic average is considered:

2/21 m (11)


150

Fig.3 Distribution of tensions in rings

Considering the (4) and (10) formulas and a safety factor ca, the formula for the admissible inner pressure is:

22

2

222

221

221

2

ia

ia

ia

iaa

zia

rr

rr

rr

rrc

p

(12)

and according with (8) and (11) formulas, the admissible outer pressure:

22

2

22

221

21

ia

a

ia

aa

Da

rr

r

rr

rc

p

(13)

For dynamic tools or for pomp installations with high vibrations or oscillations, the rings material receives alternant traction and compression tensions. In these situations ca>34. 3. THE INFLUENCE OF AXIAL FORCES Due to uneven application of the forces on the friction rings surfaces (stator and rotor), a force momentum is created from which can result a mechanical deformation. The average radius of those two rings, where the force PA is applied, is:

4

dDrm

(14)

4H

p

dDr

(15)

The value of the torque:

mpAA rrPM (16)


151

Where:

bkpPA 1 (17)

For k values (<1, =1, >1) and also the difference (rp - rm) which can be (<1, =0, > 1), the deformation SAM has positive, negative or 0 values. The same deformations can result also on the B ring. The twisting momentum is: 21 bbPM BB (18) with:

11cpPB (19)

42HB dD

b

(20)

21HB dD

c

(21)

For an ordinary shape ring, the twisting angle will be according to Gemma [6]:

i

a

m

r

rEl

rM

ln

12

3 (22)

and the maximum tensions on 1 and 2 points:

i

ai

m

r

rrl

Mr

ln

6

2max

(23)

In all situations, the angle is low and it can be considering sin . In this case:

FMa cbS (24)

The interface deformation and shape which results under the influence of the axial forces consists of the sum of individual deformations of A and B rings.

aBMAaMaM SSS (25)

To highlight the influence of deformations due to axial forces, the next situation is considered: a primary outer sealing with rotor ring from stellite and the stator ring from graphite are presented in figure 4.The sealing pressure is barpa 25 , isn’t considered the springs influence, and cF =1, D = 62,5 mm, d= 47,5 mm, b= 7,5 mm; dH = 50 mm, K=0,82. Dates:


152

A Ring:la = 15 mm; EA=2.105 MPa; mmrmA 5,27 ; mmrp 28 ;

BRing: mmlB 13 ; aB MPE 4,1 ;DB=68 mm;b1=31mm;b2=30 mm; mmcmmrmB 9;8,28 1 .

Fig. 4. Outer primary sealing

With formulas (16), (17), (22) and (24) is calculated the mechanical deformation of an A rotating ring:

d

DlE

crrrKpbS

AA

FmmpAMa

ln.

12

3

12 (26)

It’s resulting:

mSAMa 28,0 (27)

For B ring, with identical functioning conditions, the deformation is determined with formulas (18), (19), (22) şi (24):

d

DlE

crbbcbpS

BBB

FmBBMa

ln

12

3

2111 (28)

It’s resulting: mSBMa 95,7 (29)

The total deformation is: mSMa 23,895,728,0 (30)

For a 1-2 μm gap thickness at the parallel surfaces, a major fluid loss is resulting. 4.CONCLUSIONS The geometry of the primary sealing surfaces, when functioning, is affected by the sum of individual mechanical and thermal deformations of the friction rings. The total deformation is: BA SSS (31) where the rotor deformation (A) is:


153

rATaATrAMaAMA SSSSS (32) and the stator deformation (B):

rBTaBTrBMaBMB SSSSS (33)

For a functioning with parallel surfaces, the sum of the individual deformations, according to (31) formula, should be null. Even so the individual deformations depend on geometry, material and installation, so this ideal solution isn’t practically realized. All the observations of the individual deformations influence were directed to the interface configuration, which presume contacts of those two rings in D or d, returning to parallel surfaces after the degradation. The most important consequence of the deformations is the flow of the sealed fluid due to the interface modification. If, by example, through realizing a configuration due to a interface deformation, which allows the introduction of pressurized fluid into the interface, then there will be a hydrostatic discharge, in this case the lost flow will increase. Through the returning of the rings contact in D or d, the degradation will reestablish a new interface with parallel or slightly conical surfaces. REFERENCES [1]Istrate,M.,Studiul etanşărilor primare la etanşările frontale,Editura Larisa,Câmpulung,2013 [2]Istrate, M., Baldea,M., Study on axial and radial forces of the primary sealing of front sealing,Scientific Bulletin.Automotive,year XX,nr.24,Vol.B,pag.19,2014 [3]Popa, N.,ş.a. Asupra uzurii si tipurilor de uzură din etanşările axiale ale pompelor din industria petrochimică, Tribotehnica 87, Bucureşti, p. 139-143, 24-26 sept.,1987 [4]Lazăr, D.,ş.a. Influenţa mediului asupra alegerii cuplului de materiale pentru inelele etanşărilor frontale. Tribotehnica` 87, 24-26 sept., Bucureşti, vol. III, p. 79-84. [5]Popa, N.,The Reynolds equation solving for the constant central thickness hydrodynamic mechanical face seals, 3rd Vienna International Conference on Nano-Technology, march 18-20,Vienna, Austria,2009 [6]Huitric, J., Bonneau, D., Tournerie, B. Finite Element Analysis of Grooved- Face Seals for Liquid. The 13 th International Conference on Fluid Sealing, Brugge, Belgium, 7-9 Aprilie. [7]Danos, Jean-Christophe ,Lubrification thermohydrodynamique dans les joints d`etancheite a face radiales. These pour l`obtenir du Grade de Docteur de l` Universite de Poitiers, 11 dec. 2000


154

VIBRATIONS AND EQUILIBRIUM OF THE PLANAR KINEMATIC CHAINS WITH ROTATIONAL KINEMATICAL LINKS WITH

CLEARANCES

Ph. D. Lecturer Jan Cristian GRIGORE, University of Piteşti, Faculty of Mechanics and Technology, Romania, [email protected]

Ph. D. Lecturer Monica BÂLDEA, University of Piteşti, Faculty of Mechanics and Technology, Romania, [email protected]

Ph. D. Lecturer Mihaela ISTRATE, Faculty of Mechanics and Technology, University of Piteşti, Romania,[email protected]

Ph. D. Lecturer Ancuţa BĂLTEANU, University of Piteşti, Faculty of Mechanics and Technology, Romania, [email protected]

Abstract:Based on our previous work in this paper we study the vibrations of a planar chain with rotational links with clearances.We also determined the matrix equation wich leads to the equilibrium positions Keywords: Lagrange’s equations,nonlinear vibrations,multibody 1.INTRODUCTION In our previous work we proved that general matrix equation of motion has the form.

qBC

Fq

R

q

0B

Bm

T

. (1.1)

The equilibrium equations are given by

clearance joint with lkinematica rotational is if ,01

clearanceout joint with lkinematica rotational is if ,

kkT

k

kk

O

O

DD

0D, (1.2)

0FRB T . (1.3) 2. VIBRATIONS OF THE PLANAR SYSTEMS WITH ROTATIONAL KINEMATIC LINKS WITH CLEARANCES 2.1. Nonlinear vibrations. The motion of the system relative to an equilibrium position, position defined by the generalized coordinates having the values 0

iq , ni ,1 , values obtained from the system (1.2),

(1.3), is given by the equations (1.1) in which, if we make the substitution zqq 0 , one

obtains the matrix equations FRBzm T , zBCzB . By numerical solving

of this system, we obtain the time histories both of the displacements )(tzz ii , ni ,1 , and

of the reactions )(tRR ii , 212 ,1 nni . 2.2. Linear vibrations

In the case of the linear vibrations we make the development into the series of the functions B , F and by




155

retaining only the linear terms and using the notations ni

qq ii

,10

0

BB ,

ni

qqii

iiqD

,1

00

BB ,

ni

qq ii

,10

0

FF ,

ni

qqii

iiqD

,1

00

FF , RRR 0 , 000

~RBB ii D ,

020100~

nBBBB , 020100 nDDDD FFFF in the conditions of the

equality deduced from the equation (1.2) 000 FRB T one obtains the matrix equations

zFRBzBzm 000~

DT , CzB 0 , wherefrom, with the notation

01

001

0

1

01

0000~~BmBFmBBmBBFBK DD TT , we get the equalities

CzBmBzFmBBmBR TT D 01

001

0

1

01

0~

,

CBmBzKzm 1

01

0

T . The eigenpulsations for such a system are obtained from the n th degree equation in

2p 0det 2 mK p , (2.1) equation that has 1n roots equal to zero, where 1n is the number of the constraint equations, number which is equal to the number of lines of the matrix B .

(X, Y)

m1g

1O

X

Y

Fig.1. Vibrations of the bar articulated at O acted only by its own weight

As example, for the vibrations of the homogenous bar articulated at O , of length l2 , Fig.1, at which the equilibrium position corresponds to lX , 0Y , 0 , one successively deduces

the expressions

0

00

mg

FF ,

00

mgR ,

cos10

sin01

l

lB ,

l10

0010B ,

mgl00

000~0B , 0F 0D ,

J

mlJ

mgl

00

00

000

0

K , where 20 mlJJ and the equation (2.1)


156

becomes 0

lg00

0

00

0

2

0

222

2

J

mpJ

J

glmmp

mp

and, as easily can be seen, it has two roots equal

to zero and the third given by

0

2

J

mlgp . (2.2)

In the general case, if we consider that the independent variables define the column matrix 1q , and the dependent variables define the column matrix 2q , then in the linear

calculus when 0qB , keeping into account the diagonal form of the matrix m , the

system (1.1), can be brought to the form 11111 FRBqm T , 22222 FRBqm T and from here, eliminating the matrices 1q , R and using the

matrices 21

1111

1222*2 x

TT BmBBmm , CBmBBF 1111

112

* TT we obtain the matrix equation *

211

122*2 FFFBBqm TT . (2.3)

For the system drawn in Fig. 6.1 we successively obtain the expressions

1

111 0

0

m

mm ,

J22m ,

10

011B ,

cos

sin2 l

lB ,

01

mgF , 02 F and the equation (2.3)

becomes 00 mglJ and from here we obtain the eigenpulsation given by the relation (2.2). 3. EQUILIBRIUM OF THE PLANAR SYSTEMS WITH ROTATIONAL KINEMATICAL JOINTS WITH CLEARANCES The equilibrium equations are obtained from the equalities (1.1), and from the

equations

qBC

F

R

q

0B

Bm

T

in which 0q , 0q and they write in the

form (1.2) and (1.3). Thus, for a system with n elements, 1n rotational kinematical joints without clearance and 2n rotational kinematical joints with joints, one obtains nnn 32 21 equations ( 212 nn equations from (1.1) and n3 equations from (1.2)) with nnn 32 21 unknowns, name them:

12n reactions for the kinematical joints without clearance, 2n reactions for the kinematical

joints with clearance and n3 kinematical parameters of the type iX , iY , i , ni ,1 , for the n elements. By solving the system of equations (1.1), (1.2), we determine the values of the generalized coordinates 1q , 2q , …, nq3 , and the reactions generically denoted by 1 , 2 , …,


157

212 nn , values that correspond to the equilibrium positions.

In the case when the matrix of the forces F does not depend on the coordinates iX ,

iY , ni ,1 , then the matrix equation (1.2), using the expressions ki

ikki

ikkk

ik yx cossinsincos )()(E for the matrices )(i

kE separates in n3 equations with 21 22 nnn unknowns ( n angular parameters i , 12n reactions in the

kinematical joints without clearance, 2n reactions in the kinematical joints with clearance and

2n angular parameters k ). For open kinematical chains there exists the relation 21 nnn and, as a consequence, for these, the equilibrium position can be determined from the matrix equation (1.2). 4. CONCLUSIONS Based on the differential matrix equation of motion, we obtained the equations of the vibrations for a planar chain with rotational linkages with clearances. This equation is treated both in the nonlinear case as well as in the linear case. We also determined the equilibrium positions. REFERENCES [1] AMIROUCHE, F., Fundamentals of multibody dynamics,Birkhänser,Boston,Berlin, 2004. [2] ERKAYA, S., UZMAY, I., Investigation on effect of joint clearance on dynamics of four-bar mechanism. Nonlinear Dyn., 58, 179-198, (2009). [3] FLORES, P., AMBRÓSIO, J., Revolute joints with clearance in multibody systems. Comput. Struct. 82, 1359-1369(2004). [4] FLORES, P., Modeling and simulation of wear in revolute clearance joints in multibody systems. Mechanism and Machine Theory, 44, 1211-1222(2009). [5] GRIGORE,J.C.,Contribuţii la studiul dinamic al mecanismelor cu jocuri. Teză de doctorat, Universitatea din Piteşti, 2008. [6] PANDREA, N., Calculul dinamic al convertorului mecanic de cuplu„ G. Constantinescu” IFToMM Int. Symp. SYROM`89 pag. 673-679, Bucharest, Romania, 1989. [7] PANDREA, N., POPA, D., Mecanisme. Editura Tehnică, București 2000. [8] PENESTRI, E., VALENTINI, P., P., VITO, L., Multibody dynamics simulation of planar linkages with Dahl friction, Multibody Syst. Dyn. 17, 321-347(2007). [9]PFEIFFER,F.,GLOCKER,C.,Multibody dynamics with unilateral contacts.Wiley,New York(1996). [10] RAVN, P., A continuous analysis method for planar multibody systems with joint clearance. Multibody Syst. Dyn. 2, 1-24,(1998).


158

STUDY ON THE THERMODYNAMIC ASPECTS OF FRONT SEALING

PhD.Lecturer Mihaela ISTRATE,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

PhD.Lecturer Jan Cristian GRIGORE,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected] PhD.Lecturer Monica BÂLDEA,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

PhD.Lecturer Ancuța BĂLTEANU,University of Pitești,Faculty of Mechanics and Technology,Târgu din Vale Street,no.1,Pitești,Romania,[email protected]

Abstract:Thermodynamic lubrication in the radial sealing junctions emphasize the interest for knowing the thermal effects on the mechanical seals. This study is divided into two parts, dedicated to the elaboration of two numerical models for a seal thermodynamic behavior,first part discusses a three-dimensional numerical model in a permanent regime, in the second part of the work, a thermodynamic behavior model for the joints in a transitory regime. Keywords: Thermodynamic ,connection, front sealing, thermal,numerical method. 1.INTRODUCTION The technological evolution allowed the dynamic machines to reach increasing speeds of revolutions for the shafts as well as very high pressure differences between the separation mediums. These problems solicit also the mechanical sealing. The study of front sealing functioning is complex and supposes mechanical phenomena and also interface initial geometry. Besides, numerous situations revealed the thermal effects in many operating conditions as well as their influence on the fluid viscosity sealing parameters. The performance of a study regarding also the thermodynamic aspects is indispensable for the comprehension of a front sealing behavior. It is influenced by several factors. Viscosity affects always the lift and the power dissipated in the interface when the conicity modifies the interface geometry and the dynamic stability. The results are three types of behavior: stable, unstable and cyclic. The effects influencing the transition between an area where the behavior is stable and an area where the behavior becomes unstable have been widely studied. Research has shown that there is a connection between the front sealing dynamic behavior (vibratory instability) and the thermal effects which are the causes of deformations. Finally, the temperature widely influences the front sealing operation causing deformations of the sealing surfaces and of other elements of the sealing device. Film temperature and pressure variations can become serious damage causes determining surfaces direct contact or an important debit of losses. The studies modeling the front sealing THD (thermohydrodynamic) behavior should find the solution to the following equations: a)generalized Reynolds equation for film; b)energy equation for film; c)heat equation for stator and rotor.


159

These equations can be solved analytically, subject to certain simplifying hypotheses, or numerically, in a permanent and transitory regime with limiting conditions for various domains. 2.ANALYTICAL AND SEMI-ANALYTICAL PATTERNS. Permanent regime modeling. An important study is elaborated by Pascovici and Cicone.The hypotheses retained for the analytical solution of the temperature field in the interface are the following: laminar and permanent regime, asymmetric model (conicity at surfaces) and negligible radial flow (no debit). In order to solve the energy equation for film, consider that viscosity and volume mass are always constant within the film. Thus:

en TrTf er cu h

m Tdzh

T0

1

(1)

where: ff Tsi are the dynamic viscosity and the reference temperature respectively, Tm –

the average temperature on a given range, β- the thermoviscosity coefficient, r- the radial coordinate and z- the axial coordinate. The energy equation becomes:

02

22

z

TK

z

Vf

(2)

Using a double front sealing demonstrated a symmetric temperature distribution within the rotor.The stator is insulated and that the temperature dissipated from the interface is evacuated by the rotor. For the easiness of the analytical solution to the problem, they impose a rectilinear form at a φ angle to the flow line,figure 1.

Fig.1

With the hypothesis that all the surfaces exposed to the exterior are insulated,obtain the following relations for the temperatures:


160

h

z

h

z

K

rrTzrT

f

2

1,22

2

(3)

fK

rrTrT

2

22

21

(4)

r

rrTTm 3

22

2

(5)

where: Kf represents the fluid conduction coefficient. The unknown remains T2 which is determined from the thermal balance supposing that the heat evacuated into the rotor by conduction is entirely transmitted by the convection of the surrounding fluid. T2 can be therefore expressed according to the feeding fluid temperature Tf which is known.

s

f K

rr

rHh

rTT

/lncos

2sin

2 0

0

232

2

(6)

Propose an analytical model taking also the stator into account in the front sealing thermal balance, shown in figure 1. Transitory regime modeling Few analytical works approached the thermal behavior to the front sealing in the transitory regime.Elaborate[13] an analytical model using a simplified theory of heat transfer.This model takes into consideration a linear heat flow variation at start and allows the accurate determination of heat evolution in the interface and the rotor. Establish an analytical model in a transitory regime allowing the study of the thermoelastic phenomena during the sealing start period. In order to calculate the rotor deformations, uses the same approximation as Lebeck (1991) but the axial temperature variation is interpolated through an exponential function. 3.NUMERIC MODELS. Permanent regime modeling. Establish a complete THD model [4] which takes into consideration the three-dimensional calculation of the pressures and temperatures field within the film starting from the heat equation within the stator and the rotor. For the rotor, the heat equation is simplified. The equations used and the calculation of rings deformations are solved through the finite elements methods. [4] elaborates a THD behavior model with finite differences,this model solves the generalized Reynolds equation and the energy equation within the film simultaneously. Zhu[4] used a finite element program named CSTEADY for the calculation of the temperature fields in the rings and for deformations.


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Fig. 2 4.EXPERIMENTAL STUDIES Measures the temperatures in interface [4] with 5 thermocouples mounted within the stator. The average temperature varies a lot with the speed of revolutions, with the debit and with the used fluid. Digard et all.[4] elaborates an experimental study for low pressure front sealing where he examines the influence of the temperature gradient on the interface geometry.Accomplish[5] an study on the thermal behavior to a front sealing in a transitory regime. The results emphasize the considerable effect of temperature increase within the interface and show that thermal deformations can be more important than mechanical deformations caused by pressure. Among the three-dimensional thermohydrodynamic models in a permanent regime, the most recent is the model [4]. The geometrical and cinematic model used in this study is described in figure3. It represents a floating stator with three variances (1) and a rotor (2) with a revolution motion having a ω angular speed. The friction surfaces of the two rings can have the misalignments given by the angles 1 and 2 and the conicities given by the angles β1 and β2 . The rotor can present on the friction surface n sinusoidal ripples or any other deformations. Starting from this three-dimensional model determines the pressure field within the film, the temperature field within the film and within the primary sealing for a configuration of the primary sealing with flat surfaces. The calculation can be extended also to misaligned, conical, rippled surfaces or to any other surfaces with referred limited conditions, solving the problems exposed hereinabove on a rotor with flutes on the sealing surface. The study performed on the model presented in figure 3 allowed the acquaintance with the influence of the geometrical parameters and of the operating conditions on the thermal behavior in permanent regime for a front sealing.


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Fig.3 5.CONCLUSIONS The concept, design, manufacture, assembly and use of the front-type mechanical sealing systems are a direct consequence of the introduction of new technological processes, of new general use and process installations and equipments respectively. Research led first of all to the increase of the safety and reliability degree as well as to the expansion of the stress conditions and to the increase of the dynamic equipments functional parameters, such as: - the increase of temperatures and pressures lead to the introduction of the balancing sealing system , of new materials and to the introduction of new geometrical and dimensional elements; - the necessity of sealing the fluids with solid suspensions, having a corrosive-abrasive character as well as the fluids having a predisposition to coke lead to the expansion of the bellows system; - the sealing of fluids with high vapor pressures lead to the use of specific materials and to adaptors for recovery system sealing, etc.; - the necessity of implementing the interchangeability system; - the adaptation of the classical front sealing system with secondary sealing in order to comply with the ecological safety standards, etc. In order to comprehend and to find adequate solutions in the sealing systems concept, design, manufacture and selection, the following considerations are taken into account: -the sealing systems are in fact special subassemblies of the dynamic and general use equipments; - the functional and constructive, manufacture, exploitation and maintenance criteria are conditioned by a complex of parameters depending on: the characteristics of the sealed fluid, the functional operating and exploitation parameters of the basic equipments correlated with the primordial requirement for safety in operation and with the economical criteria.


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Aspects approached by the authors of the articles studied during front sealing development:the development of the processing and resolution speed for the computer graphics and specialized software rendered the evolution of the computer assisted design systems for front sealing possible. Finite element programs for sealing rings tensions and deformations or even for fluid flow through the sealing interstitium were elaborated [7, 4, 10]. Front sealing rings modeling was performed in order to comprehend the causes for sealing obsolescence [8]; The control for front sealing operation was intensified and modernized. Front sealing operation was monitored on the computer [7, 9]. Numerous works related to thermodynamic lubrication in the radial sealing junctions emphasize the interest for knowing the thermal effects on the mechanical seals. More specifically, the heat dissipated by the fluid film shearing causes important temperature gradients in the joining rings. These exercise solids thermoelastic deformations, affecting mainly the interface geometry. Having the same size order as the film thickness, deformations play an important role in the stability of radial sealing joints, and they can lead, in the worst case, to mechanic seal dysfunctions. It is very important to know the thermal phenomena governing these joints functioning, both from a fundamental and from an implementation point of view. REFERENCES [1] Crudu,I., Etanşări pentru organe de maşini in mişcare.Tribosisteme industriale, Tribotehnica 80, Bucureşti, [2] Istrate, M., Studiul etanşărilor primare la etanşările frontale, Editura Larisa, Câmpulung, 2013 [3] Istrate, M., Baldea,M., On predictive control using vibration detection method on condensation damage to the machine with single stage pump,Scientific Bulletin.Automotive,year XII,nr.16,2013 [4] Danos, Jean-Christophe ,Lubrification thermohydrodynamique dans les joints d`etancheite a face radiales. These pour l`obtenir du Grade de Docteur de l` Universite de Poitiers, 11 dec. 2000 [5] Doust,TG and Parmar,A, Transient Thermoelastic Effects in a Mechanical Face Seals. The 10 th International Conference on Fluid Sealing, Innsbruck, 3-5 Aprilie 1984, Paper F4, p. 407-421. [6] Popa, N.,The Reynolds equation solving for the constant central thickness hydrodynamic mechanical face seals, 3rd Vienna International Conference on Nano-Technology, march 18-20, 2009, Vienna, Austria [7] Etsion, I. and Auer, B.M. Simulation and Visualization of the Face Seal Motion Stability by Means of Computer Generated Movies. The 9 th International Conference on Fluid Sealing, 1-3 Aprilie 1984, Paper E1 p.153-162. [8]Salant, F.R. and Key, W.E. Development of an Analytical Model for Use in Mechanical Seal Design. The 10 th International Conference on Fluid Sealing, Innsbruck, 3-5 Aprilie 1984. [9]Schopplein, W., Zeus, O. Skliding Materials. The State of Art and Development Trends. International Sealing Conference 1986, Lenngries


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[10]Huitric, J., Bonneau, D., Tournerie, B. Finite Element Analysis of Grooved- Face Seals for Liquid. The 13 th International Conference on Fluid Sealing, 7-9 Aprilie,1992,Brugge, Belgium, [11]Dumbrava, M., Morariu, Z.,Consideraţii asupra regimului de frecare la etanşările frontale, 10-12 nov. 1988, I.P.Iaşi. [12]ZEUS, O. Cavitation Damages at Faces Materials of Mechanical Seal,Pump Congress Karlsruhe 92, p.84, 6-8 oct. 1992, p.84, Karlsruhe [13]Pascovici, M.D. and Etsion, I., The Accuaracy of the Isoviscous Solution of the Reynolds Equation in Mechanical Seals,Journal of Enegineering Tribology,1992 [14]Doane,J,C, Myrum,T, and Beard,J, An experimental-computational investigation of the heat transfer in mechanical face seals,International Journal of Heat and Mass Transfer,Vol.34,Issues 4-5,Pag.1027-1041,April-may 1991 [15]Zeus,D.,Viscous Friction in Small Gaps-Calculations for Non-Contacting Liquid or GasLubricated End Face Seals,STLE Tribology Trans., 33 (3), p. 454, 1990 [16]Buck,G.S. Heat Transfer in Mechanical Seals,6 th Int. Pump Users Symp.,April 24-28, 1989,Huston, Texas [17]Cicone,T.,Apostolescu,A., Temperature distribution in the rings of liquid face seals-evaluation of simplified analytical models using 3D FEM analysis,BALKANTRIB’O5, 5th International Conference on Tribology,June 15-18,2005, Kragujevac, Serbia and Montenegro [18] Tournerie,B.,Brunetière,N., Danos,J.C.,2D Numerical modeling of the TEHD transient behaviour of mechanical face seals,Proc. of the 17th BHRG Conf. on Fluid Sealing, 2003,York, UK.


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CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS

IN THE CANDU 6 NUCLEAR REACTOR. PART 6 - PRESENTATION OF THE DECOMMISSIONING DEVICE

Fiz. drd. Gabi ROSCA FARTAT,

Polytechnic University of Bucharest, [email protected], Ing. Constantin POPESCU,

Polytechnic University of Bucharest, [email protected], Prof. Univ. Emerit Dr. Ing. Constantin D. STANESCU,

Polytechnic University of Bucharest, [email protected]

ABSTRACT: The objective of this paper is to present a possible solution for the designing of a device for the decommissioning of the horizontal fuel channels in the CANDU 6 nuclear reactor. The decommissioning activities are dismantling, demolition, controlled removal of equipment, components, conventional or hazardous waste (radioactive, toxic) in compliance with the international basic safety standards on radiation protection. One as the most important operation in the final phase of the nuclear reactor dismantling is the decommissioning of fuel channels. For the fuel channels decommissioning should be taken into account the detailed description of the fuel channel and its components, the installation documents history, adequate radiological criteria for decommissioning guidance, safety and environmental impact assessment, including radiological and non-radiological analysis of the risks that can occur for workers, public and environment, the description of the proposed program for decommissioning the fuel channel and its components, the description of the quality assurance program and of the monitoring program, the equipments and methods used to verify the compliance with the decommissioning criteria, the planning of performing the final radiological assessment at the end of the fuel channel decommissioning. These will include also, a description of the proposed radiation protection procedures to be used during decommissioning. The dismantling of the fuel channel is performed by one device which shall provide radiation protection during the stages of decommissioning, ensuring radiation protection of the workers. The device shall be designed according to the radiation protection procedures. The decommissioning device assembly of the fuel channel components is composed of the device itself and moving platform support for coupling of the selected channel to be dismantled. The fuel channel decommissioning device is an autonomous device designed for dismantling and extraction of the channel closure plug and shield plug, extraction of the end fitting, cutting and extraction of the pressure tube. The fuel channel decommissioning device consists of following major components: coupling and locking fuel channel module, assembly valve for access to the fuel channel, storage tubes assembly for extracted components, handling elements assembly, cutting and extraction device and housing device. The design of the device and platform support is achieved according to the particular features of the fuel channel components to be dismantled in the program of nuclear reactor decommissioning according to all the safety aspects and environmental protection during the activities, resulting from the decommissioning plan developed. Key words: Candu reactor, device, decommissioning, dismantling, radiation protection, fuel channel

1. INTRODUCTION

The CANDU reactor decommissioning activities are dismantling, demolition, controlled removal of equipment, components, conventional or hazardous waste (radioactive, toxic) in compliance with the international basic safety standards on radiation protection.


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The decommissioning activities performed are administrative and technical and include the preparation, endorsement and approval of documents, obtaining permits and authorizations, providing financial resources, decontamination, dismantling, demolition, controlled removal of equipment, components, conventional or hazardous waste (radioactive, toxic), demonstrating the fulfillment of the radiological conditional or unconditional release of the facility and the ground included in the decommissioning project.

The dismantling of the fuel channel components is performed according to the nuclear reactor decommissioning documentation and the detailed schematic documentation of a the CANDU nuclear reactors fuel channel.

2. GENERAL PRESENTATION OF THE DEVICE

Many of the decommissioning activities involve the remote devices coordination to prevent the contact or some removed components proximity of the operators.

2.1 General considerations

Considering the fuel channel complexity, and that the designed operation life of a fuel channel is 30 years at 80% capacity and 24 years at maximum capacity, at the design of the channels fuel decommissioning device for shall be taken into account: - the detailed fuel channel description and its components; - the installation documents history from the operation period of the dismantled fuel channel; - adequate radiological criteria for decommissioning guidance; - safety and environmental impact assessment, including radiological and non-radiological analysis of the risks that can occur for workers, public and environment; - the proposed program description of the fuel channel decommissioning and its components; - the description of the quality assurance program; - the monitoring program, the equipments and methods used to verify the compliance with the decommissioning criteria; - the planning of performing the final radiological assessment at the end of the fuel channel decommissioning. Initial conditions for fuel channels decommissioning starting are the following: - there are no fuel bundles in the fuel channels; - the cooling system should be power off and the facility dismantled; - the feeders coupling of each feed pipes through which the cooling agent passes, located on the outside of each end fitting to be disassembled and the connection to be covered with a blind flange with four fastening screws and metallic safety lock against unscrewing; - the platform shall be in maintenance position for installation of the dismantling device.

Dismantling of the fuel channel components is performed when the initial conditions are carrier out. Channel status is exemplified in Figure 1.

Figure 1. Schematic representation of the fuel channel before dismantling


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2.2 Device assembly components presentation Device assembly for fuel channel components decommissioning is composed of the

device itself (1) and moving platform (2) that contains the device support assembly (3) for front alignment at the fuel channel (Figure 2).

1 2 3

Figure 2. Schematic representation of the device assembly components

The moving platform is necessary for moving in vertical and horizontal calandria plan

and positioning to the fuel channel which shall be decommissioned. The device support assembly is required for the positioning of the decommissioning

device at the fuel channel which shall be dismantled, for coupling of the channel.

2.3 Decommissioning device components presentation The decommissioning device for fuel channel components decommissioning is intended

for the following operations performed at the fuel channel: - the storage of the channel closure plug extracted from the end fitting; - the storage of the channel shield plug extracted from the fitting end; - the storage of the pressure tube extracted from the fuel channel; - the storage of the end fitting.

The decommissioning device for fuel channel components decommissioning consists of the coupling and locking fuel channel module (1), the access valve assembly to the fuel channel (2), the storage tubes assembly for extracted components (3), the handling elements assembly (4), the cutting and extraction device (5) and the housing device (6), exemplified in Figure 3.

1 2 3 4 5 6

Figure 3. Fuel channel decommissioning device components

The cutting and extraction device consists of the following modules: guiding-fixing

module (1), traction modules (2), guiding-fixing module at cutting (3), cutting module (4), guiding-extracting module (5) articulated elements (6) for modules connecting and command cable (Figure 4).


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5 2 3 4 3 2 1 7

6

Figure 4. Cutting and extraction device components

2.4 The coupling and locking fuel channel module The coupling and locking module is a stand-alone device, for coupling and fixing the

device at the fuel channel for performing the dismantling operations. The operation of the fuel channel coupling device is done manually by the operator. The coupling and locking module consists of the auxiliary closing piece (2), the locking cylinder (3), the safety seal (4) of the locking cylinder and is coupled to the fuel channel (1), exemplified in Figure 5.

4 3 2 1

Figure 5. The coupling and locking module to the fuel channel

After fuel channel module coupling, is mounted a protective cylindrical screen, made of

two semicircular pieces, closed with screws, covering the end fitting for the radiation protection of the operator, after extraction of the fuel channel end fitting (Figure 6).

Figure 6. The protective cylindrical screen for end fitting extracting

2.5 The access valve assembly The access valve is a structure which, by opening, enable the access of the cutting and

extraction device into the fuel channel to achieve the dismantling operations, and consist of the access valve itself (1) and the valve actuator (2), exemplified in Figure 7.

1 2

Figure 7. The access valve of the device


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2.6 The storage tubes assembly The storage tubes assembly is used to store the extracted components of the fuel

channel, as a result of the dismantling operations. This assembly consists of the radiation detector (1), the Blue tube (2), the Red tube (3), the Yellow tube (4), the Green tube (5) and the gearmotor drive (6), exemplified in Figure 8.

1 2 3 5 6 4

Figure 8. The storage tube assembly components

The storage tubes assembly is mounted on a shaft driven by a gearmotor to turning it in

order to place a tube in front of the access valve for access to the fuel channel. The storage tubes are used as follows:

- the Blue tube for storage of the pressure tube; - the Red tube for the fitting end storage; - the Yellow tube for storage of the channel closure plug and the channel shield plug; - the Green tube for storage of the extended channel closure plug.

2.7 The handling elements assembly The handling elements assembly is composed of the sleigh assembly (1), the sleigh

travel actuator (2), the stationary tube of the cutting and extraction device (3), connecting cable roller of the cutting and extraction device (4), the extracting actuator of the end fitting (5), exemplified in Figure 9.

1 2 3 4 5

Figure 9. The handling elements assembly

The handling elements assembly can operate two positions in order to place one element

front of the storage tube: - one position is when the stationary tube it is in working direction for the movement of the cutting and extraction device; cable roller realize the cable progress or tightening of the


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cutting and extraction device in time of displacement; the operations are the extraction of the channel closure plug, the channel shield plug, cutting and extraction of the pressure tube; - the second position is when the extracting actuator it is in working direction; the operations are extraction of the end fitting, installation or removal of the extended channel closure plug. 3. CONCLUSIONS

The decommissioning of the fuel channels is a complex process that requires piece by piece removal activities of components, transport and storage in dedicated facilities, preparation of records and documents specific decommissioning operations.

The presented device is a device that extracts the internal components of the horizontal fuel channels, ensuring a radiation protection during the stages of decommissioning.

The design of the device, moving platform and the device support assembly shall be achieved according to the particular features of the fuel channel components to be dismantled in the nuclear reactor decommissioning program, with respect of all security aspects, environmental protection during decommissioning activities and working procedures resulting from decommissioning plan developed.

4. REFERENCES [1] Cheadle B.A., Price E.G., “Operating performance of CANDU pressure tubes”, presented at

IAEA Techn. Comm. Mtg on the Exchange of Operational Safety Experience of Heavy Water Reactors, Vienna, 1989.

[2] Roger G. Steed, “Nuclear Power in Canada and Beyond”, Ontario, Canada, 2003. [3] Venkatapathi S., Mehmi A., Wong H., “Pressure tube to end fitting roll expanded

joints in CANDU PHWRS”, presented at Int. Conf. on Expanded and Rolled Joint Technology, Toronto, Canada, 1993.

[4] AECB, “Fundamentals of Power Reactors”, Training Center, Canada. [5] AECL, “CANDU Nuclear Generating Station”, Engineering Company, Canada. [6] ANSTO, “SAR CH19 Decommissioning”, RRRP-7225-EBEAN-002-REV0, 2004. [7] CANDU, “EC6 Enhanced CANDU 6 - Technical Summary”, 1003/05.2012. [8] CNCAN, “Law no. 111/1996 on the safe deployment, regulation, authorization and

control of nuclear activities”, 1996. [9] CNCAN, “Rules for the decommissioning of objectives and nuclear installations”, 2002. [10] IAEA, “Assessment and management of ageing of major nuclear power plant

components important to safety: CANDU pressure tube”, IAEA-TEDOC-1037, Vienna 1998.

[11] IAEA, “Assessment and management of ageing of major nuclear power plant components important to safety: CANDU reactor assemblies”, IAEA-TEDOC-1197, Vienna 2001.

[12] IAEA, “Decommissioning of Nuclear Power Plants and Research Reactors” Safety Standard Series No. WS-G-2.1, Vienna 1999.

[13] IAEA, “Nuclear Power Plant Design Characteristics, Structure of Power Plant Design Characteristics in the IAEA Power Reactor Information System (PRIS)”, IAEA-TECDOC-1544, Vienna 2007.


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[14] IAEA, “Organization and Management for Decommissioning of Nuclear Facilities”, IAEA-TRS-399, Vienna 2000.

[15] IAEA, “Selection of Decommissioning Strategy: Issues and Factors”, IAEA-TECDOC-1478, Vienna 2005.

[16] IAEA, “State of the Art Technology for Decontamination and Dismantling of Nuclear Facilities”, IAEA-TRS-395, Vienna 1999.

[17] IAEA, “Water channel reactor fuels and fuel channels: Design, performance, research and development”, IAEA-TEDOC-997, Vienna 1996.

[18] IAEA, “Heavy Water Reactor: Status and Projected Development”, IAEA-TEREP-407, Vienna 1996.

[19] Nuclearelectrica SA, “Cernavoda NPP Unit 1&2, Safety features of Candu 6 design and stress test summary report”, 2012.

[20] UNENE, Basma A. Shalaby, “AECL and HWR Experience”, 2010;


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IN THE CANDU 6 NUCLEAR REACTOR. PART 7 - FUNCTIONING OF THE DECOMMISSIONING DEVICE

Fiz. drd. Gabi ROSCA FARTAT,

Polytechnic University of Bucharest, [email protected], Ing. Constantin POPESCU,

Polytechnic University of Bucharest, [email protected], Prof. Univ. Emerit Dr. Ing. Constantin D. STANESCU,

Polytechnic University of Bucharest, [email protected]

ABSTRACT: The scope of this paper is to achieve the device functioning steps for the commissioning of the horizontal fuel channels of calandria vessel. The dismantling of the fuel channel is performed by one device which shall provide radiation protection during the stages of decommissioning, ensuring radiation protection of the workers. For the decommissioning operation design shall be taken to ensure all aspects of security, environmental protection during decommissioning operation steps and creating and implementing work procedures resulting from developed decommissioning plan. The fuel channel decommissioning device is designed for dismantling and extraction of the fuel channel and its components. The decommissioning operation consists of following major steps: platform with device positioning to the fuel channel to be dismantled; coupling and locking the device at the fuel channel; unblock, extract and store the channel closure plug; unblock, extract and store the channel shield plug; block and cut the middle and the end of the pressure tube; block, extract and store the end fitting; block, extract and store the half of pressure tube; mounting of the extended closing plug. The operations steps are performed by the Cutting and Extraction Device and by the extraction actuator from the device handling elements assembly. After each step of dismantling is necessary the confirmation its finalization in order to perform the next operation step. The dismantling operation steps of the fuel channel components are repeated for all the 380 channels of the reactor, from the front of calandria side (plane R) as well as the rear side (plane R'). Key words: Candu reactor, device, decommissioning, dismantling, radiation protection, fuel channel

1. INTRODUCTION

The dismantling of the fuel channel components is performed according to the nuclear reactor decommissioning documentation and the detailed schematic documentation of a the CANDU nuclear reactors fuel channel.

Many of the decommissioning activities involve the remote devices coordination to prevent the contact or some removed components proximity, of the operators.

Dismantling of the fuel channels represents the final phase of nuclear facility decommissioning and refers to the technical operations taken to extract the components from inside of the nuclear reactor channel. It is a complex process and requires activities such as assembly/disassembly decommissioning device, locking/unlocking the channel closure and the shield plug, pressure tube cutting, extracting of each component from the channel, as well as radioactive waste management.


173

2. GENERAL PRESENTATION OF THE DEVICE Many of the decommissioning activities involve the remote devices coordination to

prevent the contact of the operators with some removed components proximity.

2.1 General considerations The decommissioning device for the horizontal fuel channels, is a device allowing

retrieval of the internal fuel channels components of the horizontal calandria nuclear reactor, providing the biological protection and the containment of contamination. When the extraction step is completed, the decommissioning device is displaced to the transport container for transfer and storage the dismantled components in the dedicated facilities.

2.2 Device mounting on the moving platform

In order to positioning the device front of the channel and coupling to end fitting, for the fuel channel dismantling, it is necessary to mount the decommissioning device on a platform. With this platform can be performed the movement of the device, parallel with the plane of the reactor (horizontal and vertical movement), in order to position in front of one of the 380 fuel channels (Figure 1).

Figure 1. Schematic representation of the device mounting and positioning

2.2 Positioning assembly dismantling Before to start of the fuel channel decommissioning operations, should be manually

performed the positioning assembly dismantling, for all 380 fuel channels (see Figure 2). The dismantling operation stages of the positioning assembly should be repeated for all

the fuel channels, from the front of calandria side (plane R), as well as the rear side (plane R'). Before removal After removal

Figure 2. Schematic representation of the device mounting and positioning

2.3 Device coupling to the fuel channel

The coupling of the device to the fuel channel is performed manually by the operator. The coupling and fixing steps are (see Figure 3): - moving platform to position of the fuel channel to be dismantled;


174

- bring the container to the position that the coupling/extraction head is under the end fitting and then the container is lifted to a position in contact with the end fitting; - place the auxiliary piece (brown piece) for coupling/extraction head closing; - moving the locking cylinder (yellow piece) of the head of the coupling/extraction head in the "locked" position; Step 1 Step 2 Step 3 Step 4

Figure 3. Schematic representation of the device coupling steps

- mounting of the protective cylindrical screen, made from two semicircular pieces by screws joined, which cover the end fitting, for the operator radiation protection after the fuel channel end fitting extraction (see Figure 4);

Figure 4. Schematic representation of the protective cylindrical screen

2.4 Fuel channel components dismantling After the finalization of the device coupling and securing preparation operations to the

fuel channel it is possible to proceed to the fuel channel dismantling operations. The dismantling operations of the fuel channel components are performed on the operator panel of the device control panel, by the operator. 2.4.1 Fuel channel closure plug removal

The preliminary operations for the closure plug removal are: - the rotation command of the storage tube assembly so that the tube for the pressure tube storage (the blue tube) to reach the working position (coaxial with the axis of the fuel channel reactor); - the opening command of the device access valve assembly; - the movement command of the handling elements assembly that the stationary tube of the cutting and extraction device to reach the working position (see Figure 5);


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Figure 5. Schematic representation of the preliminary operations

After completing the preliminary operations, the operator command the cutting and

extraction device to moving, unlocking, extraction and storage of the channel closure plug in the yellow tube (see Figure 6).

Figure 6. Schematic representation of the channel closure plug extraction

2.4.2 Shield plug removal

After positioning of the blue tube in the working position, the operator command the cutting and extraction device to moving, unlocking, extraction and storage of the shield plug in the yellow tube (see Figure 7).

Figure 7. Schematic representation of shield plug extraction

2.4.3 Presure tube cutting For pressure tube cutting it is necessary to bring the blue tube in the working position.

The operator command the cutting and extraction device to move in the middle of the pressure tube, the positioning is performed by the encoder value (see Figure 8).

Figure 8. Schematic representation of the cutting device positioning

After cutting and extraction device positioning and fixing claws blocking (from the

guiding-fixing module), the operator can command the cutting module to start the cutting operation (see Figure 9). The cutting operation is monitored by video camera, for cutting viewing, and pyrometer for temperature recording in the cutting area.


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Figure 9. Schematic representation of the cutting operation The second step of the pressure tube cutting operation is the positioning of the cutting

and extraction device at the end of the pressure tube at the joint with the end fitting and blocking the guiding-fixing module claws (see Figure 10).

Figure 10. Schematic representation of the cutting device positioning

The operator command the cutting module to start the cutting operation (see Figure 9).

The cutting operation is monitored by video camera, for cutting viewing, and pyrometer for temperature recording in the cutting area.

After the end of the cutting operations, the cutting and extraction device is retreated in the stationary tube from the handling elements assembly.

2.4.4 End fitting dismantling

The preliminary operations to the end fitting extraction is performed by dragging the handling elements assembly that the extracting actuator of the end fitting and the red tube from the storage tubes assembly reach the working position (see Figure 11).

Figure 11. Schematic representation of the end fitting extraction preliminary operations

The operator command the extension of the extracting actuator until the coupling and

blocking with the end fitting. After coupling and blocking to the end fitting it is possible to command the withdrawal of the extracting actuator to the storing position in the red tube (see Figure 12).

Figure 12. Schematic representation of the coupling and extraction of the end fitting


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After performing the extraction of the end fitting, it is necessary to close the fuel channel, until the pressure tube extraction. To perform this operation, the operator should turn the storage tubes assembly until the green tube reach the working position (see Figure 13). In this tube is located the extended channel closure plug. When the green tube is in working position, the operator command the extension of the extracting actuator, to push from the green tube the extended channel closure plug until the closing of the fuel channel. After fuel channel closing, the extracting actuator is withdrawn to the handling elements assembly.

Figure 13. Schematic representation of the extended channel closure plug mounting

The next step is to close the access valve. The closing operation of the fuel channel is

necessary to ensure a radiation protection during the dismantling of the protective cylindrical screen (see Figure 14).

Figure 14. Schematic representation of the protective cylindrical screen dismantling

After installing of the extended channel closure plug, closing of the access valve and the

manually dismantling of the protective cylindrical screen, the operator can prepare the decommissioning device for the pressure tube extraction stage.

2.4.5 Presure tube extraction

The preliminary operations for the pressure tube extraction are: - manually coupling of the decommissioning device to the fuel channel; - protective cylindrical sleeve mounting; - the rotation command of the storage tube assembly so that the green tube to reach the working position; - the opening command of the device access valve assembly (see Figure 15);



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After completing the preliminary operations, the operator command the extension of the extracting actuator, to extract the extended channel closure plug and store it in the green tube (see Figure 16). After storage of the extended channel closure plug, the extracting actuator is withdrawn to the handling elements assembly.

Figure 16. Schematic representation of the extended channel closure plug extraction and storage

The operator command the movement the blue tube for the pressure tube storage from

the storage tube assembly and the handling elements assembly that the stationary tube of the cutting and extraction device to reach the working position (see Figure 17);

Figure 16. Schematic representation of the storage tube assembly and the handling elements assembly

positioning After positioning of the storage tube assembly and the handling elements assembly, the

operator can command the cutting module to extract and store the pressure tube into the blue tube from the storage tube assembly (see Figure 17).

Figure 17. Schematic representation of the extraction and storage of the pressure tube

The next step is to close again the fuel channel. To perform this operation, the operator

should turn the storage tubes assembly until the green tube, where is located the extended channel closure plug, and the extracting actuator from the handling elements assembly, reach the working position (see Figure 18).


The operator command the extension of the extracting actuator, to push from the green

tube the extended channel closure plug until the closing of the fuel channel. After fuel channel closing, the extracting actuator is withdrawn to the handling elements assembly (see Figure 19).


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Figure 19. Schematic representation of the extended channel closure plug mounting

The last operation after fuel channel closing is to close the access valve and withdrawal

of the decommissioning device from the front of the fuel channel. The closing operation of the fuel channel is necessary to ensure a radiation protection during the dismantling of the protective cylindrical sleeve (see Figure 20).

Figure 20. Schematic representation of the access valve closing and withdrawal of the decommissioning device

After dismantling of the fuel channel components, the charged decommissioning device

is moved with the moving platform to the transfer position, at the transport container, for the decommissioned materials storage transfer.

The dismantling operation stages of the fuel channel components are repeated for all the 380 channels of the reactor, from the front of calandria side (plane R) as well as the rear side (plane R').

3. CONCLUSIONS

The presented device is a device that extracts the internal components of the horizontal fuel channels, ensuring a radiation protection during the stages of decommissioning.

The decommissioning of the fuel channels is a complex process that requires piece by piece removal activities of the components, transport and storage in the dedicated facilities, records and specific documents preparation of the decommissioning operations.

The design of the device, moving platform and the device support assembly shall be achieved according to the particular features of the fuel channel components to be dismantled in the nuclear reactor decommissioning program, with respect of all security aspects, environmental protection during decommissioning activities and working procedures resulting from decommissioning plan developed.

4. REFERENCES

1. Cheadle B.A., Price E.G., “Operating performance of CANDU pressure tubes”, presented at IAEA Techn. Comm. Mtg on the Exchange of Operational Safety Experience of Heavy Water Reactors, Vienna, 1989.

2. Roger G. Steed, “Nuclear Power in Canada and Beyond”, Ontario, Canada, 2003. 3. Venkatapathi S., Mehmi A., Wong H., “Pressure tube to end fitting roll expanded


4. AECB, “Fundamentals of Power Reactors”, Training Center, Canada.


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5. AECL, “CANDU Nuclear Generating Station”, Engineering Company, Canada. 6. ANSTO, “SAR CH19 Decommissioning”, RRRP-7225-EBEAN-002-REV0, 2004. 7. CANDU, “EC6 Enhanced CANDU 6 - Technical Summary”, 1003/05.2012. 8. CNCAN, “Law no. 111/1996 on the safe deployment, regulation, authorization and

control of nuclear activities”, 1996. 9. CNCAN, “Rules for the decommissioning of objectives and nuclear installations”,

2002. 10. IAEA, “Assessment and management of ageing of major nuclear power plant


11. IAEA, “Assessment and management of ageing of major nuclear power plant components important to safety: CANDU reactor assemblies”, IAEA-TEDOC-1197, Vienna 2001.

12. IAEA, “Decommissioning of Nuclear Power Plants and Research Reactors” Safety Standard Series No. WS-G-2.1, Vienna 1999.

13. IAEA, “Nuclear Power Plant Design Characteristics, Structure of Power Plant Design Characteristics in the IAEA Power Reactor Information System (PRIS)”, IAEA-TECDOC-1544, Vienna 2007.

14. IAEA, “Organization and Management for Decommissioning of Nuclear Facilities”, IAEA-TRS-399, Vienna 2000.

15. IAEA, “Selection of Decommissioning Strategy: Issues and Factors”, IAEA-TECDOC-1478, Vienna 2005.

16. IAEA, “State of the Art Technology for Decontamination and Dismantling of Nuclear Facilities”, IAEA-TRS-395, Vienna 1999.

17. IAEA, “Water channel reactor fuels and fuel channels: Design, performance, research and development”, IAEA-TEDOC-997, Vienna 1996.

18. IAEA, “Heavy Water Reactor: Status and Projected Development”, IAEA-TEREP-407, Vienna 1996.

19. Nuclearelectrica SA, “Cernavoda NPP Unit 1&2, Safety features of Candu 6 design and stress test summary report”, 2012.

20. UNENE, Basma A. Shalaby, “AECL and HWR Experience”, 2010;


181

CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR THE DECOMMISSIONING OF THE HORIZONTAL FUEL

CHANNELS IN THECANDU 6 NUCLEAR REACTOR. PART 8 - PRESENTATION OF THE CUTTING AND

EXTRACTING DEVICE

Ing. Constantin POPESCU, Polytechnic University of Bucharest, [email protected],

Fiz. drd. Gabi ROSCA FARTAT, Polytechnic University of Bucharest, [email protected],

Prof. Univ. Emerit Dr. Ing. Constantin D. STANESCU, Polytechnic University of Bucharest, [email protected]

ABSTRACT: This paper present a constructive solution proposed by the authors in order to achieve of a cutting and extracting device for the decommissioning of the horizontal fuel channels in the CANDU 6 nuclear reactor. One of the most important part of the decommissioning device is the Cutting and Extraction Device (CED) which perform the dismantling, cutting and extraction of the fuel channel components. This flexible and modular device is designed to work inside the fuel channel. The main operations performed by the Cutting and Extraction Device (CED) are dismantling and extraction of the channel closure plug and shield plug, cutting and extraction of the pressure tube. The Cutting and Extraction Device (CED) consists of following modules: guiding-fixing module, traction modules, cutting module, guiding-extracting module and articulated elements for modules connecting. The guiding-fixing module is equipped with elastic guiding rollers and fixing claws in working position, the traction modules are provided with variable pitch rollers for allowing travel speed change through the fuel channel. The cutting module is positioned in the middle of the device and it is equipped with three roll knives for pressure tube cutting, having a system for cutting place video surveillance and pyrometers for cutting place monitoring temperature. The operations performed by the Cutting and Extraction Device (CED) of fuel channel are as follows: unblock and extract the channel closure plug, unblock and extract the channel shield plug, block and cut the middle of the pressure tube, block and cut the end of the pressure tube, block and extract the half of pressure tube. The Cutting and Extraction Device (CED) is fully automated, connected by wires to a Programmable Logic Controller (PLC) and controlled from a Human Machine Interface (HMI). The design of the Cutting and Extraction Device (CED) shall be achieved according to the particular features of the fuel channel components to be dismantled and to ensure radiation protection of workers.

Key words: Candu reactor, decommissioning, dismantling, radiation protection, fuel channel, cutting, extraction

1. INTRODUCTION

In the decommissioning process of a nuclear reactor CANDU-6, due to safety reasons, the protection measures of personal are required against the nuclear radiation, and using special decommissioning devices with command and control from the outside.


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2. GENERAL PRESENTATION OF THE DEVICE The decommissioning activities involve the remote devices coordination to prevent the

contact of the operators with some removed components proximity. The device presented hereunder, is a constructive solution proposed by the authors in

order to achieve of a cutting and extracting device (CED) for the decommissioning of the horizontal fuel channels in the CANDU 6 nuclear reactor.


The operations performed by the Cutting and Extraction Device (CED) of fuel channel (Figure 1) are as follows: unblock and extract the channel closure plug, unblock and extract the channel shield plug, block and cut the middle of the pressure tube, block and cut the end of the pressure tube, block and extract the half of pressure tube.

Fig.1. Schematic representation of the fuel channel before dismantling

The Cutting and Extraction Device (CED) is designed to be fully automated, connected

by wires to a Programmable Logic Controller (PLC) and controlled from a Human Machine Interface (HMI) from the decommissioning device.

Some characteristics and capabilities of the Cutting and Extraction Device (CED) device: - Length = 1320 mm - Outer diameter = 98 mm - Pipe inner diameter: minimum = 100 mm maximum = 110 mm - Pipe cutting thickness: up to 5 mm - Displacement velocity in pipe: 0.. 0.2 m/s

2.2 Device assembly components presentation

The Cutting and Extraction Device (CED) consists of following modules (see Figure 2): 1 - guiding-fixing module 2 - traction modules 3 - guiding-fixing modules at cutting 4 - cutting module 5 - guiding-extracting and connecting module 6 - flexible elements for modules connecting 7 - command cable


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5 2 3 4 3 2 1 7

6

Fig. 2. Cutting and Extraction Device (free and in pipe)

2.3. Guiding-extracting module The guiding-fixing module is equipped with elastic guiding rollers and fixing claws in

working position (Figure 3).

Fig. 3. Guiding-fixing module

This module is a self-adapted device to the differences of diameters along the pipe,

derived from thermal cycles, in reactor time life or other mechanical deformation (Figure 4). The fixing claws are piloted by an actuator and block device in the desired position.

a) b)

Fig.4. a) CED in a pipe with maximum inner diameter b) CED in a pipe with minimum inner diameter

The Cutting and Extraction Device (CED) has four such modules. This is helpful for:

fixing entire device inside the pipe, providing safety for cutting process, even if is a junction of two pipes with different diameter (see Figure 5).


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a) guiding rollers and fixing claws in “open” position b) guiding rollers and fixing claws in “close” position

Fig. 5. CED passing from thin tube at another one thicker -detail 2.4 Traction module

This module has three traction rollers. Before and after this module, in the CED structure, there is a guiding-fixing module, which allows the linear displacement (see Figure 6).

a) module – general view b) traction roller in c) traction roller in “closed” position “open” position

Fig. 6. Traction module

Each traction roller is an elastic system, with variable pitch. The traction module spins at a constant speed, linear speed variation coming from angle

variation on traction rollers (see Figure 7). The displacement is done by the principle “screw drives - nut”.

Fig.7. Traction roller position In the junction area CED movement is ensured by the traction module that is entirely in

pipe (see Figure 8).


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Traction module active Traction module passive

Fig.8. CED passing from thin tube at another one thicker – general view

2.5 Cutting module This is the most important module in the CED structure. It has three cutting rollers able

to cut up to 5 mm steel thickness. The pressure tube (PT) of CANDU 6 nuclear reactor, has 3 mm thickness.

These cutting rollers are pushed on the cutting surface by a system driven by an actuator (see Figure 9).

a) module – general view b) cutting roller cuts the pipe

Fig.9. Cutting module

Two of the rollers have in their proximity one pyrometer for temperature monitoring

during the cutting process (the red cone, see Figure 10).

Fig.10. Temperature monitoring, pyrometer cone (red)

The third cutting roller has a camera for video surveillance of the cutting process (the

magenta cone, see Figure 11).

Fig. 11. Video surveillance, camera cone (magenta)


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2.6 Guiding-extracting and connecting module This module is composed by two parts: the guiding-fixing module and connecting

module, the rigidly coupled together (see Figure 12).

Fig. 12. Guiding-extracting and connecting module

In the front of the connecting module there is a camera for video surveillance of the

connecting process (see Figure 13).

Fig.13. Frontal video camera

The connecting module has three fixing claws which are piloted by an actuator and

block the device in the connecting position with extracting plugs. The CED is prepared to extract the channel closure plug (see Figure 14) or the channel shield plug (see Figure 15).

Fig. 14. Extracting channel closure plug

Fig. 15. Extracting shield plug

2.7 Flexible elements for connecting modules The CED is a flexible device which should work inside the pipes. For to be able to pass

through the pipes, between modules there are flexible elements (see Figure 16) to allow the displacement along the pipes, even if there are deformations from the thermal cycles or from the mechanical tolerances.


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Fig.16. Flexible element

During the time life of the nuclear reactor, due to high temperature, can occur plastic

deformations of the pressure tube (PT) or lattice tube (LT). We should take in consideration that the decommissioning process is possible after 20-

25 years at the activity stop of the nuclear reactor.

2.8 COMMAND CABLE The Cutting and Extraction Device (CED) is fully automated, connected by wires (see

Figure 17) to a Programmable Logic Controller (PLC) and controlled from a Human Machine Interface (HMI).

Fig. 17. Command cable

3. CONCLUSIONS The design of the channels fuel decommissioning device should takes into account: - the detailed fuel channel description and its components; - safety and environmental impact assessment, including radiological and non-

radiological analysis of the risks that can occur for workers, public and environment. The Cutting and Extraction Device (CED) of the Candu fuel channel should become a

very helpful device in the decommissioning process, due to its capabilities and properties: - flexibility, command and control, safety; - excellent monitoring program, methods and equipment used to verify the compliance

with the decommissioning criteria; - this is a device that extracts the internal components of the horizontal fuel channels,

ensuring radiation protection during the stages of decommissioning.

4. REFERENCES 1. Cheadle B.A., Price E.G., “Operating performance of CANDU pressure tubes”, presented at


2. Roger G. Steed, “Nuclear Power in Canada and Beyond”, Ontario, Canada, 2003.


188

3. Venkatapathi S., Mehmi A., Wong H., “Pressure tube to end fitting roll expanded joints in CANDU PHWRS”, presented at Int. Conf. on Expanded and Rolled Joint Technology, Toronto, Canada, 1993.

4. AECB, “Fundamentals of Power Reactors”, Training Center, Canada. 5. AECL, “CANDU Nuclear Generating Station”, Engineering Company, Canada. 6. ANSTO, “SAR CH19 Decommissioning”, RRRP-7225-EBEAN-002-REV0, 2004. 7. CANDU, “EC6 Enhanced CANDU 6 - Technical Summary”, 1003/05.2012. 8. CNCAN, “Law no. 111/1996 on the safe deployment, regulation, authorization and















189

CONSIDERATIONS FOR THE DEVELOPMENT OF A DEVICE FOR

THE DECOMMISSIONING OF THE HORIZONTAL FUEL CHANNELS IN THE CANDU 6 NUCLEAR REACTOR.

PART 9 - CUTTING AND EXTRACTING DEVICE FUNCTIONING




ABSTRACT: This paper presents a constructive solution proposed by the authors in order to achieve of a cutting and extracting device for the decommissioning of the horizontal fuel channels in the CANDU 6 nuclear reactor. The Cutting and Extraction Device (CED) performs the dismantling, cutting and extraction of the fuel channel components. It's a flexible and modular device, which is designed to work inside the fuel channel and has the following functions: moving with variable speed, temperature monitoring and video surveillance inside the pipe, unblock and extract the channel closure plug (from End Fitting - EF), unblock and extract the channel shield plug (from Lattice Tube - LT), block and cut the middle of the pressure tube, block and cut the end of the pressure tube, block and extract the half of pressure tube. The Cutting and Extraction Device (CED) consists of following modules: guiding-fixing module, traction modules, cutting module, guiding-extracting module and flexible elements for modules connecting. The guiding-fixing module is equipped with elastic guiding rollers and fixing claws in working position, the traction modules are provided with variable pitch rollers for allowing variable travel speed through the fuel channel. The cutting module is positioned in the middle of the device and it is equipped with three knife rolls for pressure tube cutting, using a system for cutting place video surveillance and pyrometers for monitoring cutting place temperature. The Cutting and Extraction Device (CED) is fully automated, connected by wires to a Programmable Logic Controller (PLC) and controlled from a Human Machine Interface (HMI). The design of the Cutting and Extraction Device (CED) shall be achieved according to the particular features of the fuel channel components to be dismantled and to ensure radiation protection of workers. Key words: Candu reactor, decommissioning, dismantling, radiation protection, fuel channel, cutting, extraction

1. INTRODUCTION In the decommissioning process of a nuclear reactor CANDU-6, due to safety reasons, the

protection measures of personal are required against the nuclear radiation, and using special devices with command and control from the outside. The CED should perform all operations needed, as: unblock and extract the plugs, block into the pipes, cut and extract the pressure tube.


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2. GENERAL FUNCTIONING PRESENTATION The decommissioning activities involve the remote devices coordination to prevent the

contact of the operators with some removed components proximity. The operations performed by the Cutting and Extraction Device (CED) of fuel channel

are as follows: unblock and extract the channel closure plug, unblock and extract the channel shield plug, block and cut the middle of the pressure tube, block and cut the end of the pressure tube, block and extract the half of pressure tube.

2.1 Preliminary operations The preliminary operations to be performed before functioning of the Cutting and

Extraction Device (CED) are (Figure 1): - decommissioning device mounting on the moving platform (1); - moving platform to position of the fuel channel to be dismantled (2); - positioning assembly dismantled (3); - decommissioning device coupled to the fuel channel (4) and protective cylindrical screen mounted; 1 2 3 4

Figure 1. Representation of the preliminary operations before dismantling

Dismantling of the fuel channel components by the Cutting and Extraction Device

(CED) is performed when the initial operations are carrier out. Channel status is exemplified in Figure 2.

Figure 2. Representation of the fuel channel before dismantling

2.2 Fuel channel closure and shield plug removal The preliminary operations to the decommissioning device for the fuel channel closure

and shield plug removal by the Cutting and Extraction Device (CED) are (Figure 3): - the handling elements assembly movement that the stationary tube (1) with the Cutting and Extraction Device (CED) to reach the working position; - the storage tube assembly rotation that the tube for the pressure tube storage (the blue tube (2)) to reach the working position (coaxial with the axis of the fuel channel reactor); - the access valve assembly (3) opening;


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1 2 3

Figure 3. Representation of the preliminary operations

After completing the preliminary operations, the Cutting and Extraction Device (CED) is

ready to move, to unlock, to extract (1) and storage (2) of the channel closure plug in the yellow tube. The next operation for the Cutting and Extraction Device (CED) is to move, to unlock, to extract and storage of the shield plug in the yellow tube (see Figure 4). 1 2 3 4

Figure 4. Representation of the channel closure and shield plug extraction and storage

2.3 Fuel channel presure tube cutting The first step for the pressure tube cutting is to move the Cutting and Extraction Device

(CED) in the middle of the pressure tube (1) and fixing the claws blocking (3) from the guiding-fixing module (2) exemplified in Figure 5. 1 2 3

Figure 5. Representation of the cutting device positioning and fixing

After device positioning and fixing claws blocking (from the guiding-fixing module),

can start the cutting operation with the cutting module (1). In the cutting operation time (2), the operation is monitored by pyrometer (3) for temperature recording in the cutting area and by video camera (4), for cutting operation viewing (see Figure 6).


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1 2 3 4

Figure 6. Representation of the cutting operation

The second step for the pressure tube cutting is to move the Cutting and Extraction

Device (CED) (2) at the end of the pressure tube at the joint with the end fitting (1) and fixing the claws blocking from the guiding-fixing module (Figure 7). 1 2

Figure 7. Representation of the second cutting

2.4 Fuel channel presure tube extraction The pressure tube extraction should be done after the both cutting of the pressure tube, in

the middle and at the end of the pressure tube at the joint with the end fitting. The extracting operation is to move the Cutting and Extraction Device (CED) (1) at the end of the pressure tube at the joint with the end fitting, so that the guiding-fixing module go into the pressure tube and fix the claws blocking (Figure 8). The half of the pressure tube is withdrawn (2) in the decommissioning device and stored into the blue tube (3) of the storage tube assembly. 1 2 3

Figure 8. Representation of the extraction and storage of the pressure tube

After the pressure tube storage the Cutting and Extraction Device (CED) (1) is retracted

into the stationary tube from the decommissioning device handling elements assembly (2), exemplified in Figure 9.


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1 2

Figure 18. Representation of the CED retracting into the stationary tube

3. CONCLUSIONS The Cutting and Extraction Device (CED) device is a device that performs the cutting

and extracting of the internal components from the horizontal fuel channels nuclear reactor and ensure a radiation protection during the stages of decommissioning.

Due to its capabilities and properties, flexibility, remote command and control of the functions, temperature and video monitoring, the Cutting and Extraction Device (CED) should become a very helpful device in the Candu fuel channel decommissioning process.

4. REFERENCES 1. Cheadle B.A., Price E.G., “Operating performance of CANDU pressure tubes”, presented at











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IN THE CANDU 6 NUCLEAR REACTOR. PART 10 - PRESENTATION OF THE DECOMMISSIONING DEVICE

OPERATING




ABSTRACT: This paper presents a solution proposed by the authors in order to achieve of a cutting and extracting device operating panel for the decommissioning of the horizontal fuel channels in the CANDU 6 nuclear reactor. The Cutting and Extraction Device (CED) is fully automated, connected by wires to a Programmable Logic Controller (PLC) and controlled from a Human Machine Interface (HMI). The Cutting and Extraction Device (CED) performs the dismantling, cutting and extraction of the fuel channel components, moving with variable speed, temperature monitoring and video surveillance inside the pipe, unblock and extract the channel closure plug (from End Fitting - EF), unblock and extract the channel shield plug (from Lattice Tube - LT), block and cut the middle of the pressure tube, block and cut the end of the pressure tube, block and extract the half of pressure tube. All operations can be monitored and controlled from a operating panel. The PLC fully command the device in automatic or manually mode, to control the internal sensors, transducers, electrical motors, video surveillance and pyrometers for monitoring cutting place temperature. The device controller has direct access to the measured values with these sensors, interprets and processes them, preparing the next actionafter confirming the action in progress. The design of the Cutting and Extraction Device (CED) shall be achieved according to the particular features of the fuel channel components to be dismantled and to ensure radiation protection of workers. Key words:Candu reactor, fuel channel, decommissioning device, dismantling, operator panel, cutting, extraction

1. INTRODUCTION

The CANDU reactor decommissioning activities involve the remote devices coordination to prevent the contact or some removed components proximity, of the operators, in compliance with the international basic safety standards on radiation protection.

The decommissioning of the fuel channels is a complex process that requires piece by piece removal activities of components, transport and storage in dedicated facilities. Many of the decommissioning activities involve the remote devices coordination to prevent the contact of the operators with some removed components proximity.


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2.GENERAL OPERATING PRESENTATION OF THE DEVICE The decommissioning device is a electromechanical system with many freedom degrees,

capable to perform the extraction of the internal components of the horizontal fuel channels. All operations are performed under the control of a system equipped with a Programmable Logic Controller (PLC)and monitored by an operator panel type Human Machine Interface (Touch Screen HMI). The system is automated, each operation step shall be confirmed by the operator after its finalization, to preparing the next operation step.


The device assembly for the fuel channel components decommissioning is composed of the device itself and moving platform that contains the device support assembly for front alignment at the fuel channel. The platform support, for moving and positioning device front of the fuel channel, is an assembly that can be moved vertically and horizontally, in parallel plane front of calandria, to the fuel channel to be decommissioned. The platform support for moving and positioning device has 2 freedom degrees due to the X-axis movement (horizontal movement) and Z axis (vertical movement), shown in Figure 1. Z axis X axis

Figure 1. Schematic representation of the platform support degrees of freedom

The cutting and extraction device perform a forward/retreat movement along the Y axis

and a rotational movement along the Y axis, so that it has two freedom degrees, as exemplified in Figure 2.

Yaxis Rotation

Figure 2. Schematic representation of the cutting and extraction device freedom degrees

In conclusion, the decommissioning device assembly can be compared to an industrial

robot capable to execute handling operations, autonomously and automatically, under a control system, which has 4 freedom degrees, three movements on the X, Y and Z axis, as well as a rotational movement around the Y axis.


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2.2 Automation block diagram of the device The automation block diagram of the decommissioning device is exemplified in Figure

3.

Access Valve (90º Actuator, limit switches for open/close)

Storage Tubes Assembly (Motors, encoder, proximity switches )

Radiation Detector

Handling Elements Assembly (Actuators, proximity switches)

Extraction Actuator (Actuators, proximity switches)

Cutting and Extracting Device (Motors, actuators, encoders, limit switches, proximity switches, video cameras , pyrometers)

Connecting Cable Roller for Cutting and Extracting Device(Motor, encoder, tensioning limiter, proximity switches)

Device automation panel (Programmable Logic

Controller, Human Machine Interface)

Radiation level monitoring operator panel

Moving Device Platform (Motors, encoder, limit switches )

Figure 3. Schematic representation of the automation main functional blocks The main functional blocks of the fuel channels decommissioning device are:

- control unit for operation and parameters visualization (equipped with PLC and Touch Screen HMI );


198

- control unit of the moving platform (equipped with motors, encoders, limit switches ); - control unit of the access valve opening/closing (equipped with actuator, proximity switches); - control unit of the storage tubes assembly rotating (equipped with motors, encoders, limit switches); - control unit of the handling elements assembly movement (equipped with actuators, proximity switches ); - control unit for the connecting cable roller of the cutting and extracting device; - control unit of the cutting and extracting device (equipped with motors, actuators, encoders, limit switches, position switches, video camera, pyrometers); - control unit for the radiation level monitoring of the dismantled components (equipped with radiation sensor and parameters visualization operator panel);

2.3 Device hmi operator panel operating

The decommissioning device operations are performed under the control of a system equipped with a Programmable Logic Controller (PLC), monitored by an operator panel type Human Machine Interface (HMI) and for the radiation level monitoring by an specialized operator panel.

In the Touch Screen HMI operator panel are designed, using specialized software, viewing and operating screens of the decommissioning stages. The operating structure screens are designed on three operating levels.

2.3.1 Level 1 - startup screen

The Level 1 in the Touch Screen HMI operator panel arerepresented by main screen (Figure 4). In this page it is possible to turn to the level 2 screens operation, the calandria stage operations screen, the decommissioning device stage operations and the alarms screen.

Figure 4. Schematic representation of the HMI startup homepage


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2.3.2 Level 2 screens The Level 2 screens in consists of the following screens:

1. The calandria fuel channels plan screen. In this page the operator performs the following operations (Figure 5):

- selection of the front or rear of the calandria; - selection of the fuel channel to be dismantled; - vertical and horizontal movement of the platform device to the fuel channel selected, movement monitored by encoder value; - confirmation of the performed operation to preparing the next operation step.

Figure 5. Schematic representation of the calandria fuel channels plan screen

2. The general decommissioning device screen.

In this page the operator performs the following preliminary operations (Figure 6): - coupling and confirmation of the manual coupling operation finishing; - opening/closing of the device access valve; - tube selection by turning of storage tubes assembly in the working position; - execution element selection by movement of the handling elements assembly in the working position; - screen selection for working with one of the handling elements; - selection of the cutting and extraction device screen for plugs extraction and cutting operations and the extracting actuator screen for end fitting extraction extended channel closure plug mounting.


200

Figure 6. Schematic representation of the general decommissioning device screen

3. The general alarms screen.

In this screen the operator visualize and acknowledges the faults appearance when the fuel channel decommissioning operations are performed (Figure 7).

Figure 7. Schematic representation of the general alarms screen

The 5 classes of alarms, "Warnings", "Urgent", "High", "Medium" and "Low Errors" are

exemplified in Figure 8.


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Figure 8. Schematic representation of the alarms classes

2.3.3. Level 3 screens The Level 3 screens in consists of the following screens:

1. The extraction and cutting device screen. In this page the operator performs the following operations with the extraction and

cutting device (Figure 9): - selection of the operation step to be performed; - moving of the extraction and cutting device to the set-point value position; - blocking the extraction and cutting device in the working position; - extraction of the channel closure plug from the fuel channel; - extraction of the shield plug from the fuel channel; - cutting of the pressure tube in the middle and at the end of the pressure tube at the joint with the end fitting; - extraction of the pressure tube from the fuel channel and storage in the storage tubes assembly; - confirmation of each finalization step in order to perform the next operation step.

Figure 9. Schematic representation of the cutting and extraction device screen


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2. The extracting actuator screen.

In this page the operator performs the following operations with the extracting actuator (Figure 10): - selection of the operation step to be performed; - moving of the extracting actuator to the set-point value position; - extraction of the end fitting from the fuel channel; - mounting / removal of the extended channel closure plug to / from the fuel channel; - confirmation of each finalization step in order to perform the next operation step.

Figure 10. Schematic representation of the extracting actuator screen

2.3.4 Radiation level monitoring screen

The radiation level monitoring operator panel, mounted on the front of the control cabinet and connected to the radiation sensor, is necessary for the radiation level monitoring of the each extracted components from the fuel channel (Figure 11). It performs radiation level measurement, parameters calculating, reading and writing the parameters in the memory panel, achieving radiation levels tables for each fuel channel decommissioned component and PC coupling for data transmission, statistics achieving and them archiving.

Figure 11. Schematic representation of the radiation level monitoring


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2.3.5 Video displays monitoring For monitoring the decommissioning operations of the fuel channel components are

required for video cameras for the dismantling steps surveillance performed by the cutting and extraction device, and the extraction actuator (Figure 12). The supervision of these operations need three cameras mounted in the following positions: - a video camera in the cutting and extraction device, into the cutting module for monitoring the process of the pressure tube cutting; - a video camera in the cutting and extraction device, into the extraction module for monitoring the extraction channel closure plug, shield plug and pressure tube; - a video camera in the device for monitoring the extraction of the extended channel closure plug.

Figure 12. Schematic representation of the video display monitoring

3. CONCLUSIONS

The process of the fuel channels decommissioning requires piece by piece removal activities of the components, radiation protection of the operators during the dismantling, transport and storage in the dedicated facilities, records and specific documents preparation of the decommissioning operations.

The mechanical design, automation and software programsof the device shall be achieved according to the particular features of the fuel channel components to be dismantled in the nuclear reactor decommissioning program, with respect of all security aspects, working procedures, ensuring a radiation protectionof the operating personnel, environmental protection during the stages of decommissioning activities resulting from decommissioning plan developed.

4. REFERENCES

1. Cheadle B.A., Price E.G.,“Operating performance of CANDU pressure tubes”, presented at IAEA Techn. Comm. Mtg on the Exchange of Operational Safety Experience of Heavy Water Reactors, Vienna, 1989.

2. Roger G. Steed,“Nuclear Power in Canada and Beyond”, Ontario, Canada, 2003.


204

3. Venkatapathi S., Mehmi A., Wong H.,“Pressure tube to end fitting roll expanded joints in CANDU PHWRS”, presented at Int. Conf. on Expanded and Rolled Joint Technology, Toronto, Canada, 1993.

4. AECB,“Fundamentals of Power Reactors”, Training Center, Canada. 5. AECL,“CANDU Nuclear Generating Station”, Engineering Company, Canada. 6. ANSTO,“SAR CH19 Decommissioning”, RRRP-7225-EBEAN-002-REV0, 2004. 7. CANDU,“EC6 Enhanced CANDU 6 - Technical Summary”, 1003/05.2012. 8. CNCAN,“Law no. 111/1996 on the safe deployment, regulation, authorization and

control of nuclear activities”, 1996. 9. CNCAN,“Rules for the decommissioning of objectives and nuclear installations”,

2002. 10. IAEA,“Assessment and management of ageing of major nuclear power plant


11. IAEA,“Assessment and management of ageing of major nuclear power plant components important to safety: CANDU reactor assemblies”, IAEA-TEDOC-1197, Vienna 2001.

12. IAEA,“Decommissioning of Nuclear Power Plants and Research Reactors” Safety Standard Series No. WS-G-2.1, Vienna 1999.

13. IAEA,“Nuclear Power Plant Design Characteristics, Structure of Power Plant Design Characteristics in the IAEA Power Reactor Information System (PRIS)”, IAEA-TECDOC-1544, Vienna 2007.

14. IAEA,“Organization and Management for Decommissioning of Nuclear Facilities”, IAEA-TRS-399, Vienna 2000.

15. IAEA,“Selection of Decommissioning Strategy: Issues and Factors”, IAEA-TECDOC-1478, Vienna 2005.

16. IAEA,“State of the Art Technology for Decontamination and Dismantling of Nuclear Facilities”, IAEA-TRS-395, Vienna 1999.

17. IAEA,“Water channel reactor fuels and fuel channels: Design, performance, research and development”, IAEA-TEDOC-997, Vienna 1996.

18. IAEA,“Heavy Water Reactor: Status and Projected Development”, IAEA-TEREP-407, Vienna 1996.

19. Nuclearelectrica SA,“Cernavoda NPP Unit 1&2, Safety features of Candu 6 design and stress test summary report”, 2012.



205


IN THECANDU 6 NUCLEAR REACTOR. PART 11 - PRESENTATION OF THE CUTTING AND EXTRACTING

DEVICE OPERATING



Fiz. drd. Gabi ROSCA FARTAT, Polytechnic University of Bucharest, [email protected]

ABSTRACT: This paper presents a constructive solution proposed by the authors in order to achieve of a cutting and extracting device for the decommissioning of the horizontal fuel channels in the CANDU 6 nuclear reactor. The Cutting and Extraction Device (CED) performs the dismantling, cutting and extraction of the fuel channel components. It's a flexible and modular device, which is designed to work inside the fuel channel and has the following functions: moving with variable speed, temperature monitoring and video surveillance inside the pipe, unblock and extract the channel closure plug (from End Fitting - EF), unblock and extract the channel shield plug (from Lattice Tube - LT), block and cut the middle of the pressure tube, block and cut the end of the pressure tube, block and extract the half of pressure tube. The Cutting and Extraction Device (CED) consists of following modules: guiding-fixing module, traction modules, cutting module, guiding-extracting module and flexible elements for modules connecting. The guiding-fixing module is equipped with elastic guiding rollers and fixing claws in working position, the traction modules are provided with variable pitch rollers for allowing variable travel speed through the fuel channel. The cutting module is positioned in the middle of the device and it is equipped with three knife rolls for pressure tube cutting, using a system for cutting place video surveillance and pyrometers for monitoring cutting place temperature. The Cutting and Extraction Device (CED) is fully automated, connected by wires to a Programmable Logic Controller (PLC) and controlled from a Human Machine Interface (HMI). The design of the Cutting and Extraction Device (CED) shall be achieved according to the particular features of the fuel channel components to be dismantled and to ensure radiation protection of workers. Key words: Candu reactor, decommissioning, dismantling, radiation protection, fuel channel, cutting, extraction

1. INTRODUCTION The decommissioning activities of the CANDU 6 reactor, involve the remote devices

coordination helpful from safety reason and also from operating process point of view. This complex process requires piece by piece removal, transport and dedicated storage.

The decommissioning device assembly is a complex one and has three principal parts: - a flexible platform that contains the device support assembly for front alignment at the fuel channel;


206

- a complex command and storage decommissioning device; - Cutting and Extraction Device (CED).

The Cutting and Extraction Device (CED), part of the decommissioning device, should perform all operations needed, as: unblock and extract the plugs, block into the pipes, cut and extract the pressure tube.

2. GENERAL PRESENTATION OF THE DEVICE

The device presented hereunder, is a constructive solution proposed by the authors in order to achieve of a Cutting and Extraction Device (CED) for the decommissioning of the horizontal fuel channels in the CANDU 6 nuclear reactor.


The operations performed by the Cutting and Extraction Device (CED) of fuel channel are as follows: unblock and extract the channel closure plug, unblock and extract the channel shield plug, block and cut the middle of the pressure tube, block and cut the end of the pressure tube, block and extract the half of pressure tube.

The Cutting and Extraction Device (CED) is fully automated, connected by wires to a Programmable Logic Controller (PLC) and controlled from a Human Machine Interface (HMI), and has some characteristics and capabilities as follow: - CED Length = 1320 mm; - CED Outer diameter = 98 mm; - Pipe inner diameter: minimum = 100 mm maximum = 110 mm; - Pipe cutting thickness: up to 5 mm; - Displacement velocity in pipe: 0.. 0.2 m/s. 2.2 Device assembly components presentation CED is an assembly of modules connected together by elastic elements to also ensure the functioning inside pipes with small deformations as in those that make part from CANDU 6 reactor, Should be considered deformations from positioning and from thermal cycles during the long life time reactor, too. 5 2 3 4 3 2 1 7

6

Fig. 1 Cutting and Extraction Device (free and in pipe)


207

The Cutting and Extraction Device (CED) consists of following modules (see Fig. 1): 1 - guiding-fixing module 2 - traction modules 3 - guiding-fixing modules at cutting 4 - cutting module 5 - guiding-extracting and connecting module 6 - flexible elements for modules connecting 7 - command cable

The cutting and extraction device perform a forward/retreat movement along the Y axis and a rotational movement along the Y axis, so that it has two freedom degrees, as exemplified in Figure 2. Y axis Rotation

Figure 2. Schematic representation of the cutting and extraction device degrees of freedom

2.3 Automation block diagram of the device

The system is designed to control the actuators and sensors, to supervise the proper performance of the operations execution.

The automation block diagram of the decommissioning device is exemplified in Figure 3.

Cutting and Extracting Device (Motors, actuators, encoders, limit switches, proximity switches, video cameras , pyrometers)

Connecting Cable Roller for Cutting and Extracting Device(Motor, encoder, tensioning limiter, proximity switches)

Device automation panel (Programmable Logic

Controller, Human Machine Interface)

Figure 3. Schematic representation of the automation functional blocks


208

The main functional blocks of the fuel channels decommissioning device are: - control unit for operation and parameters visualization ( PLC and Touch Screen HMI from the decommissioning device ); - control unit of the cutting and extracting device (equipped with motors, actuators, encoders, limit switches, position switches, video camera, pyrometers); - control unit for the connecting cable roller of the cutting and extracting device;

2.4 Device hmi operator panel operating The PLC fully command the device in automatic or manually mode, to control the internal sensors, transducers, electrical motors, video surveillance and pyrometers for monitoring cutting place temperature.

In the decommissioning Touch Screen HMI operator panel are designed, using specialized software, viewing and operating screens of the Cutting and Extraction Device operating stages.

The main page for Cutting and Extraction Device operating is exemplified in Figure 4.

Figure 4. Schematic representation of the cutting and extraction device screen

In this page the operator performs the following operations with the extraction and

cutting device (Figure 8): - selection of the operation step to be performed; - moving of the extraction and cutting device to the set-point value position; - blocking the extraction and cutting device in the working position; - extraction of the channel closure plug from the fuel channel; - extraction of the shield plug from the fuel channel; - cutting of the pressure tube in the middle and at the end of the pressure tube at the joint with the end fitting;


209

- extraction of the pressure tube from the fuel channel and storage in the storage tubes assembly; - confirmation of each finalization step in order to perform the next operation step.

In the general alarms screen the operator visualize and acknowledges the faults appearance when the fuel channel decommissioning operations are performed (Figure 5).

Figure 5. Schematic representation of the general alarms screen

2.5 Video displays monitoring

For monitoring the decommissioning operations of the fuel channel components are required for video cameras for the dismantling steps surveillance performed by the cutting and extraction device.

The supervision of this operation need a video camera in the cutting and extraction device, into the cutting module for monitoring the process of the pressure tube cutting (the magenta cone, Figure 6).

Figure 6. Video surveillance, camera cone (magenta)

In the front of the connecting module there is a camera for video surveillance of the

connecting process (see Figure 7). It also used for:


210

- monitoring the extraction channel closure plug, shield plug and pressure tube; - monitoring the extraction of the extended channel closure plug.

Figure 7. Frontal video camera

3. CONCLUSIONS

The mechanical design, automation and software programs of the device should takes into account: - the detailed fuel channel description and its components; - safety and environmental impact assessment, including radiological and non-radiological analysis of the risks that can occur for workers, public and environment.

The Cutting and Extraction Device (CED) of the fuel channel decommissioning device should become a very helpful device in the decommissioning process, due to its capabilities and properties: - flexibility, command and control, safety; - excellent monitoring program, methods and equipment used to verify the compliance with the decommissioning criteria; - this is a device that extracts the internal components of the horizontal fuel channels, ensuring radiation protection during the stages of decommissioning.

Due to performance monitoring system, this device could be used to improve the next generation of this kind of devices, analyzing each step and the aspects from DFMEA analysis of device.

4. REFERENCES

1. Cheadle B.A., Price E.G., “Operating performance of CANDU pressure tubes”, presented at IAEA Techn. Comm. Mtg on the Exchange of Operational Safety Experience of Heavy Water Reactors, Vienna, 1989.





2002.


211

10. IAEA, “Assessment and management of ageing of major nuclear power plant components important to safety: CANDU pressure tube”, IAEA-TEDOC-1037, Vienna 1998.












212

CONSIDERATIONS REGARDING THE FRETTING PHENOMENON USING LEAF SPRINGS

Stefan GHIMIȘI, Constantin Brâncuși University of Targu Jiu,[email protected]

Abstract. The fretting phenomenon represents particulary and complex form of wear who is; generaly, and/or weary of fretting who is produced on the load contact in a relative oscialatory movement lay small amplitude.A simultaneoustly applied tangential force and normal into contact appears a adhesion force Keywords: fretting, oscillatory motion, elastic blade 1. INTRODUCTION

Fretting is now fully identified as a small amplitude oscillatory motion which induces a harmonic tangential force between two surfaces in contact. It is related to three main loadings i.e. fretting-wear, fretting-fatigue and fretting corrosion.

The main parameters were reported to be amplitude displacement, normal load, frequency, surface roughness and morphology, and residual stresses. More recently fretting has been discussed using the third-body concept and using the means of the velocity accommodation mechanisms introduced by Godet et al.[1,2]

Fretting regimes were first mapped by Vingsbo. In a similar way, three fretting regimes will be considered: stick regime, slip regime and mixed regime. The mixed regime was made up of initial gross slip followed by partial slip condition after a few hundred cycles. Obviously the partial slip transition develops the highest stress levels which can induce fatigue crack nucleation depending on the fatigue properties of the two contacting first bodies. Therefore prediction of the frontier between partial slip and gross slip is required. The type of surface damage that occurs in fretting contact depends on the magnitude of the surface normal and tangential tractions. In existing fretting models the relative displacement is assumed to be accommodated mainly micro slip in the contact surface.

2. EXPERIMENTAL MEANS For the study of the fretting phenomenon in case of elastics assemblages spring slides with multiple sheets, I used the experimental stall from fig.1. .[3] The stall permits testing for one slide and for spring slides with multiple sheets, too.

2.1. Description of the stall The elastic blade is assembled on the rigid support, using the upper plate with screws.

It oscillates because of the rod and crank eccentric mechanism. This mechanism is actioned by the electrical engine, assuring the necessary conditions for the fretting phenomenon to take place. The contact is charged with the assistance of 4 screws through the agency of some helical springs and through the agency of some radial-axial bearings with conic rolls. The helical springs beforehand standard permit a charge with a normal and known force, the presence of the radial-axial bearings assuring the discharge of friction between the screw and the superior plate.


213

Fig.1. Experimental stall Fig.2.The system eccentricity

The stall can be used for the testing at fretting of some couples by different materials. This stall can be adapted for study of the lamellar springs with many sheets. The blades used in experiments have 560x56x2 mm and are realized by spring steel having hardness 55 HRC. The rod-crank mechanism permits a displacement at the end (extremity) of the 20 mm lamella and can modify this displace by changing of the system eccentricity (Fig.2). The system is actioned through the agency of electrical engine having revolution of 750 rot/min.

The experimental stand was used in order to study the wear state caused by the blade to blade contact, specific for leaf springs with multiple sheets, by a low-amplitude oscillatory motion. In this case I have studied the dependence of the fretting phenomenon of normal push force and a certain request length, obtaining for each case specific traces. At a variation of the push normal force between 200 and 250 N, the duration of exposure ranged between 40000 and 60000 stress cycles[4].

Traces of wear obtained from tests at a normal force of 250 N, reveal, as expected, an increase in the size of the used area, at an increased request length. Also, all traces of fretting wear tests were identified by the presence of "red powder", at the contact between the two blades.

The marks obtained for 250 N and 40000 cycles are shown in Figure 3 for the same normal force and for 50,000 to 60,000 cycles traces obtained are given in Figure 4 and 5.


214

Fig.3.a Traces of wear F = 250N, Nc = 40,000 cycles - upper blade

Fig.3.b Traces of wear F = 250N, Nc = 40,000 cycles - lower blade



215




The marks obtained for 200 N and 40000 cycles are given in Figure 6 for the same normal force and for 50,000 to 60,000 cycles traces obtained are given in Figure 7 and 8.


216

Fig.6. Traces of wear F = 200N, Nc = 40,000 cycles - upper blade




217



For 40,000 cycles, the traces obtained are distinguished by small areas of adhesion between the two blades. These adhesion areas increased significantly between 50,000 and 60,000 cycles. At 60,000 cycles areas with much more "red powder" were found on the blade, which proves a proportional increase of the wear during application. For the same length request, we can see an increase in the wasted area, if the normal push force is higher.


218

The marks obtained were taken with a video camera and were subsequently processed on the computer. Examples of fretting wear are presented in fig. 9,10,11.

Fig.9 Traces of fretting for a normal load of 250 N and 50000 load cycles

Fig.10 Traces of fretting for a normal load of 250N and 60000 load cycles

Fig.11. Traces of fretting for a normal load of 200 N and 50000 load cycles


219

CONCLUSION The experimental stall supports experimental tries for the study of fretting. We can

determine the different size of the fretting areas and we can compare them with the theoretical results.

We can identify the presence of "red powder" on the blades subjected to fretting and also we can see a wear increase, proportional with the length of the request. For the same length request, we can see an increase in the wasted area, if the normal push force is higher.

REFERENCES

[1] Ghimisi S., Experimental investigation of the fretting phenomenon-dependence of number cycles, Baltrib’09, V International Scientific Conference, Lithuanian University of Agriculture, Kaunas, Lithuania, 19-21 decembrie 2009, PROCEEDINGS, ISSN 1822-8801, pag.226-230 [2] Johnson K.L.,Contact Mechanics,Cambridge University Press,Cambridge,1985,pp.202- [3] Ghimisi S, Transition in the fretting phenomenon based on the variabile coefficient of fretting, Fiability& Durability, nr 2/2010, pag.89-92, Editura Academica Brancuşi, Târgu Jiu, ISSN 1844ICMERA 2010, 2-4 december Bucharest, Romania, Publisher Institute of Electrical and Electronics Engineers(IEEE), China, ISBN 978-1-4244-8867-4,pag.308-312 [4] Ghimisi S, Study of the transition in the fretting phenomenon, Baltrib’09, V International Scientific Conference, Lithuanian University of Agriculture, Kaunas, Lithuania, 19-21 decembrie 2009, PROCEEDINGS, ISSN 1822-8801, pag.230-236


220

STRESSES DETERMINATION METHOD IN MOVING PARTS OF THE MARINE ENGINES

Assist.Prof.Eng.PhD. MOROIANU CORNELIU,Naval Academy of Constanţa,

ROMANIA,E-mail: [email protected]

Abstract: The moving parts of internal-combustion engines endures the highest and the most complex stresses. The tensile compression, bending and twisting stresses appear under the action of gas pressure forces and inertic forces in the elements of the crankshaft. One to the bending stresses, the crankshaft is the strained compromising the coaxility of necks and bearing bushes. Taking into account these points (considering the points), the determination of the stresses to witch it is put, becomes a defining element for designing and testing to strength. This work presents a numerical modelling (MathCad program) for determining the stresses to which the moving parts is put for the marine two-stroke engines. Key words: numerical modeling, marine internal-combustion engine, moving parts, stresses.

1. INTRODUCTION One to the installing clearances, high speed of the rise of pressure during the combustion and change of application direction of forces, the stress of the crankshaft of the internal combustion engine is like a shock. The variable forces induce the fatigue phenomenon dangerous especially to the passing from the arm to the necks and the torsional vibration stress is also dangerous. Knowing the state values of the motive fluid in the characteristically points of the duty cycle, we can determine the values of the show and effective parameters of the duty cycle as well as the values of the main building dimensions of the engine. We obtain the duty cycle diagram based on the data obtained for the real volumes of the motive fluid in the characteristically points of the duty cycle (Alexandru C,1991).

0 0.5 1 1.50

2

4

6

8

p k( )

yG

V k( ) xG

Fig. 1. The characteristical points of the duty-cicle.


221

2.THE DETERMINATION OF KINEMATIC VALUES OF PISTON AND CONNECTING ROD Having established the type of crank and connecting rod assembly, determined the crank star and chosen the firing sequence we shall develop the kinematic calculation of driving mechanism taking into account the characteristical values of piston and connecting rod (shift, speed, acceleration),(GrunwaldB , 1980).

0 36 72 108 144 180 216 252 288 324 360

1.32 103974.4628.8283.262.4

408

753.6

1.099 1031.445 1031.79 103

2.136 103

xp a vp a 100ap a 10

a a a

Fig 2. The characteristical values for piston resulted from the kinematic calculation. 3. THE DYNAMIC CALCULATION OF DRIVING MECHANISM The dynamic calculation of driving mechanism assumes the determination of gas pressure forces, of inertic forces of masses being in rotating and translation motion as well as of their components. In figure 3, 4 and 5. I presented the diagrams obtained for a marine Sulser RTA84 engine.

0 100 200 300 400

2 103

2 103

4 103

6 103

F k( )

Fp k( )

Fit k( )

k

Fig. 3. The inertic forces of masses in rotating and translation motions.


222

0 100 200 300 400

2

4

6

8

p k( )( )

k Fig. 4. The gas pressure force.

0 100 200 300 400

1 103

1 103

2 103

M k( )

Mm

Mm.mot

k

Fig. 5. Instantaneous momentum and resulted momentum. 4. THE DETERMINATION OF STRESSES OF THE CRANKSHAFT NECKS

The calculation of the direction of resulted force straining the crankpin neck is made according to the positive or negative values of the involved forces and at the values so obtained it was added 180° to achieve the correlation of force variation with the neck surface which takes over those forces. In Figure 6 the polar diagram of stress is presented.

The diagram shores that the highest stresses of crankpin neck are recorded in the scale II for PM=90°… 180°, while the lowest stresses are in the scale IV for PM=270°… 360°.


223

3 103 2 103 1 103 0 1 103 2 1032 103

0

2 103

4 103

6 103

0

Y k( )

y k 1500 ( )

y k 1000 ( )

CircleY 1500 5 ( )

0

G k( ) x k 1500 ( ) x k 1000 ( ) CircleX 1500 5 ( )

Fig. 6. The diagram of forces loading the bearing neck of the crackshaft.

To determine the stresses of crankshaft bearing, it is necessary to specify the planes on which the forces determining these stresses act.

0 100 200 300 400

500

1 103

1.5 103

turning angle [grd RAC]

Rp

[kN

]

Rp.m ed 3

Fig. 7. The polar diagram of forces loading the base bearing

The total loads of forces acting on a crankshaft bearing are determined taking into account the arrangement of cylinders, the position of contiguous cranks round about the axis of rotation and the firing sequence (Dragalina A, 1993).


224

The determination is made based on the moment equations of pressure forces and inertia forces of masses with rotating and translation motions being also included the forces generated by counterweights. In Fig. 7 and 8 there are the diagrams of forces loading the bearing neck as well as the polar diagram of force loading the base bearing(Bobescu G, 1997). .

2 103 1 103 0 1 103 2 103

2 103

1 103

1 103

2 103

Fig. 8. The polar diagram of forces loading the base bearing. 5. CONCLUSION The paper presents a type of numerical modeling represented by a program written in MathCad program for determining the stressed at which the moving parts neck of marine engines are loaded, but not only. Such a program is useful for designing the engines as well as for testing the strength of the crankshafts. REFERENCES [1]Alexandru C.;Marinepropulsionmachineryand equipment, Technical Publishing

House,Bucharest, 1991,p.57. [2]Grunwald B.; Theory, calculation and construction of engines formotor vehicles,

Technical Publishing House,Bucharest, 1980,p. 124 . [3]Dragalina Al.; Applicationscalculationofmarine diesel engines,Publisher of Naval

Academy"Mircea cel Bătrân", Constanța,1993, p.20-65. [4]Bobescu G. s.a., Enginesfor automobiles -designguide,University"Transilvania", Braşov,

1997,p.65. [5]Moroianu C.s.a., Marine Engines,Technical Publishing House,Bucharest, 1996, p.305.


220

STRESSES DETERMINATION METHOD IN MOVING PARTS OF THE MARINE ENGINES

Assist.Prof.Eng.PhD. MOROIANU CORNELIU,Naval Academy of Constanţa,

ROMANIA,E-mail: [email protected]

Abstract: The moving parts of internal-combustion engines endures the highest and the most complex stresses. The tensile compression, bending and twisting stresses appear under the action of gas pressure forces and inertic forces in the elements of the crankshaft. One to the bending stresses, the crankshaft is the strained compromising the coaxility of necks and bearing bushes. Taking into account these points (considering the points), the determination of the stresses to witch it is put, becomes a defining element for designing and testing to strength. This work presents a numerical modelling (MathCad program) for determining the stresses to which the moving parts is put for the marine two-stroke engines. Key words: numerical modeling, marine internal-combustion engine, moving parts, stresses.

1. INTRODUCTION One to the installing clearances, high speed of the rise of pressure during the combustion and change of application direction of forces, the stress of the crankshaft of the internal combustion engine is like a shock. The variable forces induce the fatigue phenomenon dangerous especially to the passing from the arm to the necks and the torsional vibration stress is also dangerous. Knowing the state values of the motive fluid in the characteristically points of the duty cycle, we can determine the values of the show and effective parameters of the duty cycle as well as the values of the main building dimensions of the engine. We obtain the duty cycle diagram based on the data obtained for the real volumes of the motive fluid in the characteristically points of the duty cycle (Alexandru C,1991).

0 0.5 1 1.50

2

4

6

8

p k( )

yG

V k( ) xG

Fig. 1. The characteristical points of the duty-cicle.


221

2.THE DETERMINATION OF KINEMATIC VALUES OF PISTON AND CONNECTING ROD Having established the type of crank and connecting rod assembly, determined the crank star and chosen the firing sequence we shall develop the kinematic calculation of driving mechanism taking into account the characteristical values of piston and connecting rod (shift, speed, acceleration),(GrunwaldB , 1980).

0 36 72 108 144 180 216 252 288 324 360

1.32 103974.4628.8283.262.4

408

753.6

1.099 1031.445 1031.79 103

2.136 103

xp a vp a 100ap a 10

a a a

Fig 2. The characteristical values for piston resulted from the kinematic calculation. 3. THE DYNAMIC CALCULATION OF DRIVING MECHANISM The dynamic calculation of driving mechanism assumes the determination of gas pressure forces, of inertic forces of masses being in rotating and translation motion as well as of their components. In figure 3, 4 and 5. I presented the diagrams obtained for a marine Sulser RTA84 engine.

0 100 200 300 400

2 103

2 103

4 103

6 103

F k( )

Fp k( )

Fit k( )

k

Fig. 3. The inertic forces of masses in rotating and translation motions.


222

0 100 200 300 400

2

4

6

8

p k( )( )

k Fig. 4. The gas pressure force.

0 100 200 300 400

1 103

1 103

2 103

M k( )

Mm

Mm.mot

k

Fig. 5. Instantaneous momentum and resulted momentum. 4. THE DETERMINATION OF STRESSES OF THE CRANKSHAFT NECKS

The calculation of the direction of resulted force straining the crankpin neck is made according to the positive or negative values of the involved forces and at the values so obtained it was added 180° to achieve the correlation of force variation with the neck surface which takes over those forces. In Figure 6 the polar diagram of stress is presented.

The diagram shores that the highest stresses of crankpin neck are recorded in the scale II for PM=90°… 180°, while the lowest stresses are in the scale IV for PM=270°… 360°.


223

3 103 2 103 1 103 0 1 103 2 1032 103

0

2 103

4 103

6 103

0

Y k( )

y k 1500 ( )

y k 1000 ( )

CircleY 1500 5 ( )

0

G k( ) x k 1500 ( ) x k 1000 ( ) CircleX 1500 5 ( )

Fig. 6. The diagram of forces loading the bearing neck of the crackshaft.

To determine the stresses of crankshaft bearing, it is necessary to specify the planes on which the forces determining these stresses act.

0 100 200 300 400

500

1 103

1.5 103

turning angle [grd RAC]

Rp

[kN

]

Rp.m ed 3

Fig. 7. The polar diagram of forces loading the base bearing

The total loads of forces acting on a crankshaft bearing are determined taking into account the arrangement of cylinders, the position of contiguous cranks round about the axis of rotation and the firing sequence (Dragalina A, 1993).


224

The determination is made based on the moment equations of pressure forces and inertia forces of masses with rotating and translation motions being also included the forces generated by counterweights. In Fig. 7 and 8 there are the diagrams of forces loading the bearing neck as well as the polar diagram of force loading the base bearing(Bobescu G, 1997). .

2 103 1 103 0 1 103 2 103

2 103

1 103

1 103

2 103

Fig. 8. The polar diagram of forces loading the base bearing. 5. CONCLUSION The paper presents a type of numerical modeling represented by a program written in MathCad program for determining the stressed at which the moving parts neck of marine engines are loaded, but not only. Such a program is useful for designing the engines as well as for testing the strength of the crankshafts. REFERENCES [1]Alexandru C.;Marinepropulsionmachineryand equipment, Technical Publishing

House,Bucharest, 1991,p.57. [2]Grunwald B.; Theory, calculation and construction of engines formotor vehicles,

Technical Publishing House,Bucharest, 1980,p. 124 . [3]Dragalina Al.; Applicationscalculationofmarine diesel engines,Publisher of Naval

Academy"Mircea cel Bătrân", Constanța,1993, p.20-65. [4]Bobescu G. s.a., Enginesfor automobiles -designguide,University"Transilvania", Braşov,

1997,p.65. [5]Moroianu C.s.a., Marine Engines,Technical Publishing House,Bucharest, 1996, p.305.


225

THE MARINE HEAVY FUEL IGNITION AND COMBUSTION BY PLASMA

Assist. Prof. Eng. PhD. MOROIANU CORNELIU, Naval Academy of Constanţa,

ROMANIA, E-mail: [email protected]

Abstract: The continuous damage of the used fuel quality, of its dispersion due to the increasing viscosity, make necessary the volume expansion and the rise of the e electric spark power used at ignition. A similar situation appears to the transition of the generator operation from the marine Diesel heavy fuel to the residues of water-fuel mixture. So, it feels like using an ignition system with high specific energy and power able to perform the starting and burning of the fuels mentioned above. Such a system is that which uses a low temperature plasma jet. Its use involves obtaining a high temperature area round about the jet, with a high discharge power, extending the possibility of obtaining a constant burning of different concentration (density) mixtures. Besides the action of the temperature of the air-fuel mixture, the plasma jet raises the rate of oxidation reaction as a result of appearance of lot number of active centers such as loaded molecules, atoms, ions, free radicals. Key words: fuel, viscosity, burning, plasma jet.

1. INTRODUCTION

The ignition of air-fuel mixtures in the burning points of marine team generators is performed under certain conditions of temperature and pressure, usually from an energy (power) source formed of an electric spark of high voltage 10-15 kV. The reaction taking place at the level of swirling plasma jet lead to the generation of an overbalance concentration of atoms and free radicals (H, CH3, O, OH) and to large quantities of products which couldn’t change (couldn’t take part in the oxidation reaction), such as CO, H2. The use of plasma jet for ignition and burning reduces the ignition delay time, raises the fuel combustion (burning) speed, giving stability to burning. As a feature of combustion process of hydrocarbons in the plasma jet is that their burning assumes the performance of thermal pyrolysis and dissociation (decomposition) as well as the generation of nitric acid and some cyanide components in the combustion products. At the plasma temperature of 3000° K the essential reactions lead to an overbalance of hydrogen atoms as part of the plasmochemical reactions. It can assume that, in case of the plasmochemical reactions, a high swirling burning is generated at the level of plasma jet (plasma-fuel), containing a large quantity of active products, having in the same time a high temperature and a direct influence upon the hydrocarbon oxidation in the air-fuel mixture.

2. THE PHYSICO-CHEMICAL PRINCIPLES OF BURNING START UNDER THE PLASMA JET OF OIL FUELS

Through the heat conducted (transmitted) by the plasma jet to the burning fuel, its combustion speed rises, due to the intensification (enhancement) of heat transfer (flow) from the jet to the air-fuel mixture. It is of interest the general case, namely, the heat action of plasma jet at the burning of one oil fuel drop. The burning/combustion process of a drop includes three distinct stages:



226

- self-ignition delay generated by heat transfer (flow) to the drop till the reacting boiling point (temperature);

- burning of transient (volatile) components in the drop and generation of ecosphere; - the cenosphere burning; provided that the heat transfer from them environment to

the drop is performed by convection and radiation; the drop temperature can be determined by the following formula:

]hhdk

)TT.(N.[

c.d.

6

dt

dT21

kgug

pkkk

k ; (1)

in which:

5,0

eu R.56,02N ; r

kgkk

e

VV.d.R

; (2)

and

2

51 d.

p.I.10.6,4h ; )TT.(a..ah 4

k4

gkg2 ; (3)

where: rk, dk – are the density and the diameter of the drop; cpk – thermal conductivity of fuel; lt – thermo conductibility coefficient of the environment; mt – viscosity; I – heat intensity of plasma jet; p – pressure; s – Stefan Bolhzman constant; ag – air emission factor from the determination stand; ak – the blackening degree of drop surface ak=0.05-0.30 for heavy fuels.

The variation of drop diameter as a result of the evaporation and burning processes of the volatiles can be expressed by:

]hhd

)TT(N..[

N.

2

dt

dd21

k

kgug

uk

k ; (4)

in which:

1uu H.NN0

; B

)B1ln(H1

; (5)

Nkgpg H/)TT.(cB ; (6)

where:


227

cpg – is the thermal conductivity of fuel vapors at p=ct; HN – heat necessary for fuel vaporization.

The temperature of the drop can be determined by:

])d

d(1[KTT 3

k

kpfk

0

; (7)

where: Tf – boiling temperature; kp – factor of proportionality; dko – the initial diameter of drop;

Supposing that the burning of coke formed is developed on the entire surface of the drop and from the point of view of physical phenomenon it is an oxidation of carbon, we can write:

2oxcox

cox C.k.75.o

dt

dd

; (8)

where: kx – is the constant of chemical reaction rate; dcox, rcox – the diameter and the density of coke particles; Co2 – represents the oxygen amount necessary for burning a particle of coke.

goo T.R

p.21,0C

2

2 ; (9)

where: Ro2 – excess of burning air.

The motion equations to the axial and radial direction for the fuel drop with spherical symmetry passing through all the burning stages are given by:

kdzz2kk

gvz Re.c).VV.(d.8

6

dt

dkg

k ); (10)

kdkk

2k

rrk Re.c).VrVr.(d.8

..6

dt

dv ; (11)

-in witch the resistance (consistency/strength) coefficient of the drop is given by:

cd = 27.Re-0,84 - for Rek 80; (12)

cd = 0,271.Reo,271 - for 80Rek104; (13)


228

cd =2, - for Re 104; (14)

-the viscosity coefficient of gas is determined by:

3

1

pa2

pa

33

2

pa3g

)zV.G(

l.).d.(012,0 ; (15)

where: l3 – is the length of the anode plate of the experimental system.

rpa, Gpa – are the density and the plasma consumption of the experimental system.

The set of the differential equations has been resolved by RUNG-KUTT method determining the axial and radial coordinates such as:

t

0

kkk dt.VzZZ0

; (16)

.dt.Vrrrt

0

kkx 0 (17)

3. CONCLUSION

The tests (experiments) have been performed for three types (classes) of fuel: fuel oil, marine Diesel heavy and light fuels. There are presented the variations of drop diameter and Reynolds number for loads of plasma generator Gp = 0,1; 0,2; 0,5 ; 1,0 ; 2,0 . 103 kg/s. On the increase of burning air consumption, the burning time of fuel drops rises because of the reductive of plasma temperature and also the heat received by fuel drop. The dependence of complete combustion of fuel drops is given by the intensity of voltaic arc for generating of plasma. The time of perfect combustion tn is suddenly reduced on the increase of the intensity of voltaic arc up to the value of 10 A, then its reduction is not important when the intensity of voltaic arc increases up to 24 A. An essential factor of total time of combustion is the initial diameter of the drops (the quality of atomization). When the drop diameter increases from 50 to 300 µm, the combustion time rises 6-7 times. REFERENCES [1] Chmela F.G., Kapus P.E., Advanced Engine Technology for Low Emission, Proceedings of 1995 Seoul Motor Show Seminar, 1995, p. 1-14; [2] Kyaw Z.H., Watson H.C., Hydrogen Assisted Jet Ignition for Near Elimination of NOx and Cyclic Variability in the SI Engine, 24th International Symposium on Combustion, the Combustion Institute, 1992, p.1449-1455; [3] Bromberg L, Cohn D., Rabinovich A, Heywood J, Emissions reductions using hydrogen from plasmatron fuel converters, J. Int. Hydrogen Energy 26, 1998, p.1115-112.

contents · significance. their further development is of great practical importance to improve the...

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