Hidraulica magazine 1 2013

ISSN 1453 – 7303 “HIDRAULICA” (No. 1/2013) Magazine of Hydraulics, Pneumatics, Tribology, Ecology, Sensorics, Mechatronics

CONTENTS

• EDITORIAL Petrin DRUMEA

5 - 6

• HYDRAULIC SWITCHING CONTROL – OBJECTIVES, CONCEPTS, CHALLENGES AND POTENTIAL APPLICATIONS

Rudolf SCHEIDL, Helmut KOGLER, Bernd WINKLER

7 - 18

• IMPROVEMENTS IN HYDRAULIC IMPACT MECHANISMS CONTROLLED BY ROTATABLE DISTRIBUTORS

Claudia KOZMA, Liviu VAIDA

19 - 26

• HYDRAULIC AND PNEUMATIC CYLINDER FAILURES , THE EFFECT OF FLUID CLEANLINESS ON COMPONENT LIFE

Patrick Adebisi Olusegun ADEGBUYI, Ioan-Lucian MARCU

27 - 30

• PNEUMATIC PRESSURE SERVOREGULATOR WITH PIEZOELECTRIC ACTUATION

Gabriela MATACHE, Ioana ILIE, Radu RADOI

31 - 41

• MODELING OF A THREE WAY ROTATABLE FLUID DISTRIBUTOR USED TO COMMAND AND CONTROL A HYDRAULIC ROCK DRILL

Claudia KOZMA, Banyai DANIEL VASILE

42 - 51

• THE ANALYSIS OF FLOW LOSSES THROUGH DYNAMIC SEALS OF HYDRAULIC CYLINDERS

Gabriela MATACHE, Stefan ALEXANDRESCU, Adrian PANTIRU, Gheorghe SOVAIALA, Mihai PETRACHE

52 - 60

• DOUBLY FEED INDUCTION GENERATOR FOR BIOMASS COMBINED HEAT AND POWER SYSTEMS

Curac IOAN, Craciun BOGDAN IONUT.Banyai DANIEL VASILE

61 - 64

• A NEW MODEL OF PNEUMATIC TRANSDUCER USED IN THE DRYING STAGE OF THE CERAMIC PRODUCTS OBTAINING

Murad Erol, Dumitrescu Catalin, Haraga Georgeta, Dumitrescu Liliana

65 – 69

• THEORETICAL CONSIDERATIONS REGARDING THE MECHANISM FOR ADJUSTING THE CAPACITY OF THE PUMS WITH RADIAL PISTONS Lepadatu Ioan, Dumitrescu Catalin

70 - 78

• CAVITATION EROSION RESISTANCE FOR A SET OF STAINLESS STEELS HAVING 10 % NICKEL AND VARIABLE CHROMIUM CONCENTRATIONS Ilare BORDEAȘU, Mircea Octavian POPOVICIU

79 - 85

• VIRTUAL INSTRUMENT FOR PLOTTING SERVO-VALVES CHARACTERISTICS AFTER SIGNIFICANT MAINTENANCE OPERATIONS Radu RĂDOI, Iulian DUȚU

86 - 89

• THE INFLUENCE OF ANGLE OF TILT OF THE SEPARATORS AND THE AIR COURSE VELOCITY ABOUT QUALITATIVE COEFFICIENT AND THE EXPLOATATION AT THE CLEANING AND SORTING OF THE CORN PULSES Constantin POPA, Mihaela-Florentina DUȚU, Iulian DUŢU

90 - 96

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MANAGER OF PUBLICATION

- PhD. Eng.Petrin DRUMEA - Manager - Hydraulics and Pneumatics Research Institute in Bucharest,

Romania

CHIEF EDITOR - PhD.Eng. Gabriela MATACHE - Hydraulics and Pneumatics Research Institute in Bucharest, Romania

EXECUTIVE EDITORS

- Ana-Maria POPESCU - Hydraulics and Pneumatics Research Institute in Bucharest, Romania

- Valentin MIROIU - Hydraulics and Pneumatics Research Institute in Bucharest, Romania

SPECIALIZED REVIEWERS - PhD. Eng. Heinrich THEISSEN – Scientific Director of Institute for Fluid Power Drives and Controls IFAS,

Aachen - Germany

- Prof. PhD. Eng. Henryk CHROSTOWSKI – Wroclaw University of Technology, Poland

- Prof. PhD. Eng. Pavel MACH – Czech Technical University in Prague, Czech Republic

- Prof. PhD. Eng.Alexandru MARIN – POLITEHNICA University of Bucharest, Romania

- Assoc.Prof. PhD. Eng. Constantin RANEA – POLITEHNICA University of Bucharest, Romania

- Lecturer PhD.Eng. Andrei DRUMEA – POLITEHNICA University of Bucharest, Romania

- PhD.Eng. Ion PIRNA - General Manager - National Institute Of Research - Development for Machines and

Installations Designed to Agriculture and Food Industry – INMA, Bucharest- Romania

- PhD.Eng. Gabriela MATACHE - Hydraulics & Pneumatics Research Institute in Bucharest, Romania

- Lecturer PhD.Eng. Lucian MARCU - Technical University of Cluj Napoca, ROMANIA

- PhD.Eng.Corneliu CRISTESCU - Hydraulics & Pneumatics Research Institute in Bucharest, Romania

- Prof.PhD.Eng. Dan OPRUTA - Technical University of Cluj Napoca, ROMANIA

Published by: Hydraulics & Pneumatics Research Institute, Bucharest-Romania Address: 14 Cuţitul de Argint, district 4, Bucharest, cod 040557, ROMANIA Phone: +40 21 336 39 90; +40 21 336 39 91 ; Fax:+40 21 337 30 40 ; E-mail: [email protected] Web: www.ihp.ro with support of: National Professional Association of Hydraulics and Pneumatics in Romania - FLUIDAS E-mail: [email protected] Web: www.fluidas.ro HIDRAULICA Magazine is indexed in the international databases:

HIDRAULICA Magazine is indexed in the Romanian Editorial Platform::

HIDRAULICA Magazine: E-mail: [email protected] Web: www.fluidas.ro/hidraulica

4


EDITORIAL About the technical level For a long time people talk about the technical level of products, services, and even of people, but most times it is mixed with technological or quality features. Most of the technical level elements are set by the designers, who predict some technical performances, some work methodologies, based on the level of knowledge in the time of conception, virtually even on their own technical level. It is obvious and has to be pointed out that the technology level of manufacturers determines products price and quality in terms of achieving the expected technical level..

Ph.D.Eng. Petrin DRUMEA MANAGER INOE 2000 – IHP

These things happen also in the fields of hydraulics and pneumatics. Manufacturers and users of hydraulic equipment have concluded that it is important that items related to technical level to be normalized, otherwise there was not possible professional maintenance and an acceptable economic performance. In this way competition between manufacturers is focused on superior performance that can be achieved in terms of predetermined dimensions and areas of basic parameters, such as pressure, also default. Producers who fail to get small weights, small pressure loss, long lifespan, high functional safety, visual aesthetic good appearance and regular functional and maintenance requirements will have big problems on the globalized competitive market. The user will buy cheap equipment, with functional and energy performances appropriate for the integrator equipment, with good reliability and maintenance costs as low as possible. From this brief enumeration it is clear that the technical level is implied, it is eliminatory in our choices and it is provided by all serious manufacturers. Technical level of hydraulic equipment creates prerequisites for obtaining high technical performance for complex equipment, and in most cases this, combined with high quality of products, leads to increased technical and economic performance of the integrated assembly. Issues related to raising the technical level of hydraulic equipment are not simple, even if sometimes they seem so, but there always has to be made a technical and economic choice having to be noticed what an increase in performances such as pressure or frequency inside the machine as a whole involves, considering that the entire structure must be changed and possibly the maintenance and safety of the entire system will suffer. There should be included in the technical level also the news inside the field such as hydraulics digitization, use of new clean working fluids, introducing electronic and mechatronic parts in the structure of equipment and systems and even the use of some new materials. In the end we can say that the technical level should not be considered as a single element of the production but only within certain general considerations, in close connection with technological processes and quality assurance methods.

5


EDITORIAL Despre nivelul tehnic De foarte mult timp oamenii discuta de nivelul tehnic al produselor, al serviciilor, ba chiar si al persoanelor, insa de cele mai multe ori acesta este amestecat cu elemente tehnologice sau de calitate. Cele mai multe elemente ale nivelului tehnic sunt stabilite de catre proiectanti, care preconizeaza anumite performante tehnice, anumite metodologii de lucru, pornind de la nivelul cunoasterii in perioada de conceptie, practic chiar de la nivelul lor tehnic. Este evident si de subliniat ca nivelul tehnologic al fabricantilor determina pretul si calitatea produselor in conditiile realizarii nivelului tehnic asteptat.

Dr.Ing.. Petrin DRUMEA DIRECTOR INOE 2000 – IHP

Aceste lucruri se intampla si in domeniul hidraulicii si pneumaticii. Producatorii si utilizatorii de echipamente hidraulice au ajuns la concluzia ca este bine ca elementele legate de nivelul tehnic sa fie normalizate, pentru ca altfel nu mai era posibila o mentenanta profesionala si o performanta economica acceptabila. In felul acesta competitia intre producatori se concentreaza pe performantele superioare care se pot realiza in conditii de gabarit prestabilite si pe domenii ale unor parametri de baza, precum presiunea, de asemenea prestabiliti. Producatorii care nu reusesc sa obtina greutati mici, pierderi de presiune mici, durata de viata mare, siguranta functionala ridicata, aspect vizual, estetic bun si cerinte functionale si de intretinere normale, vor avea mari probleme pe piata concurentiala globalizata. Utilizatorul va cumpara echipamentele ieftine, cu performante functionale si energetice adecvate utilajului integrator, cu o buna fiabilitate si cu costuri de mentenanta cat mai mici. Din aceasta scurta enumerare reiese clar ca nivelul tehnic este subinteles, el este eliminatoriu in alegerile noastre si este asigurat de catre toti producatorii seriosi. Nivelul tehnic al echipamentelor hidraulice creaza premizele obtinerii unor utilaje complexe de mare performanta tehnica, iar in cele mai multe cazuri acesta combinat cu o calitate buna a produselor conduc la cresterea performantelor tehnico-economice ale complexului integrator. Problemele ridicarii nivelului tehnic al echipamentelor hidraulice nu sunt simple, chiar daca uneori par asa, insa intotdeauna trebuie facuta o alegere tehnico-economica in care trebuie vazut ce implica o crestere a unor performante de tipul presiunii sau frecventei in interiorul masinii in intregul ei, tinand cont ca trebuie modificata intreaga structura si ca s-ar putea ca mentenanta si siguranta intregului sistem sa aiba de suferit. De asemenea trebuie incluse in nivelul tehnic si noutatile din domeniu precum digitalizarea hidraulicii, utilizarea unor fluide de lucru noi, nepoluante, electronizarea si mecatronizarea echipamentelor si sistemelor si chiar si utilizarea unor materiale noi. In final se poate spune ca nivelul tehnic nu trebuie considerat ca un element singular al productiei ci doarin interiorul unor aprecieri generale, in legatura stransa cu procesele tehnologice si cu metodele de asigurare a calitatii.

6


HYDRAULIC SWITCHING CONTROL – OBJECTIVES, CONCEPTS, CHALLENGES AND POTENTIAL APPLICATIONS

Rudolf SCHEIDL1, Helmut KOGLER1, Bernd WINKLER2 1 Johannes Kepler University Linz, [email protected]. 2 Linz Center of Mechatronics, [email protected]

Abstract: Hydraulic switching control operates via the switching of valves. Numerous principles exist, some of which are already routinely applied, some others have been studied, but many more are possible. A successful realization of many switching techniques requires fairly advanced hydraulic components, foremost fast switching valves, fast check valves and compact accumulators which can resist high load cycles, and a sound understanding of the relevant processes by advanced modelling, simulation, and experimental analysis. Switching control can bring the following advantages: lower costs, higher robustness, better standardization, easier control due to better repeatability and less hysteresis; better energetic efficiency by the application of energy saving converter principles; generation of very fast motion for relatively high loads. Some elementary switching principles and switching converter principles are described in this paper. The authors expect an early industrial application of novel switching techniques in heavy duty and in agricultural machines. Keywords: Hydraulic switching control, switching converters, buck converter, electric hydraulic analogy.

1 Introduction

The idea to employ switching systems for realizing some sort of hydraulic transformers is an old idea as the hydraulic ram of Montgolfier from 1796 and Pollard’s work on pulsating hydraulic power transmission (1964) [1] show. The major motivation for the current attempts to realize hydraulic switching control is the success of switching control in modern electric drive technology.

Hydraulic switching control is a sub-domain of digital hydraulics, which is characterized by performing control only by the use of components with discrete states (see [2]). In hydraulic switching control these discrete state components are switching valves. Control input parameters are some timing characteristics of these valves which can be: the pulse-width of a pulse width modulated valve switching, the duration of individual pulses, switching frequencies, or the phase shifts of switching pulses of different valves, to mention just a few.

The basic motivations for switching control are:

• To use the simple component switching valve instead of the more complex servo or proportional valves, to achieve one or several of the following goals: lower costs, higher robustness, better standardization

• Easier control: better repeatability, less hysteresis • Better energetic efficiency by the application of energy saving converter principles • Generation of very fast motion for relatively high loads

Like in power electronics, the variety of possible switching concepts is very large and there is not just one optimal solution for many applications but rather a few optimal for a small class of applications. Some concepts evaluated so far are:

7


• Non energy saving concepts: Elementary switching control concepts [3], the LCM ‘ Digi-

Actuator’ [4], combinations of pulse code and switching control concepts [5]. • Energy saving switching control principles: the Gall&Senn principle [6], buck- [7], resonance-

[8], wave- [9], motor-converters [10, 11]. Some hydraulic switching principles are routinely applied since decades in the following applications: ABS break systems, variable valve timing for compressor valves [12], clutch actuation of gears in passenger cars [13], on-line gap adjustment of continuous casting segments [14, 15]. All these drivers require relatively low power and flow rate, respectively. Fast switching valves for such power ratings can be realized much easier than for higher flow rates and also pulsation and noise excitation levels are relatively low. The application of switching control to high power drives is facing the following main challenges:

• Appropriately powerful fast switching and check valves, and fast, compact, and reliable accumulators.

• The sound understanding and the mastering of pulsating phenomena and of acoustic noise. • Control algorithms for switching control, which not only provide a proper control of the

intended motions, force, or pressure but which also cope with the challenges of switching control, in particular with pulsation.

Switching control is also a useful technology for other than mechanical actuation. Not only modern diesel fuel injection, in particular common rail systems, apply switching principles but also the cooling of rollers employs this technology to realize the so called thermal crown in steel and aluminium rolling [16, 17]. There is no limitation of the application of such switching control concepts in principle. Currently, the only reasonable question is which applications are best suited for an early application of these technologies. This is not just a technical matter but depends even more on the disposition of end users to take the front runner role. To convince potential end users a single advantage, like just being more efficient, may be a too weak argument, there must rather be several advantages. In the next chapter some switching control concepts are presented. In chapter 3 areas of early applications of hydraulic switching control are discussed. An outlook on the further development of switching control will be given in chapter 4.

2 Switching control concepts

2.1 Electric – hydraulic analogy and duality Hydraulic and electric machines are power converters with one mechanical port, a shaft or a rod. Electrical machines employ a dynamical principle for torque or force generation, respectively. This corresponds to hydrodynamic machines, like turbines or centrifugal pumps. Hydrostatic machines, however, employ a static principle which is dual to the dynamic principles. In the dynamic principles machines, torque or force depend on the flow variable, i.e. on the electric current or on the flow rate, respectively; in hydrostatic principles machines torque or force depend on the pressure. We limit this investigation to hydrostatic machines. The duality of static and dynamic machines has significant consequences for switching control. A very elementary circuit for switching control of an electrical motor is shown in Figure 1a. The current i of the electric motor which is driven by a PWM voltage signal has a small fluctuation due to the flattening effect of the motor inductance. The angular speed’s (ω) fluctuation is even smaller, because the inertia of the rotor and of attached moving mechanical parts have a flattening effect too. Both flattening effects are very expressed due to the high switching frequencies of modern drives, which are in the order of 104 Hz. That leads to a quite constant speed.

8


UE L

R M, Jω

ii~M

tTper

∆itTper

∆ω

a)

t

Tsw

∆U t

Tsw

∆F tTsw

∆v

F, v

pS

mb)

Figure 1: Elementary switching control circuits of a) an electrical motor, b) a hydraulic actuator

Hydraulic elementary switching control, as shown in Figure 1b, intrinsically generates higher speed fluctuations because the acceleration corresponds to the hydraulic force F which is approximately a rectangular signal. This requires extra devices to get rid of the strong pressure pulsation due to switching, particularly if the mechanical inertia of the system to be actuated is relatively small. A further main difference between electrical and hydraulic switching systems is the high capacity of hydraulic systems. In Figure 1b, for instance, that capacity stems from oil compressibility in the chambers of the cylinder and in the connecting lines. One hardly can get rid of this capacity, which causes energetic losses in combination with the resistance of the valve at higher switching frequencies. There is some resonance effect resulting in an energy optimal switching frequency (see [18]). In electrical switching systems parasitic capacities are a negligible problem unless switching frequencies are extremely high. On the components level a strong analogy exists between electrical and hydraulic switching systems. This concerns mainly the key component valve, the performance of which is a limiting factor for advancements in switching control. The availability of fast electric ‘valves’ is a major reason why electrical engineering leads the way. The progress there, for instance with respect to switching frequency or power range, was guided by advancements in power semiconductors technology.

2.2 Elementary switching control principles In elementary switching control just one or more valves are put in front of a hydraulic cylinder or motor, respectively. There are no other components, like pulsation attenuation devices or transmission lines, which exhibit a significant dynamics at the actual operating condition, between valves and the actuator. The simplest circuit is shown in Figure 2, comprising just two 2-2 way valves. Switching frequency is a major operation parameter for the performance of such a system since high frequencies flatten pressure and velocity pulsation significantly.

uP

t

t

uT

t

uT one way mode

two way mode

Fload

VP

VT

pS

pT

p

s

uP

tTsw

κ Tsw

max(s).

min(s).

1

pmax(p)

min(p)s.

Figure 2: Elementary switching control schematic (left) and typical signals in one way mode of operation (right)

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One way mode In the so called one way mode of operation, only one switching valve is actuated. To lift the load Fload, valve VP is actuated, e.g. employing PWM control with a frequency ω /2π = 1/ TSW and a duty cycle κ, to lower the load VT is activated. The one way mode switching cannot provide better efficiency than resistance control with continuously operating valves. If there is no flow from the tank line to the cylinder in a load lifting operation, continuous or switching control give the same energetic efficiency η.

SS

load

pp

ApF

==η (1)

The fluctuation in chamber pressure p and in piston velocity s depends on several parameters. In [19] the following approximate value for the nondimensional pressure fluctuation ψfl = (max(p)-min(p))/pS is derived under the assumptions of constant supply pressures pS, pT, a very fast switching, and that the nonlinear state equation can be linearized around the mean values of pressure and speed and at position x0 = ξ0 smax.

( ) ( )επ

ξκκψ

ω

ωω 21

0 cafcg

bload

fl−

+−

= (2)

The meaning of the nondimensional quantities in this result is given by the following set of relations

ωω

ωωωω

ωω

ω

ψκψ

ψψψ

ωω

ωε

τξωτξψ

cf

cafcgvaa

agsmFf

pp

sAQa

smpc

sAVb

Ep

ddvt

sspp

loadm

loadm

m

m

mm

loadload

N

SN

SSS

=−

=−

=

−+−===

Α=======

;;12

11;;

;;;;/;;

2maxmax

2maxmax

0

max

(3)

The fluctuation of the nondimensional speed v around its mean value vm is of order O(1/ω 2). ψm is

the nondimensional pressure mean value. pS is the system pressure, s the piston position, smax the cylinder stroke, t is the physical, τ the nondimensional time, V0 is the dead volume of the cylinder chamber at s = 0, m the moved mass, A the piston area, QN the nominal flow rates of both valves (at nominal pressure loss pN). An analysis of Eqs. (2), (3) shows that pressure fluctuation can be reduced by a high switching frequency ω , by a valve with a low nominal flow rate QN, and by a high dead volume V0. QN, however, is determined by the required speed of the system, higher V0 limits the stiffness of the system which may deteriorate the control performance of a closed loop drive. Therefore, the only independent parameter to improve fluctuation is switching frequency. Large drives which are running relatively low maximum speed require smaller ω to stay below a certain pressure fluctuation level than high speed, short stroke drives. Two way mode In this mode pressure and tank valves are switched alternately. Under certain conditions this may yield better or worse efficiency than the one way mode of operation. A better efficiency is obtained if the oscillation of the load m are such that in phases when VT is switched on oil is sucked from the tank line, a worse efficiency in opposite case.

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a) 0 2 4 6 8 10 12 14 16 18 20

-0.2

0

0.2

0.4

0.6

0.8

1

1.2

τ

ψη

v

b) τ0 2 4 6 8 10 12-0.02

0

0.02

0.04

0.06

0.08

0.1

0.12

v

η

c) τ

v

η

0 10 20 30 40 50 60 70 80 90 100-0.1

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

Figure 3: 1 Results of a numeric computation of an elementary switching system in dual mode; switching

frequency is: a) identical to natural frequency of the system; b) twenty times the ntaural frequency ; c) wobbled between one tenth and five times the natural frequency; values of the other parameters: aω = 1; cω = 1; fload = 0.5; κ

= 0.5; (b+ξ0) = 1.

Figure 3 shows results of numerical simulations of an elementary switching system according to Figure 2 in two way mode at different switching frequencies. High efficiency is obtained if this frequency is below and up to the natural frequency of the hydraulic drive’s free oscillation, determined by the load inertia m and the compliance of the hydraulic cylinder. Going much beyond that frequency reduces fluctuation but efficiency drops and may even fall below that of resistance control because part of the flow from the pressure line is transferred to the tank line. This can be avoided with check valves as has been proposed in [6]. The major drawback of this type of switching control is that a good efficiency is bound to high pressure fluctuation, because only if pressure p reaches nearly system and tank levels in each switching cycle, a good efficiency can be obtained. This requires the frequency not to go much beyond the natural frequency of the system and, therefore, also the speed fluctuation will be high in many cases. The considerations presented so far were based on the assumptions that • the supply lines have constant pressure and are not disturbed by the strongly pulsating flows

going to or coming from the switching system • the line between the valves and the actuator are ideal, thus do not exhibit significant

impedances. Particularly for very high frequencies this requires adequate components, e.g. fast response accumulators, and special care in the design of the hydraulic system.

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2.3 Switching converters Converters use some intermediate system between the switching valves and the actuator to resolve the trade-off problem of elementary switching control mentioned before. Figure 4 shows schematics of four switching converters. Wave Converter This converter type was investigated in [9, 20]. By resonant switching standing waves are generated in a properly designed pipe network. This network acts as a filter that annihilates lower order pulsation (several low order Fourier Spectrum components) such that at the exit port only very small pressure or flow rate oscillations occur. The output pressure is proportional to the pulse width κ. The resonance condition requires the length L of the first order twin pipe to have half wavelength. For a switching frequency of 100 Hz L is approx. 6 m. This length condition is the main drawback for a practical implementation unless switching frequencies > 500 Hz are possible. Resonance Converter It employs an oscillating mass in form of some piston which is supported by a spring. By switching the cylinder chamber alternately to pressure-, tank-, and exit-line the piston oscillates and hydraulic fluid is transmitted to the consumer. The system operates close to the resonance frequency of the spring mass system utilizing the frequency response characteristics close to the resonance point for flow rate control. More information and results are given in [8, 21]. This system can be made fairly compact for switching frequencies beyond 100 Hz. It is a step-down and step-up converter, thus even output pressures exceeding the system pressure can be realized. The optimal timing of the consecutive switching of the valves is critical for efficiency. Buck Converter This is a very simple system and corresponds directly to most electrical switching power supply devices. More can be found in [7, 22]. It is a step-down converter but can also recuperate energy from the hydraulic system. The inductance pipe’s length depends on the switching frequency, the higher the shorter this pipe. Motor Converter It follows same principles as the buck converter, only the inductance element is not the fluid inertia of a pipe but the rotary inertia of a pump-motor unit (see, e.g. [7, 8]). The advantage is that this inertia can be controlled independently from the capacitance and resistance of the system. The disadvantages of this concept are the costs and weight of the pump-motor unit and the hydraulic capacitance in such units which is a source for losses at higher switching frequencies.

Wave Converter

Resonance Converter

Buck (Step-down) Converter Motor Converter

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pressure

line

tank line

VP

VT

VCHKP

VCHKT

inductance pipe

Acc

exit port

pressure line

tank line

VP

VT

VCHKP

VCHKT

inductance motor

Acc

exit port

Figure 4: Four types of switching converters

2.4 Exemplary results of the hydraulic buck converter (HBC) An extensive study of the HBC, its principles, modelling, components design, and control, can be found in the thesis [22], some results concerning its application to mobile robots in [23, 24, 25], and the application of several HBCs in parallel to avoid the accumulator (Acc in Figure 4) in [26]. Figure 5 presents some steady state performance results of a low power HBC prototype. This converter can also recuperate energy which yields a substantial energy improvement over a resistance control concept.

Figure 5: HBC prototype and its steady state efficiency results; (system data: inductance pipe: length: 1.2 m, diameter 3 mm, oil: Shell Tellus 15 cSt @ 40°C; switching valves: LCM FSVi (QN=10 l/lin @ 5 bar); switching

frequency: 100 Hz; accumulators: piston accumulators, developed by the authors; find all data in [22] and [24]).

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Figure 6: HBC (as shown in Figure 5) in motion control, lifting an lowering a load (Cylinder: 63 mm x 45 mm piston/rod size, stroke xmax = 600 mm); applied control concept: flatness based controller (details in [22]).

Assessment of the HBC: Pros: This is a simple system which can reduce energy consumption considerably; energy recuperation is possible. Cons: The size of the inductance pipe may be a problem, unless higher switching frequencies are possible; winding of the inductance pipe to a coil deteriorates efficiency; the accumulator is a highly loaded system and makes the system soft; this requires some sophisticated control to achieve a good dynamical performance; arrangement of several HBCs in parallel and running them in a phase shifted operation is a means to skip the accumulator and improves dynamical performance; performance can become as good as with fast servo-valves, yet with much lower energy consumption (see [26] for details). Such multiple converter concepts facilitate also a simple standardization, since power rating can be easily adjusted by an appropriate number of converter units. A schematic of such multiple converters is shown in Figure 7.

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Figure 7: Schematic of a multiple (N) HBC system driving one cylinder and computed performance comparison with a servo drive for the case of a ramp like motion (for details see [26]), some data: The cylinder areas are 1.2

and 0.6 m² and the dead volume in the piston chamber is 150 l, which is sufficient for a good pressure attenuation in case of a phase shifted HBC configuration. A typical velocity for positioning the piston is 4 mm/s, which requires a flow rate of about 300 l/min in this case. Such large flow rates cannot be reasonably handled by

a single hydraulic buck converter with state of the art switching valves. Thus, the presented simulations consider 6 HBCs in parallel, which operate at a switching frequency of 50 Hz. The pipe inductance of one

individual HBC is about 2.5 m with a hydraulic diameter of 10 mm.

3 Favourable domains of early applications of switching control

Practical implementation of hydraulic switching control is a challenging undertaking despite its long history as a basic idea. It is a fairly new technology, if it has to meet today’s requirements on hydraulic drives, since it needs also very advanced components and system understanding. Implementation in advanced machinery or plants requires some risk taking. The question is, which branches and areas of applications are most favourable for switching control. The authors see the following domains:

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• Metal production systems: they require high forces, high dynamics and high precision,

the conditions are often very crude, the systems, particularly in steel production, very extended. Servo-hydraulic drives show excellent performance in terms of dynamics and precision; their drawbacks are: high oil cleanliness requirements, which puts high burden on system installation and maintenance, wear of valve metering edges, drift problems concerning valve zero position, and high energetic losses, for instance due to the leakage of zero lap or under lapped servo valves, and high costs. Digital hydraulics and switching hydraulics are a promising alternative, to avoid some of the mentioned deficiencies of servo-drives. This requires switching valves with a fast dynamic response, and appropriate control concepts. Energy saving is not yet a topic of utmost importance. But there is a gradual increase of awareness of customers and, may be, stronger legislative regulations in the one or other field. In steel production, for instance, the political pressure to reduce CO2 consumption and the fact that the core processes, like blast furnaces or steel making by basic oxygen furnaces are already highly optimized processes put some pressure on other energy consuming processes, like the many mechanical drives and actuators, even though their relative share of total energy consumption is very low.

• Agricultural machinery: there is an ongoing trend of mechatronization; this affects also drives and actuators; in [27] the example automatic level control system of a pick-up employing hydraulic actuation is presented; agricultural machines technologies must be low cost, capable of standing the often hard operating conditions, and very service friendly. Switching valves are definitely much more robust than servo or proportional valves and cheaper. An important aspect is also standardization of components leading to a smaller range of product variants which is a major condition for low cost production; this is highly supported by switching control since impedance forming is done by other means but the valve’s nominal flow rate or the spool position. Energy saving is definitely an aspect of high relevance in future; firstly, to save fuel costs, but secondly, to enable higher actuation functionality under the given power limitations of the tractor’s engine or the power transmissions. Currently, there are initiatives (for instance, [28]) to install high voltage power supply systems to overcome power limitations of mechanical and hydraulic transmissions between tractors and implements and to improve efficiency. This should be a strong motivation to think about new concepts for hydraulic drives for such machinery.

• Tool machines: Hydraulic servo drives have been a dominant technology in tool machines till the upcoming of modern speed variable electrical drives. Hydraulic drives have lost ground and are mainly limited to such drives where either very high forces or high compactness are required. There is a trend toward hybrid drives, combining speed variable electrical motors with a hydrostatic transmission (see, e.g., [29, 30]) for high load applications. The hydrostatic transmission provides force amplification but can also provide, for instance, gear shift, load holding, and fast emergency stop. In such symbiotic drives switching, if needed, fast switching, are basic operations. Particularly in the latter case, typical components and hydraulic processes of hydraulic switching control become relevant. In [31] a hydraulic micro positioning drive for tool machine applications is mentioned. Even though this particular case employs servo valve technology for control, switching control is an interesting alterative in terms of cost, compactness, and robustness. In particular compact actuators will gain more importance for increasing the functionality of tool machines under the condition of already very complex systems which leave little room for additional components. For the success of hydraulic actuation in this field new solutions for the complete hydraulic system must be found. The objectives are to get an overall compact and modular actuation system for functions with high forces, little room space, ultimate response, and high robustness. Hyrid combinations with electrical drives are also a direction of novel advantageous drive solutions which would benefit from advanced digital hydraulic components.

16


4 Conclusions

Hydraulic switching control is an alternative to existing analog control techniques with several advantages. In contrast to many other hydraulic innovations it requires a whole bunch of innovations for a successful realization: principles, components, and a new system design approach which considers the coupling between dynamical processes, hydraulic pulsation, mechanical oscillation acoustic noise and the design of the system. If a critical mass of such innovative elements exists and can be obtained from several vendors, a rush of digital hydraulic systems will occur. This and the energy saving options will enable new functionality for several machine systems at fairly low cost and with adequate performance. Also electro-hydraulic hybrid drives and actuators will benefit from advancements of hydraulic switching control, both on the components and on the system level.

Acknowledgement

This work was sponsored by the Austrian Center of Competence in Mechatronics (ACCM) which is a COMET K2 center and is funded by the Austrian Federal Government, the Federal State Upper Austria, and its Scientific Partners.

REFERENCES

[1] F. H. Pollard, “Research Investigation of Hydraulic Pulsation Concepts”. First Quarterly Progress Report (RAC-933-1), Republic Aviation Corporation, Farmingdal, L.I., N.Y., June 1963.

[2] M. Linjama, “Digital Fluid Power – State of the Art.” Proc. of the Twelfth Scandinavian International Conference on Fluid Power, Volume 2(4), SICFP'11, May 18-20, 2011, Tampere, Finland.

[3] R. Scheidl, G. Hametner, “The role of resonance in elementary hydraulic switching control”, Proc. Instn. Mech. Engrs. Vol. 217 Part I: J. Systems and Control Engineering, 2003, pp. 469-480.

[4] A. Plöckinger, R. Scheidl, B. Winker, “Combined PWM- and Hysteresis Switching Control For A Digital Hydraulic Actuator”, The Third Workshop on Digital Fluid Power, DFP’10, October 13-14, 2010, Tampere, Finland.

[5] M. Huova, A. Plöckinger, “Improving resolution of digital hydraulic valve system by utilizing fast switching valves”. The Third Workshop on Digital Fluid Power, DFP10, October, 2010, Tampere, Finland.

[6] H. Gall and K. Senn, „Freilaufventile - Ansteuerungskonzept zur Energieeinsparung bei hydraulischen Linearantrieben“, Ölhydraulik und Pneumatik, 38/1994, Nr. 1-2.

[7] R. Scheidl, B. Manhartsgruber, H. Kogler, B. Winkler, M. Mairhofer, “The Hydraulic Buck Converter - Concept and Experimental Results”. Proc. 6th IFK, 6. International Fluid Power Conference, Dresden, 31.3.-2.4. 2008.

[8] R. Scheidl, G. Riha, “Energy Efficient Switching Control by a Hydraulic ‘Resonance-Converter’”. In C.R. Burrows, K., A. Edge (Eds.): Proc. Workshop on Power Transmission and Motion Control (PTMC’99), Sept. 8-11, Bath, 1999, pp. 267-273.

[9] R. Scheidl, D. Schindler, G. Riha, W. Leitner, “Basics for the Energy-Efficient Control of Hydraulic Drives by Switching Techniques”. In J. Lückel (ed.): Proc. 3rd Conference on Mechatronics and Robotics, Oct. 4-6, Paderborn, Teubner, Stuttgart, 1995.

[10] J. Dantlgraber. “Hydro-Transformer”. European patent application (PCT) Intern. Publication No. WO 00/08339, 2000.

[11] F. Wanga, L. Gua, Y. Chena, “A continuously variable hydraulic pressure converter based on high-speed on–off valves”, Mechatronics, Vol. 21(8), 2011, pp. 1298–1308

[12] D. M.Deffenbaugh et al., “Advanced Reciprocating Compression Technology”. Final Report SwRI, Project No. 18.11052, DOE Award No. DE-FC26-04NT4226, U.S. Department of Energy, December 2005.

[13] K. Murakami, T. Wakahara, I. Fukunaga, H. Sakai, “The Hydraulic Control System for a New Electronic 4WD System”. JSAE Review, Volume 17, Number 4, October 1996 , pp. 447-447.

[14] R. Brandstetter, „Untersuchung von Antriebskonzepten für die automatische Gießdickenverstellung bei Stranggießanlegen mit segmentierter Strangführung", Master thesis, Johannes Kepler University Linz, June 1996.

17


[15] “SMART Segment & DynaGap SoftReduction”,

http://www.industry.siemens.com/datapool/industry/industrysolutions/metals/simetal/en/SMART-Segment-Dyna-Gap-Soft-Reduction-en.pdf

[16] GRIP Engineering AG. Spray Bars for Steel Rolling Mills. Product brochure. 2006. [17] N. Chakraborti, B. Siva Kuamr, V. Satish Babu, S. Moitra, A. Mukhopadhyay, “Optimizing Surface

Profiles during Hot Rolling: A Genetic Algorithms Based Multi-objective Optimization”. Dagstuhl Seminar Proceedings 04461, Practical Approaches to Multi-Objective Optimization, http://drops.dagstuhl.de/opus/volltexte/2005/245.

[18] R. Scheidl, G. Hametner, “The role of resonance in elementary hydraulic switching control”. Proc. Instn. Mech. Engrs. Vol. 217 Part I: J. Systems and Control Engineering, 2003, pp. 469-480.

[19] M. Garstenauer, B. Manhartsgruber, R. Scheidl, “Switching Type Control of Hydraulic Drives - A Promising Perspective for Advanced Actuation in Agricultural Machinery”, New Fluid Power Applications and Components, Society of Automotive Engineers, Warrendale, pp. 37-47, 9-2000.

[20] Schindler D., “Numerische und experimentelle Untersuchungen über pulsierende Strömungen in Rohren und Ventilen als Grundlage für die Entwicklung energieeffizienter Schalttechniken in der Ölhydraulik“, PhD thesis, Johannes Kepler University Linz, 1995.

[21] G. Riha, „Beiträge zur Entwicklung eines energiesparenden hydraulischen Schaltkonverters“, PhD thesis Johannes Kepler University Linz, 1998.

[22] H. Kogler, “The Hydraulic Buck Converter - Conceptual Study and Experiments”, PhD thesis, Johannes Kepler University Linz, 2012.

[23] E. Guglielmino, C. Semini, H. Kogler, R. Scheidl, D.G. Caldwell, “Power Hydraulics - Switched Mode Control of Hydraulic Actuation”, Proc. 2010 IEEE/RSJ International Conference on Intelligent Robots and Systems (IROS 2010), Oct. 18-22, 2010, Taipei, Taiwan.

[24] H. Kogler, R. Scheidl, M. Ehrentraut, E. Guglielmino, C. Semini, D.G. Caldwell, “A Compact Hydraulic Switching Converter for Robotic Applications”, Proc. Bath/ ASME Symposium on Fluid Power and Motion Control - FPMC2010,, Sept. 15-17 2010, Bath, UK, pp. 56-68.

[25] E. Guglielmino, C. Semini, Y. Yang, D.G. Caldwell, R. Scheidl, H. Kogler, “Energy Efficient Fluid Power in Autonomous Legged Robots”, Proc. 2009 ASME Dynamic Systems and Control Conference, Renaissance Hollywood Hotel, October 12-14, 2009, Hollywood, California.

[26] H. Kogler, R. Scheidl, “The hydraulic buck converter exploiting the load capacitance”. Proc. 8th International Fluid Power Conference (8. IFK), March 26 – 28, 2012, Dresden, Germany, Vol. 2(3), pp. 297 – 309.

[27] B. Winkler, R. Scheidl, “Automatic Level Control in Agricultural Machinery”. In Proceedings of the 2nd International FPNI PhD Symposium on Fluid Power, Modena, Italy, 2002.

[28] M. Geißler, Th. Herlitzius, “Mobile Working Machine With An Electrical 4-Wheel Drive”. Seminar MobilTron 2011 - Mannheim, September 28 – 29, 2011.

[29] T. Neubert, „Drehzahlveränderbarer Verstellpumpenantrieb in Kunststoff – Spritzgießmaschinen“, Ölhydraulik und Pneumatik Nr.45, 10/2001.

[30] P. Ladner, K. Ladner, R. Scheidl, H. Strasser, “Investigation of a closed electro-hydraulic hybrid drive”. In Proceedings of the 11th Scandinavian International Conference on Fluid Power, SICFP’09, IJune 2-4, 2009, Linköping, Sweden.

[31] B. Winkler, R. Haas, “A Hydraulic Micro-Positioning System for Industrial Mill Centers. Proceedings of 13th Mechatronics Forum International Conference”, September 17-19, 2012, Linz, Austria, pp. 855 – 862.

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http://drops.dagstuhl.de/opus/volltexte/2005/245


IMPROVEMENTS IN HYDRAULIC IMPACT MECHANISMS CONTROLLED BY ROTATABLE DISTRIBUTORS

Claudia KOZMA1, Liviu VAIDA2 1 Technical University of Cluj Napoca, [email protected] 2 [email protected]

Abstract: Descriptions of two known hydraulic impact mechanism are realized. One of the impact structures is to be patented. Improvements in the constructive structure and in the operating hydraulic scheme of these impact mechanisms are analyzed. The aim is to simplify the hydraulic scheme and the command and control structure, to minimize the hydraulic parasitic capacities and to attain an adjustable impact frequency. In brief, the goals are to improve the constructive structure of a hydraulic impact device or a hydraulic rotary percussive mechanism controlled by a rotary distributor and the operating principle of the mechanism. Keywords: impact mechanism, percussive device, rock drill, hydraulic hammer, control system, control valve, rotary valve, rotatable distributor

1. Constructive structures for hydraulic impact mechanisms using rotary distributors

An older patented impact mechanism, Fig.1, comprises, according to [1], an impact piston, a rotating distributor and a hydraulic turbine. The impact piston delimitates in the impact mechanism housing two working chambers, C1 and C2. The hydraulic rotating distributor is a 3/2-ways control valve and is used to feed with high-pressure fluid flow the working chamber C2 and alternately to discharge it. The mobile element of the used hydraulic distributor is a rotating spool with longitudinal grooves. For the rotation of the rotating spool it is chose a hydraulic turbine. The working or drive chamber C2 is connected alternately through the longitudinal grooves of the

Fig.1. A constructive structure of a hydraulic drilling mechanism using a hydraulic rotary distributor,

according to [1], where: PV1 – hydraulic pressure source (volumetric pump); AH1, AH2 – hydraulic accumulators; T2 –tank; C1,

C2 – work chambers; I – pressure line; II – discharge line.

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mailto:[email protected]



rotating spool with the line I and line II. The working chamber C1 communicates permanently with the hydraulic pressure source PV1. The line I is for supplying with hydraulic fluid under pressure by means of a hydraulic pressure source PV1. A hydraulic accumulator AH1 is also mounted on the line I. The line II is connected to the hydraulic rotary mechanism taken picked for the rotation of the dislocation tool or of the drill steel. The rotary mechanism and the percussive mechanism use two separate hydraulic pressure sources, necessary to realize the rotation of the dislocation tool and the collision of the tool with the target material (rock, concrete, asphalt). A recent impact mechanism [2] presented in [3] and [4] comprises, Fig.2, essentially, an impact piston and a rotary distributor with a rotatable spool. The impact mechanism housing is formed by two bodies, B1 and B2. The impact piston delimitates likewise two working chamber in the impact mechanism housing, C1 and C2. The working chamber C1 is in permanent connection with the hydraulic pressure source PV. The working chamber C2 communicates alternatively by means of the rotatable spool with the hydraulic pressure source PV and with the tank T. The rotatable spool forms with the piston body a 3/2-ways hydraulic rotary distributor. The rotary displacement of the spool is provided by a hydraulic gear motor MRD2-212D.

The impact mechanism and the rotary mechanism can be mounted in series using the same hydraulic pressure source or can function independently with separate hydraulic pressure sources. Both cases are exemplified in the prior art. The mounting in series of the percussion mechanism and the rotary mechanism appears earlier than the 80’s. It is affirmed in [5] that the said mounting in series implies the need of a single hydraulic pressure source thus simplifying the pumping station and the hydraulic scheme. Contrary, as the rotary hydraulic mechanism is supplied with fluid flow from a separate pump and through a separate circuit, as in [6], the working of the rotary mechanism is realized independently from the percussive mechanism but the resulted hydraulic scheme is more complex with the other corresponding consequences. The operation cycle of the impact mechanisms from Fig.1 and Fig.2 is controlled by feeding with working liquid along the line I, respectively from the pump PV to the hydraulic linear motors. It is to be underlined that for the second impact mechanism, Fig.2, it is needed additionally an external separate element to rotate the rotatable spool in contrast with the first impact mechanism. The hydraulic turbine, Fig.1, which rotates the rotatable spool, is incorporated in the impact mechanism housing and is operated hydraulically by the fluid spent in the working chamber C2. The rotatable

Fig.2. A constructive structure of a hydraulic drilling mechanism using a hydraulic rotary distributor,

according to [2], where: PV – hydraulic pressure source (volumetric pump); B1, B2 – bodies or housings of the impact mechanism;

T1 – tank; C1, C2 – work chambers; MRD2 – hydraulic gear motor.

20


spool, Fig.2, is rotated by means of a rotary mechanism actuated in this case hydraulically, MRD2-212D, and supplied with hydraulic fluid from a separate hydraulic pressure source (not illustrated).

2. Operating principles for hydraulic impact mechanisms using rotary distributors

The operation principle of the impact mechanism with the construction scheme in Fig.1 is illustrated in Fig.3. The liquid flow sent from the hydraulic pressure source PV1 to the hydraulic linear motor MHL is divided into two flows. One liquid flow is conveyed to the working chamber C1 through an existing permanent communication. The rest of the liquid flow is sent through the rotary distributor DHR to the working chamber C2. The working liquid spent to obtain a certain piston stroke is farther sent to the return line the destination being the tank T3. Not all the working fluid arrives in the tank T3 due to the non-return valve NRV and the throttling valve. The remained working fluid is reflected. At this time the working position of the rotary distributor is switched and the connection with the working chamber C2 through the rotary distributor is blocked. In consequence, the flow will charge the hydraulic accumulator AH2 and will start the hydraulic turbine. The hydraulic turbine communicates according to [1] through two tangential holes with a discharge line connected to the tank T2.

As long as the hydraulic pressure source PV1 is supplying with working fluid the impact mechanism, the hydraulic turbine HT operates on the base of the liquid sent to the working chamber C2. Moreover, the hydraulic accumulator AH2 tends to discharge and feed the hydraulic turbine. In consequence the hydraulic turbine continues to rotate the fluid distributor at substantially the same speed. The rotatable spool of the distributor DHR is rotated at the maximum rate achieving the highest frequency of impacts delivered by the impact piston to the tool. The main part of the fluid withdrawn by the rotary mechanism MHR through the 4/3-ways linear hydraulic distributor DHL is used to rotate the fluid distributor when the supply with high-pressure fluid flow from the hydraulic pressure source PV1 to the hydraulic percussive mechanism is cut off. The said main part of the liquid is conveyed by mean of the flow governor through line II and to the hydraulic turbine. The rest of liquid passes through the flow restrictor FR to the tank T3. The achieved frequency is lower in this case and increases back when feeding with liquid to the hydraulic percussive mechanism is resumed. Another way to redistribute the liquid spent by the hydraulic rotary mechanism MHR to rotate the drill tool is revealed through [7]. In [7], the flow governor misses and the fluid flow discharged from the rotary mechanism is directed to the impact piston serving in increasing the impact piston velocity. Thereafter, as the rotation of the hydraulic turbine HT is started, the hydraulic turbine receives an extra fluid flow only when the hydraulic percussive mechanism is supplied with high pressure fluid

Fig.3. The operating principle of the hydraulic impact mechanism

presented in [1], where: T1, T2, T3, T4 – tanks; PV1, PV2 – hydraulic pressure sources; AH1, AH2 – hydraulic accumulators; C1, C2 – work chambers;

DHR – hydraulic rotary distributor; DHL – hydraulic linear distributor; MHL – hydraulic linear motor; MHR – hydraulic rotary

motor; HT – hydraulic turbine; RG – reduction gear; FG – flow governor; NRV – non-return valve; TRV – throttling valve; FR –

flow restrictor.

21


flow from PV1. The increase in fluid flow in the line II leads in closing of the non-return valve NRV and in increasing of the rotational speed of the rotary distributor. The flow restrictor FR can provide a pressure difference sufficient for opening the non return valve NRV and the main part of the liquid from the hydraulic rotary motor MHR will flow through it to the hydraulic turbine HT. The liquid that enters the hydraulic turbine actuates the turbine causing the rotation of the rotatable spool of the rotary distributor DHR. The rest of liquid passes through the flow restrictor FR to the discharge line. When the feed of fluid in the impact mechanism is resumed, the entire fluid flow sent to the working chamber C2 to advance the impact piston charges the hydraulic accumulator and rushes to the hydraulic turbine at the retreating piston stroke. The rotational speed of the fluid distributor is substantially increased now. The operating of the impact piston maintains the rotation of the rotary distributor at maximum speed.

a)

b)

Fig.4. The hydraulic impact mechanism presented in [1], when: a) both work chambers C1 and C2 are supplied with high pressure fluid flow;

b) the work chamber C2 is discharged and the high pressure fluid flow is directed to the line II.

22


Fig.5. The operating principle of the hydraulic impact

mechanism presented in [2], where: T1, T2 – tanks; PV – hydraulic pressure source; C1,

C2 – work chambers; DHR – hydraulic rotary distributor; DHL – hydraulic linear distributor; MHL – hydraulic linear motor; MHR – hydraulic rotary motor; MRD – hydraulic gear motor; RG – reduction gear.

The Fig.4 b is uncertain and improper taking into account that the fluid flow discharged from the work chamber C2 is adjustable according to [1] and in the line II a certain required pressure is provided to cause the rotation of the hydraulic turbine.

Omitting the hydraulic rotary mechanism MHR from the Fig.5, it can be said that the liquid flow sent from the hydraulic pressure source PV to the hydraulic linear motor MHL is likewise divided into two flows. One liquid flow is conveyed permanently to the working chamber C1 through certain communication channels. According to one of the working positions of the rotary distributor DHR, the rest of the liquid flow is conveyed through it to the working chamber C2. As the hydraulic rotary distributor allows the communication of the working chamber C2 with the hydraulic pressure source PV through the pressure line, the impact piston executes a work stroke and delivers an impact on the drill tool (drill steel or drill bit). The working liquid spent to obtain a certain piston stroke – work stroke – is farther sent to the return line through the rotary distributor DHR back to the tank T1, according to the other working position of the distributor.

In addition it can be used also a rotary hydraulic motor operated by a 4/3-ways control valve. The control valve serves to stop and reverse the rotation of the drill tool. By mounting the hydraulic rotary mechanism in series with the hydraulic percussive mechanism, the liquid flow sent from the hydraulic pressure source PV to the hydraulic linear motor MHL is divided into three flows, Fig.5. The hydraulic fluid used by the rotary hydraulic motor MHR to rotate the drill tool is discharged to the tank T2. To be capable of executing percussive action independently from the hydraulic rotary motor MHR there can be chose two hydraulic pressure sources feeding the mechanism separately. Examples of hydraulic rotary percussive mechanism with two hydraulic pressure sources are [6], [7] and with one hydraulic pressure source are [2], [3]. According to the design structures of both impact mechanism, as the impact piston advances to apply a strike the fluid flow from the working chamber C1 joins to the fluid flow sent to the working chamber C2 resulting in increasing the velocity of the impact piston. By directing the fluid flow from the working chamber C2 to the tank T1 through the rotary distributor DHR the impact piston is forced to do a reverse stroke, as the working chamber C1 continuously communicates with the hydraulic pressure source PV. A representation of the high pressure liquid zones and low pressure liquid zones in the mechanism’s structure for each working position of the hydraulic rotary distributor is presented in Fig.6.

3. Similarities and dissimilarities

Similarities in the described impact mechanisms

Both structures, Fig.1 and Fig.2, incorporate a rotary spool and a differential impact piston delimitating two working chambers, C1 and C2. The drive chamber controlled by the hydraulic rotary distributor is the largest one, C2. The rotary spool presents on the external cylindrical face longitudinal grooves. The longitudinal grooves are disposed as opening through which the drive chamber C2 communicates with the hydraulic pressure source of the impact mechanism and as

23


openings through which the drive chamber C2 communicates with the tank. Both hydraulic rotary distributors operate as 3/2-ways control valve.

Dissimilarities in the described impact mechanisms

The rotary distributor from Fig.1 is mounted separately from the hydraulic linear motor. Conversely, the rotary distributor from Fig.2 is mounted in the hydraulic linear motor. Moreover, the element comprising the rotary spool of the rotary distributor is chose to be the impact piston itself. To be mentioned, the impact piston does not act itself as a valve unlike the valveless impact devices. The rotary spool from the Figl.1 is rotated by means of a hydraulic turbine. The rotary spool from the Figl.2 is rotated by means of a hydraulic gear motor. The hydraulic turbine uses the fluid directed to it from the spent fluid to advance the impact piston.

a)

b)

Fig.6. The hydraulic impact mechanism presented in [2], when: a) both work chambers C1 and C2 are supplied with high pressure fluid flow;

b) the work chamber C2 is discharged and the high pressure fluid flow is directed to the tank. 24


a) b)

Fig.7. The operating principle of the hydraulic impact mechanism presented in [2], modifying the actuating method of the rotary distributor:

T1, T2, T3 – tanks; PV, PV1, PV2 – hydraulic pressure source; C1, C2 – work chambers; DHR – hydraulic rotary distributor; DHL – hydraulic linear distributor; MHL – hydraulic linear motor; MHR – hydraulic rotary motor; HT – hydraulic turbine; RG – reduction gear; FG – flow governor; NRV – non-return valve; TRV –

throttling valve. a) with the hydraulic motor MHR connected to the pressure line; b) with the hydraulic motor MHR connected to the discharge line;

With other words, the element rotating the rotatable spool is driven internally. The hydraulic gear motor can use high pressure fluid flow from a separate hydraulic pressure source or item from the hydraulic pressure source of the impact mechanism. In Fig.2 the method of feeding or acting the hydraulic gear motor is undecided. The longitudinal grooves are set for the structure from Fig.2 to eight equidistantly, alternatively and reversely disposed. According to the description of the rotary spool from [2], as the spool is rotated the longitudinal grooves communicate four by four with the work chamber C2 through radial orifices drilled in the piston body. The arrangement for the longitudinal grooves of the rotary spool from Fig.1 is not revealed.

4. Discussions

To opt for the method of rotating the spool with a hydraulic turbine (a reaction turbine) it is needed to take into account the effect of the mounting a hydraulic accumulator at the relief of the fluid flow from the impact mechanism. In Fig. it is illustrated the structure with the rotatable spool mounted in the piston body and actuated by a hydraulic turbine without including in the impact mechanism housing hydraulic accumulators.

When the hydraulic motor MHR is connected to the pressure line of the hydraulic pressure source PV the hydraulic turbine does not rotate, Fig.7 a. Conversely, when the hydraulic motor MHR is connected to the discharge line as in Fig.7 b the hydraulic turbine HT rotates but at low speed with a stopped impact mechanism. Thus, the hydraulic turbine rotates continuously at low rotations per minute until the impact mechanism is started. As the feeding of the impact mechanism starts, the rotational speed of the hydraulic turbine increases. A redistribution of the liquid sent to the tank T1 can be obtained by means of the flow governor. Accordingly, the working liquid spent by the rotary mechanism to rotate the tool can be sent to the hydraulic turbine through the non-return valve of the flow governor FG. Also, the working liquid

25


spent by the percussive mechanism to strike the tool can be blocked and reflected and further directed to the hydraulic turbine to rotate it. The working liquid conveyed from the hydraulic percussive mechanism to the non-return valve NRV acts to close it in each case presented in Fig.7. The throttling valve TRV controls the fluid flow directed to the tank T1. To modify the rotational speed of the rotary distributor the throttling valve TRV has to be adjusted. In consequence, by controlling the throttling valve TRV of the flow governor FG it is attained an adjustable rotational speed of the rotatable spool. More important it is attained an adjustable impact frequency for the impact piston of the impact mechanism. A fully open throttling valve reduces the fluid flow entering the hydraulic turbine and the rate of rotation of the rotatable spool leading to a minimum frequency of impacts delivered by the impact piston. A fully closed throttling valve increases the fluid flow entering the hydraulic turbine and the rate of rotation of the rotatable spool leading to a maximum frequency of impacts delivered by the impact piston. The frequency of impacts delivered by the impact or strike piston is controlled by varying the rotational speed of the rotary distributor by means of the flow throttling valve TRV. By mounting a hydraulic accumulator on the discharge line similarly as in Fig. 3 the liquid is forced from the working chamber C2 by the impact piston, the non return valve and the throttling valve to the flow governor FG, the hydraulic accumulator AH2 and the hydraulic turbine HT. The hydraulic accumulator AH2 discharges and feeds the hydraulic turbine to maintain a desired rotational speed of the rotatable spool. One simplification of the described construction from Fig.1 is the use of the hydraulic reaction turbine fixed at the rotatable spool of the rotary distributor. The arrangement of the flow governor in the manner of that illustrated in Fig.3 ensures that the rotational speed of the hydraulic turbine can be controlled independently of the hydraulic motor MHR. Moreover, the mentioned arrangement facilitates the conveying of the flow spent by the impact mechanism to the hydraulic turbine and permits the adjustment of the rotational speed of the hydraulic turbine and more of the impact frequency of the impact mechanism. The rotation speed of the hydraulic turbine can be maintained by means of a hydraulic accumulator mounted as illustrated in Fig.3 on the discharge line of the impact mechanism. One simplification of the second described construction, Fig.2, is the mounting of the rotatable spool of the hydraulic distributor in a cylindrical bore made in the piston body, more precisely in the smaller piston rod. Channels are provided in the piston body and in the smaller piston rod to assure a first communication of the working chamber C2 with the internal piston cavity (the cylindrical bore) and a second communication of the internal piston cavity with the discharge line through which the fluid flow is sent to the tank. The said simplifications allow reduction in weight and size and furthermore simplify the pumping station, reduce the rate of high pressure liquid consumed, increase the efficiency and improve the reliability of the impact mechanism.

REFERENCES (Arial, 11pt, Bold)

[1] J.I. Neroznikov, N.N. Shvets, s.a., “Hydraulic drilling machine“, Patent Number(s): US 5064003, 1991 [2] L. Vaida, C. Kozma, Generator hidraulic de vibraţii pentru perforatoare hidraulice rotopercutante,

registered to OSIM: A/10018/2012 [3] C. Kozma, „A constructional and functional improvement in hydraulic rotary percussive drill”, Calimanesti-

Caciulata, Romania, 9-11 November, 2011, „Proceedings of 2011 International Salon of Hydraulics and Pneumatics – Hervex”, ISSN: 1454-8003

[4] C. Kozma, L. Vaida, „Hydraulic schemes for impact devices. A control system for impact mechanisms using a rotatable distribution valve – Part 2”, Calimanesti-Caciulata, Romania, 7-9 November, 2012, „Proceedings of 2012 International Salon of Hydraulics and Pneumatics – Hervex”, ISSN: 1454-8003

[5] R.J. Perraud, “Rotary percussion hydraulic drilling machine“, Patent Number(s): 4291771, 1981 [6] V. Uitto, “Method and arrangement for controlling percussion rock drilling”, Assignee: Sandvik Tamrock

Oy, Pat. No.: US 2005/0006143 A1, 2005 [7] B. Cadet, “Device for hydraulic power supply of a rotary apparatus for percussive drilling”, Assignee:

Montabert S.A., Pat. No.: US 6883620 B1, 2005

26


HYDRAULIC AND PNEUMATIC CYLINDER FAILURES , THE EFFECT OF

FLUID CLEANLINESS ON COMPONENT LIFE Patrick Adebisi Olusegun ADEGBUYI1, Ioan-Lucian MARCU2

1Faculty of Engineering, Lagos State University, Ojo. P.m.b 1087 Apapa –Lagos, Nigeria e-mail: [email protected], [email protected] 2Technical University of Cluj Napoca, e-mail: [email protected]

Abstract: This article reviews the situation of hydraulic and pneumatic failures of cylinders. It also identifies the various component malfunctions that may lead to these failures. Furthermore, the effect of fluid cleanliness on cylinder component life cycle was examined.

Keywords: Hydraulic, Pneumatic, Failures, Cleanliness Life

1. Introduction

The application of cylinders may allow fluids such as cutting fluids, wash down fluids, etc to come in contact with the external area of the cylinder. These fluids may attack the piston rod wiper and or the primary seal and this must be taken into account when selecting and specifying seal components.

Dynamic seals will wear. The rate of wear will depend on many operating factors.

Wear can be rapid if a cylinder is miss-aligned or if a cylinder has been improperly serviced. Seal wear is very important in the application of cylinders and could lead to failure.

Piston-rods: Possible consequences of piston-rod failure or separation of the piston rod from the piston include but are not limited to.

- Piston rod or attached load thrown off at high speed

- High velocity fluid discharged

- Piston rod extending when pressure is applied on the piston retract mode

Piston rods or machine members attached to the piston may move suddenly and without warning as a consequence of other conditions occurring to the machine such ass:

- Failure of the pressurized fluid delivery system ( hoses, fitting, valves, pomp, compressors) which maintain cylinder position

- Catastrophic cylinder seal failure leading to sudden loss of pressurized fluid

The use of cushions should be considered for cylinder applications when the piston velocity is expected to be over 4inches/second. These cushions are normally designed to absolve the energy of a linear applied loud.

A rotating masse has considerably more energy than the same masse moving in a linear mode.

All these could lead to hydraulic and pneumatic cylinder failure.

Proper alignment of the cylinder piston rod and it’s matting components on the machine should be cheeked-in both the extended and retracted positions.

Improper alignment will result in excessive rod stand and/or cylinder bore wear.

27




Another source of failure is internal leakages. Piston seal leak (by-pass) 1-3 cubic inches per minute is considered normal for piston ring construction. Virtually no static leak with lip seal seals on piston should be expected. Piston seals wear is a usual course of piston seal leakage and eventual cylinder failure.

Contamination in a hydraulic system can result in a pored cylinder bore, resulting in rapid seal wear which may lead to cylinder failure.

2. Effect of fluid cleanliness on component life

This is an important factor for consideration in cylinders operating in an environment wear air drayed materials are present such as : fast draying chemicals paint or weld splatter or other hazardous conditions such as excessive heat should have shields installed to prevent damage to the piston rod and piston rod seals.

Many factors can reduce the service life of hydraulic components. Contamination of hydraulic fluid by insoluble particles is one of these factors. To prevent particle contamination from cutting short component life, an appropriate fluid cleanliness level must first be defined and then maintained on a continuous basis. [5]

Particle Contamination And Its Consequences Particle contamination in hydraulic fluid accelerates wear of system components. The rate at which damage occurs is dependent on the internal clearances of the components within the system, the size and quantity of particles present in the fluid and system pressure. Typical internal clearances of hydraulic components are shown in table 1.

Table 1.

COMPONENT TYPE TYPICAL INTERNAL CLEARANCE IN MICRONS

Gear pump 0.5 – 5.0 Vane pump 0.5 – 10 Piston pump 0.5 – 5.0 Servo valve 1.0 – 4.0 Control valve 0.5 – 40 Linear actuator 50 - 250

Particles larger than a component's internal clearances are not necessarily dangerous. Particles the same size as the internal clearance cause damage through friction. But the most dangerous particles in the long-term are those that are smaller than the component's internal clearances. Particles smaller than 5 microns are highly abrasive. If present in sufficient quantities, these invisible 'silt' particles cause rapid wear, destroying hydraulic components.

28


Quantifying Particle Contamination Some level of particle contamination is always present in hydraulic fluid, even in new fluid. It is the size and quantity of these particles that we are concerned with. The level of contamination, or conversely the level of cleanliness considered acceptable, depends on the type of hydraulic system.[5] Typical fluid cleanliness levels for different types of hydraulic systems, defined according to ISO, NAS and SAE standards, are shown in table 2.

Table 2.

TYPE OF HYDRAULIC SYSTEM

MINIMUM RECOMMENDED CLEANLINESS

LEVEL

MINIMUM RECOMMENDED

FILTRATION LEVEL IN

MICRONS (βχ ≥ 75) ISO 4406 NAS 1638 SAE 749

Silt sensitive 13/10 4 1 2 Servo 14/11 5 2 3-5 High pressure (250–400 bar)

15/12 6 3 5-10

Normal pressure (150-250 bar)

16/13 7 4 10-12

Medium pressure (50 -150 bar)

18/15 9 6 12-15

Low pressure (< 50 bar)

19/16 10 - 15-25

Large clearance 21/18 12 - 25-40 ISO 4406 defines contamination levels using a somewhat complicated dual scale numbering system. The first number refers to the quantity of particles larger than 5 microns per 100 milliliters of fluid and the second number refers to the number of particles larger than 15 microns per 100 milliliters of fluid. The complicated part is that the quantities of particles these numbers represent are expressed as powers of the numeral 2. For example, a cleanliness level of 15/12 indicates that there are between 214 (16,384) and 215 (32,768) particles larger than 5 microns and between 211 (2,048) and 212 (4,096) particles larger than 15 microns, per 100 milliliters of fluid. Defining A Target Cleanliness Level As an example, let’s assume that we have a normal-pressure system and using table 1.2 we define our target cleanliness level to be ISO 16/13. Having established the minimum fluid cleanliness level required for acceptable component life in this type of system, the next step is to monitor the actual cleanliness of the fluid to ensure that the target cleanliness level is maintained on a continuous basis. This involves taking fluid samples from the system at regular intervals and testing them for cleanliness.

Testing Fluid Cleanliness There are two ways of testing fluid cleanliness. The first involves sending a fluid sample to a laboratory for analysis. The lab results contain detailed information on the condition of the fluid. The information normally included in a fluid condition report, along with typical targets or alarm limits, are shown in table 3.

29


Table 3.

CONDITION CATEGORY RECOMMENDED TARGETS

OR ALARM LIMITS Fluid cleanliness level Within targeted range chosen for the

system or recommended by the manufacturer (ISO 4406)

Wear debris level (Al) 5 ppm, (Cr) 9 ppm, (Cu) 12 ppm, (Fe) 26 ppm, (Si) 15 ppm

Viscosity ± 10 % of new fluid Water content < 100 ppm Total Acid Number (TAN) + 25% of new fluid Additive level − 10% of new fluid

The second way to test a fluid’s cleanliness level is to use a portable, electronic instrument designed for this purpose. This method is convenient and results are almost instant, however it shouldn’t be considered a total substitute for lab analysis because the results do not include wear debris levels, viscosity, water content and other useful data. But when the two methods are used in combination, the frequency of lab analysis can be reduced. Whichever method is employed, it is important that the equipment used to capture and contain the sample is absolutely clean. If you are taking multiple samples from different systems, take care not to cross-contaminate one fluid sample with another, and never take samples from drain plugs or other low lying penetrations in the system, otherwise the results will be unreliable. Ideally, samples should be taken from the return line, upstream of the return filter, with the system working at operating temperature.

3. Conclusion Monitoring and maintaining fluid cleanliness involves a continuous cycle of testing and corrective action in order to reduce component failure .Cleanliness is also an important factor hence cylinders should be protected from contaminants entering the ports. Also before making connections to cylinder ports, piping should be thoroughly cleaned to remove all chips or burns which might have resulted from threading or flaring operations.

4.References [1] Banyai, D., Vaida, L., (2009), Synoptic view of the latest trends in hydraulic actuation, Buletinul Institului Politehnic din Iasi.

[2] Merkle et all, (1990), Hydraulics, Festo Didactic KG.

[3] Nekrasov, B., (1969), Hydraulics, Peace publishers, Moscow.

[4] Parker, (2005), Cylinder safety guide, Del plaines.

[5]Drumea P., Matache G., Lepadatu I. – Metode de crestere a fiabilitatii utilajelor prin ungerea cu doze precise de lubrifiant – HERVEX 2001, pag.38-42,ISSN1454-8003

[6] Pop,I. et all, (1999) Conventional Hydraulics, U.T Pres, Cluj-Napoca.

[7] Pop, I., et all, (1999), Modern Hydraulics. Pneumatics, Ed. U.T Pres, Cluj-Napoca.

30


PNEUMATIC PRESSURE SERVOREGULATOR WITH PIEZOELECTRIC ACTUATION

PhD.Eng. Gabriela MATACHE1, PhD.St.Eng. Ioana ILIE2, PhD.Eng. Radu RADOI3

1,2,3Hydraulics and Pneumatics Research Institute, Bucharest Romania, [email protected] 1. Introduction For achieving a modern and competitive economy, it is required to create highly improved automation equipments using some performant tehnological methods. Until now the domain of regulation and control equipment has known a remarkable evolution, reaching max.operational frequencies below 10 Hz. These kind of equipments have the following advantages: - are plain and compact; - perform an easy and very accurate adjustment of the output pressure, as a result of the precision of positioning of the electro pneumatic convertors used (couple motor, electromagnet or piezoelectric motors) - present a short and speedy feedback; - low electric consumption at actuation; - may perform the regulation of the output pressure from distance on a wide range; - maintains constant the output pressure. The firms which manufacture such devices are developing a new domain of mechatronics, that of active and intelligent materials of the following type: piezoelectric, magnetostrictive, with shape memory , integrated in the servoequipments elaborated[1]. 2. Presentation of the pneumatic pressure servoregulator with piezoelectric actuation The pneumatic pressure servoregulator with the functional role of adjusting the output pressure, depending on the electric drive size (given tension Uc, prescribed or softare processed), promoted by the present Ph.D final thesis, use in the actuation level a bimorph lamella type piezoelectric actuator, fig. 1. In what regards its structure, the servoregulator proposed has a classic mechanical structure – aimed objective, similar with that of the piloted pressure regulators at which the drive pressure was installed manually, an electromechanical convertor (piezoelectric actuator) a pressure trnasducer and an electronic drive system (SEC)- for generating and installing the drive tension Uc. The modern constructions of such regulators have the pressure transducer and the electronics integrated (encapsulated) which allowed the achievemtn of some compact structures, with a low electric consumption and high operational performance, which confers them the character of mechatronic products (due to data processing and an adequate software for the intelligence level). The operation of the pneumatic pressure servoregulator with piezoelectric lamella in relation with the constructive and functional diagram from fig.1 is described below. The value of the regulated output pressure may be monitored by means of a pressure transducer, whose signal is compared with the input electric signal of reference.

31



Fig. 1 If the value of the regulated output pressure is lower than the corresponding value of the signal of reference, the electronic drive system SEC will generate a higher drive signal Uc which determines the approach of the flap or piezoelectric lamella and implicitly the increase of pressure in the actuation chamber pcom. This increase leads to the opening of the circuit entry exit, by the displacement downward of the main valve 3 which means a larger flow section, a higher flow which determines an increase of the output pressure in the volumes from below the servoregulator. In the cases in which output pressures are higher the direction of the moves and their effects are reverse. The pneumatic pressure servoregulator consists of the body 4 which has two ports ‘’i’’and ‘’e’’(of equal dyameters), corresponding to the entry and exit of the compressed air. In the body 4 is mounted the main valve 3 which is maintained in normal position - Uc=0 closed by the spring 2. When is pressure supplied on port ‘’i’’ this enters both chamber ‘’a’’ and the port ‘’b’’ the crossing track 9 , calibration nozzle 10, through the nozzle 17 reaching the piezoelectric flap- lamella 15. In the same time a certain drive pressure pcom corresponding to the input pressure is installed through

32


port ‘’f’’and in the pilot chamber ‘’c’’ delimited by membrane 5 of the servoregulator. Subsequently the mechanical subassembly from below the mebrane 5 is identical with those of the piloted pressure regulators. In the moment when the piezoelectric lamella (flap)15, fixed on the body 6 of the servoregulator by means of the support plate 12 with some clamping screws 14, is supplied with a drive tension,it occurs a distortion of the lamella meaning an approach of the nozzle 17, where according to the principle ‘’nozzle - flap”, is generated a proportional drive pressure - pcom required for achieving the membrane imbalance 5. Cause of this membrane imbalance 5, it is reached by means of the rod 18, the move away of the main valve 3 from its seat enclosed in part 19. In this moment the section of the flowing circuit between the chamber ‘’a’’ and port ‘’e’’ opens, obtaining an increasing output pressure. When it is reached the desired value at the exit pressure, permanently present in chamber d’’of the body 4 by means of port ‘’m’’ it takes place a displacement of the membrane 5 corresponding to a new balance position of the forces on mebrane. This balance of foreces will have effect upon the displacement of the main valve 3, leading to the installation of a flowing section between valve and the seat 19 corresponding to the regulation of the output pressure at the necessary value, imposed by UC.[3]. This is the operational mode involving pressure regulation. Additionally the servoregulator ensures the maintaining of a constant value of the regulated pressure, irrespective of the air consumption from down under. If at exit is connected a closed chamber of constant volume, case which is rare in use, but is extreme for the regulation function, the output pressure increase tendency above the regulated value, leads to the diminishing of the flowing section seat – chamber, even until its entire closing. If downstream the consumption-flow of air in a chamber with variable volume is stabilized when installing the regulated pressure, the flowing section is preserved. In the case when the air consumption decreases, the flow from the stabilized mode becoming in excess, this leads to an increase of the regulated pressure. This output pressure increase tendency permanently installed in chamber ‘’d’’ will unbalance the membrane which produces the decrease of the flowing section, meaning a subsequent flow decrease at a level equal with consumption followed by a preservation of the regulated pressure value. In evolutions of reverse directions, when the consumed flow increases appears a tendency of diminishing the regulated pressure, the drive pressure installed by Uc generates an increase of the flowing section, meaning a flow increase as well and its accordance to the value of the consumed one. 3. The calculation of design of the main level and the equations afferent to the mathematical model Congruent with the structure from fig. 1 for the sizes which appear in the mathematical model, are used the following symbols: P0, T0 – air pressure and temperature in normal conditions; P1, T1 – input pressure and temperature; P2, T2 – output pressure and temperature; P3, T3 , V3 – pressure, temperature and volume from the drive chamber; Aij – the geometric areas of the functional flowing sections; αij- the flow coefficients corresponding to the flowing sections;

jim - the mass air flows through sections; x – the opening variation of the flap, equal with the distortion of the piezoelectric lamella;

y,y,y - position, speed and acceleration of the central mobile assembly (valve 3, rod 18, membrane 5); Farc=F0 +k .y – the force of the compression spring 2; m = m1 +m2 +m3– the mass of the mobile assembly valve + membrane; m3 = 1/3 spring s mass.

33


All the status parameters like pressure P=p+1,013 temperature T = t[0C] + 273,15[K] are at their absolute values. Other notations will be defined in the calculation diagrams or during the elaboration of the model.[3] Are adopted the following simplifying hypothesis, unanimously accepted in pneumatics [1]: - The instant massic flow through the flowing sections it is considered equal with that from the

stationary mode, for the same value of the relation between the downstream-upstream pressures;

- The air flow through the ports takes place without any heat exchange with the outside environment (adiabatic evolution n=k=1,4);

- The air evolution in the functional chambers in transitory mode occurs without any heat exchange (adiabatic evolution);

- The influence of the temperature variations from chambers upon the dynamic parameters (flow, pressure) is negligible, so that the temperatures are considered constant and equal with the normal temperature (T= 293,15 K; t = 200C );

- The pressure losses on internal circuits of the servoregulator are negligible, taking into account their relatively short length and the low air viscosity;

- The elastic force of the nemetallic membrane is negligible (its distortion is stood by the geometrical shape, not by the material), the active surfacxe is considered constant and equal with the initial geometreic value;

- The flow coefficients are considered constant; - The flowing forces are considered negligible in comparison with the forces developed by

pressures. As functional parameters imposed by rule, according to the theme of design of a pneumatic pressure servoregulator, are: - presure regulated at output: (p2)min... (p2)max=(0,05....5,8) bar - supply input pressure: (p1)min... (p1)max= (1...6) bar a. The calculation of the nominal dyameter - Dn Taking into account that the flow m must be obtained in the most disadvantageous conditions of flowing through the sections from the input ouput route of the servoregulator, are first used the flow calculation relations:[1]

−

⋅⋅α⋅=

skg

PP

PP

TPS15545,0m

21

7117,1

i

e

4235,1

i

e

i

iD (1)

- for 1PP528,0

i

e ≤< (subsonic flowing mode);

⋅⋅α=

skg

TPS04046,0m

i

iD , (2)

- for 528,0PP0

i

e ≤< ( sonic flowing mode)

where: Pi, Pe

2m

N are the input and namely the output pressures from the servoregulator

(corresponding to pressures P1, P2), S [m2] is the nominal flowing section, Ti [K] air input temperature Ti = T1 = 293,15 [K], flow coefficient αD = 0,6...0,8 depending on the total loss of pressure on the input-output route, predominating the local pressure losses.

34


On this route there are two sudden variations of section, a decrease from the nominal section

4DS

2n

n⋅π

= at the flowing section dydS ⋅⋅= 3π and an increase from S to Sn. In the dame time

there are two changes in what regards the flowing direction with angles of 900 and three transformations of section shape (of jet). Since P1 [i T1 are the same, it means that only the

function

1

2

PPF from the paranthesis of the subsosnic flow relation (1) remains in discussion. Its

value decreases at the increase of the P2/P1 relationship value. The numeric coefficients of the relations from the subsonic flowing mode (1) and from the sonic flowing mode (2), even if different, the lowest flow it results to be that given by the relation (1). For the required massic flow: m , at

pressure pi in the most disadvantageous case of relation between pressures i

e

pp

, Ti=T1=293,15 K

and αD = 0,7, nominal section Sn [ ]2m must have the value:

7117,1

1

2

4235,1

1

21

1

15545,0

−

⋅⋅⋅

⋅=

PP

PPP

TmS

D

n

α

Knowing the nominal section it is calculated the nominal dyameter Dn:

ππ n

nn

nS

DD

S⋅

=⇒⋅

=4

4

2

b. The calculation of the maximum opening stroke – yd . the continuity of the flowing section ( fig. 2) imposes:[2]

dDn

D ydD⋅⋅⋅=

⋅⋅ 32

2

1 4παπα , or correspinding to the real flowing sections:

dn ydD ⋅⋅=⋅ 32 3,264,0 (3)

For the circular section as it is Sn, αD1= 0,815, and for the flowing section plane valve- seat S (the lateral surface of a cylinder) αD2= 0,732. Although the section from downstream of S is of a ring shape, is affected by the rod dyameter – d4, generally it is adopted d3= Dn. With the relation 3 it is calculated yd:

32

21

4 dD

yD

nDd ⋅⋅⋅

⋅⋅=

παπα

c. The dimensioning of the main valve (fig. 2). Being known the dyameter d3 it is calculated from the condition of resistence at the contact pressure of the non metallic sealing zone (from rubber) the dyameter d22. The most detrimental situation is when the valve seals the seat and p2=0. In relation with the distribution of pressure p1 on the valve, the dimensioning relation becomes:

1ac

ac322 pp

pdd−

⋅= (4)

Because it was ignored the force Farc of the spring, in reality reduced in relation with the force developed by pressure p1 it will be taken the admisible contact pressure: pac = (15...18).105 N/m2.

35


Due to constructive considerations, for placing the rubber ring in the valve (by means of vulcanization, pasting or crimping) the dyameter d2 it is dimensioned at the size: d2 = d22 + (3...4) mm. The calculation of the dyameter d1 it is made taking into account the balance of forces on the valve. For diminishing the forces developed by the different pressures p1 and p2 the valve is balanced (pressure p2 is brought under the valve too), solution which imposes the sealing of this chamber with a seal ring.For a more accurate evaluation of these forces it was stated the hypothesis that the transition from pressure p1 to p2 it is not step but ramp type, with the geometrical limits defined by the dyameters of the valve seat d22, namely d3 [3]

Fig. 2

( ) ( ) ( ) ( )

)(2222

444424

1221

242

222

223

223

1222

1

2212

122

124

23

223

222

21222

221

ppddpdp

dp

dp

dp

pddd

pdd

pdd

ppddp

−⋅=⋅−⋅+⋅+⋅−⋅−⇒

⇒⋅⋅

+−⋅⋅

=−⋅+−⋅⋅+

+−⋅πππππ

)2(2

)(2

)( 24

23

222

223

222

121

21 ddd

pdd

pppd ⋅−+−+⋅=−⋅

36


Knowing the numerical values of the dyameters d22, d4,d2,d3, it results the following dimensioning formula:

212121 4650)( ppppd ⋅−⋅=−⋅ (5)

Generalized for any ∆p=pi-pe,the relation (5) becomes :

pppppp

∆−=∆=−

12

21

ppppppd ∆⋅+⋅=∆−⋅−⋅=∆ 464)(4650 111

21 , meaning

pppd ∆⋅+⋅=∆ 464 12

1 (6) The friction force from the seal, input by this ring will have the value:

(7) )( 211 ppbdFe −⋅⋅⋅⋅= πµ d. The force developed by pressure on valve and the friction force from the seal Considering the sense of force Fs developed by the pressures that charge it (in bars), with preset geometric elements (în cm), force Fs (în daN) has the expression:

( ) ( ) ( ) ( )

4444242

122

1221

24

232

23

222

12222

221 dpddpddpddppddpFs

πππππ⋅−−⋅−−⋅+−⋅

++−⋅= (8)

e. The dimensioning of the valve rod. The dyameter of the rod d4 is chosen constructively. For having a domain of the regulated output pressures close to the drive pressure values, the rod dyameter should be smaller. f. The calculation of the force on the membrane of the pneumatic pressure servoregulator In fig. 3 is shown the distribution of pressures on the elastic membrane of the servoregulator. Upon the assembly membrane rod act two forces developed by pressure: on the upper part of the membrane the drive pressure p3 and on the lower part the output pressure (regulated) p2.

Fig. 3

37


As a result the force generated on the membrane is :

Fm = ( )2

24

25

3

25

44pddpd

⋅−⋅

−⋅⋅ ππ

4. The study of the piloting-command level and the equations of the model The force developed by the piezoelectric lamella, used as electro mechanic convertor was calculated and checked experimentally. From the relation (9), it results the equation of connection displacement lamella x – drive tension Uc of shape:

⋅

+⋅⋅⋅⋅⋅⋅⋅−

=3

224 11

231 hT

SIEUbld

x lE

zp

c (9)

The equation force bimorph – drive tension:

lhbUEd

F cp ⋅⋅⋅⋅⋅−= 312

(10)

Fig. 4

38


Fmax bore by the piezoelectric lamella, studied experimentally demonstrated that for the max.drive pressure it is sufficient even at values of F/2, for balancing the forces given by these pressures:

pFF ⋅≈ 2 (preliminarily imposed condition).[2] In the same time in relation with the max.displacements of the bimorph –x the domain of the openings flap- lamella it was chosen as a safety measure between 0 and x/2. For the mode in which it was considered the distribution of pressure on lamella - fig 4, force Fp for x0=0, when p3=p3max≅p1

22max3 100

4−⋅⋅⋅= dpFp

π (11)

where: p3max [bar]; d0 [mm] The max.drive pressure in relation with the force F developed by the lamella (13) may be:

[ ]barxF

xF

xF

dFp 592,1

2)2(2

4/2/

20

20

20

20

max3 ⋅=⋅⋅

=⋅⋅⋅

=⋅

=πππ

(12)

Where is maintained the condition that pFF ⋅≈ 2 , with F [N] and x0 [mm]. The calcxulation relationships of the flow in sonic mode, respectively subsonic mode, for the hypothesis of the adiabatic evolution k=1.4, are [1]: For the nozzle – flap couple (flowing section A30) due to the flow of air right in the atmosphere in most of cases the flowing mode is sonic (the result being obtained experimentally too):

11

30330 1

2 −+

+⋅⋅

⋅=

kk

kk

RTAPm (13)

The flowing mode through section A13 (confirmed experimentally) proves to be constantly subsonic, cause never 528,0/ 13 ≤PP (pressures in their absolute value) [3]

−

⋅

−⋅

⋅=

+k

kk

PP

PP

kk

RTAPm

1

1

3

2

1

313113 1

2 (14)

From the continuity of flows, for d0 =0,8 mm in stabilized flowing mode: 3013 mm = (subsonic mode – sonic mode) may be found xl for different absolute input and drive pressures P1 and P3 with the relation:

7117,1

1

34235,1

1

3

3

18553,0

−

⋅⋅=

PP

PP

PPxl (15)

39


4. The mathematical model of the pneumatic pressure servoregulator For structuring the mathematical model of the pneumatic pressure servoregulator, it is used the schematic presentation from fig. 5 in which are configured the forces (previously determined) interferring in the operation of the servoregulator. The balance of forces at the level of the central mobile subassembly (valve - rod - membrane) has the following form: 0)y(sign)FF(FGFF iearcmaxms =⋅+−−++ (16) For eliminating pressure p3 from the equations of the mathematical model and their correlation with Uc – the drive-supply tension of the lamella:

2883,0

3

15765,0

3

1c

300l P

PPP

8553,0U1009,3xxxx

−

⋅=⋅⋅−=−= −

(17)

Fig.5

40


5. Conclusions The pneumatic pressure servoregulator differs from a classic proportional regulator, at which the regulated pressures are installed due to a mechanic prompt, performed by hand. In its structure it was introduced a piloting step, operationally based on the pneumatic circuit nozzle-flap, which determine drive pressure p3 depending on the position of the flap –xl materialized by a piezoelectric actuator with bimorph lamella. The study of this circuit, in relation with the real flow modes of the air through its two sections of reference (A13 [i A30) allowed the determination of function p3 = p3(xl), which by means of the force of distortion of the lamella: x = x(Uc) and its constructive connection : xl=x0-x, was brought to the form p3=p3(Uc). The study of the piloting step,has shown that after certain openings xl the drive pressure - p3 does not decrease below certain values. In this situation for tensions Uc< 30 V have no regulation effect in the field Uc = (30...60) V are obtained desired regulated pressure values. The mathematical model set and its numerical simulation proves the theoretical possibility of achieving a pneumatic pressure servoregulator with proportional regulation characteristic. On the servoregulator obtained after design, the mathematical model allowed with an error caused by approximations, the achievement in theory of a proportional characteristic of regulation, whose validity will be subsequently studied on a physical model realized at the dimensions obtained in the stage of design of the pneumatic piezoelectric pressure servoregulator. REFERENCES [1[ Radcenco, Vs., Alexandrescu, N. – Calculation and design of the pneumatic elements and diagrams of automation, Technical Publishing house Bucuresti 1985 [2] Belforte, G., Gastaldi, L., Sorli, M. – Gli azionamenti piezoelettrici nel comando di valvole pneumatiche – O+P, septembre1998 [3] Matache, G – The study and improvement of the equipement regulating pressure in the pneumatic systems – Ph.D thesis

41


MODELING OF A THREE WAY ROTATABLE FLUID DISTRIBUTOR USED

TO COMMAND AND CONTROL A HYDRAULIC ROCK DRILL

CLAUDIA KOZMA1, BANYAI DANIEL VASILE2 1,2.Tech.Univ. of Cluj-Napoca, Faculty of Mechanical Engineering, Department of Thermal Engineering, 3400 Cluj-Napoca, Romania; e-mail: [email protected] Abstract: A hydraulic command system comprising a combination of two hydraulic half bridge elements of type A and E is analyzed. The system can be used to command and control an impact mechanism for hydraulic rock drills. The A+E half bridges combination reflects the presence of the active and passive hydraulic resistances. Under the form of a three-way valve with angular command displacement, the A+E circuit is used to control a hydraulic differential motor. Unlike a linear valve, for this rotary valve, shock waves are reduced. Moreover, its positioning in a structure, as it is proposed, leads to minimized parasitic hydraulic capacities. Two methods of calculus the circular section area through which the flow crosses the rotary valve are developed. The characteristics of the three way rotary valve are expressed mathematically and graphically. The hydraulic rotatable distributor – linear hydraulic motor subassembly is mathematical analyzed. Key words: hydraulic rotary valve, three ways valve, pressure-flow curves, half bridge, hydraulic resistance, steady-state characteristics, valve coefficients, area 1. Introduction A hydraulic impact mechanism presented in [1], [2] and [12], is analyzed. The percussion piston of the impact mechanism is controlled by use of an innovative command and control structure which is a rotatable distributor. The system formed by the rotatable distributor and the hydraulic motor of the impact piston minimizes the parasitic hydraulic capacities, the weight, and improves the dynamic behavior of the piston. The present paper is a completion of the mathematical modeling from [3].

2. The hydraulic rotating distributor A rotatable spool which is to be patented is presented and analyzed in figure 1. The rotatable spool presents axially slots, notated with P and T, made circumferentially and positioned alternatively and reversely. These slots operate as input and return flow distribution slots.

Fig.1. A rotating spool 1 with: T and P –

slots; u – shoulder; s – shaft.

42



The sleeve of the rotatable spool is constituted by the prominence bp of the piston 2, as suggested in figure 2. Thus, the rotatable spool 1 is contained within the impact piston 2 and the shaft s imparts a rotation movement to the spool. The rotation motion of the rotatable spool can be provided by means of a hydraulic motor, an electric motor or a hydraulic turbine, through the shaft s. The slots P and T communicate with the working chamber C2 of a hydraulic linear motor MHL, as in figure 2 and figure 3, by means of a number of orifices OK3 realized

through the piston prominence bp. The circumferentially slots are brought alternatively in communication through four orifices OK3 with the chamber C2. Thus, the working chamber C2 is fed with hydraulic fluid through the orifices OK3 and the evacuation of the fluid from the chamber C2 is realized through the same orifices OK3. Until the chamber C2 is connected to the return line through the slots T and the orifices OT, a shoulder u confines the fluid fed through the orifices OS, the annular chamber delimitated by the shoulder u and the spool head acts as an accumulator. A hydraulic scheme of the subassembly rotary valve – linear hydraulic motor is illustrated in figure 3. The subassembly is proposed by [1] to be used in rock drilling, with a proper rotary mechanism, hydraulic accumulators and other hydro-mechanical elements needed to perform the drilling process. By use of the four inlet slots P and four return slots T of the rotatable distributor DHR (figure 3), the working chamber C2 is supplied intermittently with hydraulic fluid from a hydraulic pressure source PV and consecutively returned to the tank T. By supplying fluid under pressure and rotating the spool 1, the piston executes a linear oscillatory motion. A safety valve SS is mounted in the high pressure circuit. In figure 4, a hydraulic scheme of the subassembly rotary valve – linear hydraulic motor outlines the command and control structure of the motor MHL. The hydraulic resistances Ri and Re of the command half bridge have variable flow cross areas but there can’t be made displacement commands as for a linear distributor. When a stepper motor is used, working positions as illustrated in

Fig.3. The hydraulic scheme of the system

comprising: ME – electrical motor; CPV – pump coupling; PV – volumetric pump; SS – safety valve;

T1, T2 – tank; DHR – hydraulic rotary distributor; MHL – linear hydraulic motor; C2, C1 are the motor

chambers.

Fig.2. The rotating spool positioned in the body of a motor piston, where: 2 – the impact piston; r – piston rod; bp – piston prominence; OK3, OS, OT –

orifices.

43


figure 5 can be obtained. Using other type of motor the opening and the closing of the orifices OK3 are realized continuously. The command half bridges is of type A+E. The input resistances Ri are formed by the slots P and the piston prominence bp. In a similar manner, the discharge slots T form the output resistance Re when the motor chamber C2 is connected to the tank.

A type hydraulic bridge is represented by two hydraulic resistances both with variable sectional cross areas. E type hydraulic bridge is always used with hydraulic linear differential motors [13]. This combination of half bridges A+E can be studied with the rest of the combinations (matrix of hydraulic bridges) in [4], [5] where they are symbolically and constructively presented and analyzed. In figure 5, a radial view of the spool-piston assembly is presented correspondingly to the two working positions of the rotating distributor.

Fig.6. Radial view through the rotatable distributor

DHR,where: θv – the central angle of the rotary spool; rv – the

spool radius.

Fig.7. Top view of the spool slot, where: SK3 –

the section of the orifice OK3; θK3-1, θK32 – central angles.

a) the feeding position; b) the exhaust position. Fig.5. The working positions of the rotatable distributor with their corresponding sectional views

through the rotating spool and its sleeve.

Fig.4. The A+E half bridge and the hydraulic linear

motor MHL, where: QC1, QC2 – supply/exhaust flow; Ri, Re – hydraulic resistances; QDHR – the distributor input

flow; QS, pS – supply flow, respectively, supply pressure.

44


3. Equations and mathematics

The flow equations of the hydraulic rotating distributor are:

( ) ( )

( )

−−≥≥⋅⋅⋅

+≤≤−⋅⋅⋅=

brxpxAC

brxppxACQ

KvCvKd

KvCDHRvKdC

323

3232

202

202

ρ

ρ (1)

Where: QC2 – the flow sent to the motor through the distributor DHR; Cd – the discharge coefficient; AK3 – the flow passing area through an orifice OK3; xv – the spool displacement; α is the piston area ratio; pDHR – the pressure of the fluid delivered to the distributor DHR; pC2 – the pressure of the fluid sent to the motor MHL; ρ – the fluid density; b – the slot wide; rK3 – the radius of the orifice OK3. The notations that appear in the flow expressions (1) are presented also in figures 4, 6 and 7. In figure 6 and figure 4 it is illustrated a radial view through the impact piston – rotating spool subassembly, revealing four orifices OK3 displayed circularly and eight grooves. From the top view of one spool groove, figure 7, the parameters needed in cross area computation are marked.

3.1 The calculus steps of the flow passing area

The spool displacement is an instant variable that can be calculated regarding the central angle of the rotary spool, θv, using the expression (2).

°⋅⋅

=180

vvv

rx

θπ (2)

The parameter xv is used also to determinate the flow passing area AK3. For xv varying between 0 and 2rK3+b, the effective passing area to motor chamber can be calculated with the expressions (3), where central angles θK3-1, θK32 are in radian. The variation of the dimensionless flow passing area is presented in figure 8. This variation begins with a full obstructed orifice OK3 and by keeping rotating the spool in the same direction the area function will have a peak when the orifice OK3 is full opened.

Fig. 8. The variation of the passing area corresponding to one spool slot.

45


( )

+≤≤+

−−⋅−⋅

−−⋅

+≤≤

+−−

+−

−⋅

≤≤

−−

−

−

−

+−−

+−

−⋅

≤≤

−−

−

−⋅

≤≤

−⋅−⋅

−⋅

=

brxbrr

brxr

brx

brxrr

bxrr

bxr

rxb

rrx

rrx

rbxr

rbxr

bxrr

rxr

rx

r

rxr

xrr

xrr

xA

KvKK

Kv

K

KvK

KvKK

vK

K

vK

K

Kv

K

Kv

K

Kv

K

vK

K

vK

K

vKK

Kv

K

Kv

K

KvK

vK

K

vKK

vK

333

3

3

323

333

3

3

3

23

3

3

3

3

3

3

3

3

3

23

33

3

3

3

23

33

3

3

323

3

2,arccos2sin21

21arccos2

2,2

arccos2sinarccos2

2,

2

arccos2sinarccos2

2

arccos2sinarccos2

,2

arccos2sinarccos2

0,arccos2sin21

21arccos2

π

π

π

(3)

By substituting in figure 7 the central angles θK3-1, θK32 with 2θK3-1, respectively θK32, the expression (3) can be rewritten as:

( ) ( )( )

⋅<<⋅+−⋅

<⋅−⋅=−−−

−−−

3323232323

313131323

32,cossin

,cossin

KvKKKKK

KvKKKKvK

rxrrrxrxA

θθθπθθθ (4)

The functions sine and cosine from the expression (4), for 0<xv<rK3 are developed below:

33

313

21cosK

v

K

vKK d

xr

xr ⋅−=

−=−θ (5)

( )

−⋅

⋅=

−−=−

333

23

23

13 14sinK

v

K

v

K

vKKK d

xd

xr

xrrθ (6)

The functions sine and cosine from the expression (4), for rK3<xv<2rK3 are developed below:

12

cos3

23 −⋅

=−K

vK d

xθ (7)

−⋅

−−⋅=

⋅

−⋅⋅

=−3333

23 11122

12

sinK

v

K

v

K

v

K

vK d

xdx

rx

rx

θ (8)

Where: dK3 – the diameter of the orifice OK3. To bring the area function AK3 into a simpler and compact form, it can be used the steps from [6], where the fluid passing area through a circular plan surface is theoretically computed. In this manner, the function sine, cosine and arcos are developed using binomial and Maclaurin (Taylor) theorems from [7].

46


By expressing the functions (5), (6), (7), (8) as Binomial and Taylor series, the obtained expressions depend on the number of binomial and maclaurin terms selected (usually the

first three or four terms, omitting the higher powers). For a flow cross area variation from zero to full opening, the calculus area expression is:

( )

⋅−⋅

⋅⋅=

3

23

3

233 10

3134

K

v

K

vKK d

xdx

dxvA

(9)

For a flow cross area variation from the full opening to a complete obstruction, the calculus area expression is:

( )

−⋅−⋅

−⋅−⋅=

3

23

3

233 1

8311

34

4 K

v

K

vKvK d

xdx

dxA π (10)

By applying binomial and Maclaurin theorem approximations are made. To obtain a better accuracy the error of the approximations can be adjusted by modifying the number of series coefficients.

In figure 9, the resulting cross area variation from the closed position to the full opening of a resistance is illustrated. The diameter is set of 80 mm.

3.2 The null operating point of the rotatable distributor

In figure 10, it can be observed that for a null command displacement (xv), this is when the spool is symmetrical positioned in the distributor body, the flow sent to the motor is zero. Thus, the null operating point of the rotatable distributor is characterized by:

( ) 0, 22 == vCvC xpxQ (11) Based on the figure 10 the statement written referring to the expression (11) is valid for any control pressure, the following condition must be fulfilled:

SPSFC pp ⋅=− α2 (12) The condition (12) has to be disposed because in a steady state analyze the pressures in the linear hydraulic motor must set the piston in equilibrium.

Fig.9. The cross area variation in the rotating hydraulic distributor, obtained with binomial theorems,

from zero to a full opening.

Fig.10. The pressure-flow curves for three different command displacements of the rotatable spool: xv/xvmax =-1; xv/xvmax = 0;

xv/xvmax =1.

47


3.3 The pressure-flow curves of the rotatable distributor

In figure 11 it is illustrated the pressure-flow curves for different command displacements of the rotary spool of the hydraulic rotatable distributor DHR which is used to command and control the linear hydraulic motor MHR (figure 3). The pressure-flow curves of a linear three way control valve are presented in figure 12 and are identical with those of the rotary valve.

Similar pressure-flow curve plots are revealed and analyzed in many research labors as [8], [9], [10].

3.4 The rotatable distributor coefficients and the linearized flow equation

Usually, the nonlinear algebraic equation (1) which describes the pressure-flow curves illustrated in figure 10 can be seen as a Taylor series and expressed regarding to the valve null operating point. The linear equation which is obtained by developing the equation (1) as a Taylor polynomial of degree 1 is easier to analyze. Similarly with the steps used in [8], the partial derivatives which appear in the Taylor polynomial are obtained by differentiation or graphically from the pressure-flow curve plots (figure 9 or 10). The rotatable distributor coefficients are defined by these partial derivatives and the computed values are:

0=−xvQK (13) 0=−pQK (14)

∞=pK (15) The flow gain or the partial derivative of the load flow with respect to rotary spool displacement, as is defined a valve flow-displacement coefficient, is the most important parameter for a valve [5-9] but in the same time questionable [11]. The valve pressure drop is always changing yielding to a variable flow gain [11]. As the pressure drop on rotatable distributor changes, the system flow gain will not be constant. Flow-displacement coefficient of a valve is proportional with system flow gain. Based on the results (13) and (14), the general linearized flow equation is:

( ) 0, 22 =∆ CvC pxQ (16)

Fig.11. The pressure – flow curves of the rotatable hydraulic distributor for different spool command

displacements.

Fig.12. The pressure – flow curves of a linear

hydraulic distributor for different spool command displacements.

48


The flow differentiation is determined for an infinitesimal variation, in the vicinity of the null operating point. In consequence, a zero result (16) will show that for infinitesimal variation of the rotating spool displacement and for infinitesimal variation of the control pressure in steady state regime, theoretically, there will not be changes in the load flow. 4. Mathematical modeling of the percussion mechanism In figure 13 it is created a hydraulic scheme for the percussion mechanism where the parameters used in the mathematical modeling of the impact system are presented. 4.1 The continuity equations for steady one-dimensional flow The continuity equations of the liquid fluid in the linear motor chambers are:

( )2111110

1 CCipCepCCC

C ppCpCt

Vt

pE

VQ −⋅+⋅+∆

∆+

∆∆⋅= (17)

( )122220

2 CCipCCC

C ppCt

Vt

pE

VQ −⋅+∆

∆+

∆∆⋅= (18)

By summing the continuity equations (17) and (18), and by using the notations (20), (21), (22) and (23) it gives:

( )

⋅+⋅⋅++⋅⋅

⋅⋅= LoadpLoadpLoadpLoad

pLoad pCxApC

ExA

Q 2_1_ 121

α (19)

Where:

212 CC

LoadQQ

Q+

=

(20)

12 CCLoad ppp −= (21)

121_ CCLoadpLoad pppC ⋅+=⋅ α (22)

12_ CepLoadpLoad pCpC ⋅=⋅ (23) Where: E – the liquid bulk modulus of elasticity; mp – the total inertial mass reduced to impact piston mass; pLoad – the pressure drop on the hydraulic linear motor; QLoad – the average volumetric flow rate supplied to the motor chambers; xpmax – the maximum stroke of the impact piston.

Fig.13. The symbolical hydraulic scheme of the

subsystem, where: A, αA – the piston areas with α the ratio between them; Cep is the external leakage coefficient; Cip is the internal leakage coefficient; pC1 and pC2 are the pressures in the supply, respectively

the discharge lines; xp is the impact piston displacement.

49


4.2 The equation of movement The equation of impact piston movement is:

21 pCpCppp FFxxm =+⋅+⋅ µ (24) Where: μ – the friction coefficient; FpC1, FpC2 – the pressure forces acting on the impact piston active surfaces. The following notation is made:

123_ CCLoadpLoad pppC ⋅−=⋅ α (25)

Using the notation (25), the equation (24) can be rewritten as:

ppLoad

ppLoad

pLoad x

CAx

CAm

p ⋅⋅

+⋅⋅

=3_3_

µ (26)

4.3 The system characteristic equation The system characteristic equation is:

( ) ( )

( )

+⋅+

⋅

⋅⋅+⋅

⋅

⋅⋅+

+

⋅⋅

⋅⋅⋅

+⋅⋅

⋅⋅⋅

⋅=⋅

αµ

µ

1

2

3_

2_

3_

2_

3_

1_max

2

3_

1_max

ACACC

sCA

mCC

sCA

CE

xA

sCA

mCE

xA

sXsQs

pLoad

pLoadep

pLoad

ppLoadep

pLoad

pLoadp

pLoad

ppLoadp

pLoad(27)

The system characteristic equation gives a natural pulsation ωn in the presence of damping and leakages defined as:

( )( )ppLoadp

pLoadeppLoadmCxACCCAE

n ⋅⋅⋅

⋅⋅+⋅+⋅⋅=

1_max

2_3_2 1 µα

ω (28)

And a damping ratio δ defined as:

( )( )µα

µδ⋅⋅+⋅+⋅⋅

⋅⋅⋅⋅

⋅

⋅

⋅+⋅=

2_3_2

1_max

1_

2_

max 121

pLoadeppLoad

ppLoadp

pLoad

pLoad

p

ep

p CCCAE

mCxACC

xACE

m

(29) These two characteristic parameters describe the dynamic behavior of the analyzed subassembly through the expressions (28) and (29). 5. Discussions and conclusions An innovating command structure for a hydraulic impact mechanism used in industrial applications such as rock, concrete or asphalt penetration is presented, described and a mathematical modeling is developed. The command structure is essentially a rotary spool with circular grooves and mounted in a piston prominence. Cylindrical orifices are radial drilled in the piston prominence to allow the fluid flow from the rotary spool grooves into a working motor chamber. The rotary movement of the hydraulic rotary spool and the hydraulic feeding of the motor assure a linear oscillatory movement for the impact piston. Two methods of calculus the flow cross area through the hydraulic resistances of the hydraulic rotating distributor are revealed.

50


The pressure – flow curves of the hydraulic rotating distributor are computed. The hydro mechanical impact mechanism is assumed to be a dynamic system with constant parameters unlike the real situation according to which the coefficients are determined by the impact mechanism position. Moreover, the liquid bulk modulus is assumed to be constant in the present mathematical modeling. The obtained linearized flow equation of the hydraulic rotating distributor equals zero. A mathematical interpretation for this result is that for small deviation of the rotary spool from its steady symmetrical position in the piston prominence, the hydraulic linear motor – hydraulic rotary distributor subassembly behaves as a hydraulic linear motor.

6. Reference [1] L. Vaida, C, Kozma, Generator hidraulic de vibraţii pentru perforatoare hidraulice rotopercutante,

registered to OSIM: A/10018/2012 [2] C. Kozma, L. Vaida, A constructional and functional improvement in hydraulic rotary percussive drill,

HERVEX ISSN 1454 – 8003 (2011) [3] C. Kozma, Mathematical modelling of a hydraulic vibro percussive system, SIDOC Project – Doctoral

students’ session (2012) [4] M. Ivantysynova, Design and Modeling of Fluid Power Systems, Purdue University (2012) [5] L. Deacu, D. Banabic, M. M. Rădulescu, C. Raţiu, Tehnica hidraulicii proportionale, Ed. Dacia ISBN

973-35-0058-5 (1989) [6] R. B. Walters, Hydraulic and electro-hydraulic systems, Publisher: Springer, ISBN-13: 978-

1851665563 (1991) [7] V S. Gourley, Binomial expansion, power series, limits, approximations, Fourier series, University of

Surrey (2007) [8] C. Kozma, The static and dynamic analysis of a hydraulic 3/2 valve with linear displacement.

Pressure-flow curves, HIDRAULICA ISSN 1453 – 7303 (2012) [9] H. E. Merrit, Hydraulic control systems, Publisher: John Wiley & Sons Ltd, ISBN 0-471-59617-5

(1976) [10] K. E. Rydberg, Hydraulic servo systems – Course, TMHP51, IEI, Linköpings Universitet, 2008 [11] J. L. Johnson, How to interpret valve specifications, http://hydraulicspneumatics.com (2005) [12] C. Kozma, L. Vaida, Hydraulic schemes for impact devices. A control system for impact mechanisms

using a rotatable distribution valve – part2, HERVEX ISSN 1454 – 8003 (2012) [13] Banyai D.V., Metode noi în sinteza maşinilor hidraulice, cu volum unitar variabil şi reglare electro-

hidraulică, thesis (2011) Address for correspondence 1. Ing. Claudia Kozma, PhD.: Technical University of Cluj-Napoca, Faculty of Mechanical Engineering, Department of Thermal Engineering, 3400 Cluj-Napoca, Romania Tel: ++40 264 401777 Fax: ++40 264 401777 e-mail: [email protected]

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ISSN 1453 - 7303 “HIDRAULICA” (No. 1/2013) Magazine of Hydraulics, Pneumatics, Tribology, Ecology, Sensorics, Mechatronics

THE ANALYSIS OF FLOW LOSSES THROUGH DYNAMIC SEALS OF HYDRAULIC CYLINDERS

Gabriela MATACHE1, Stefan ALEXANDRESCU1, Adrian PANTIRU1, Gheorghe SOVAIALA1, Mihai

PETRACHE2

1INOE 2000– IHP Bucharest, e-mail:[email protected] 2SC.LYRA SRL Bucharest, e-mail: [email protected]

Abstract: To hydraulic actuators most commonly known as cylinders a feature seen in terms of performance is the product availability depending on dynamic seals. This paper intends to clarify the issue of losses through mobile sealing elements of parts in relative linear motion piston – cylinder liner and rod – piston guide. In the second part of the paper there is presented a comparison of these losses to those of other hydraulic equipment. Keywords: hydraulic cylinders, piston seals, flow losses.

1. Introduction

As it is known, hydraulic cylinders also known as actuators or linear hydraulic motors convert the pressure energy of oil supplied by a pump into mechanical energy capable of providing an active energy of translational motion and thus mechanical work, [1]. From a constructive point of view there is a variety of such hydraulic devices, in figure 1 being shown a double acting and unilateral rod cylinder mounting with front flange.

Figure1. The components of a hydraulic cylinder 1 - cylinder liner, 2- piston, 3t - piston socket mounted near the rod, 3p - piston socket mounted

near the piston end, 4 - cylinder rod, 5 - fixed rod seal, 6 - piston guide ring, 7- nut, 8 - washer,10- O ring, fixed lid seal, 11- lid, 12 - rod guide ring, 13 - tie bar, 14 - nut, 15 - washer, 16 - scraper, 17-

rod socket, 18 - mounting lid.

52


When using hydraulic motors, components that wear out are sockets 3p, 3t, 17, guide rings 6,12, rod 4 and scraper 16 as a result of friction towards the fixed parts. Elements that first wear are the sockets and the negative effect materializes through loss of flow inside and outside of the product. Size of these losses determines the acceptability criteria for mobile sealing gaskets and thus their life span. The variety of sealing elements for rod and piston correlated to the criteria of the cylinder working speed, maximum working pressure and dimensions determined the manufacturers to avoid stipulating in booklets the acceptable flow loss amounts over the lifetime of these products and thus the time for their replacing.

2. Sealing process

As it is known, until now, there has not been developed a general theory of sealing explaining all aspects of this process. Seals are machine parts assemblies aimed at closing as tight as possible a space containing a hydraulic environment under pressure. Sealing system, in the case of hydraulic cylinders, must maintain pressure and avoid loss of fluid to the low pressure side, both inside the cylinders, by isolating chambers Cp and Ct,- figure1- and outside the rod during its movement outwards. In the case of hydraulic cylinders sealing is achieved due to pressing when mounting against the contact surface and also due to the radial force resulting from the pressure of the sealed environment itself. This way to achieve the pressing force from within the hydraulic system is called the “self sealing”.

In Figures 2 and 3 are shown elements with self sealing (O ring and socket), [2]

Figure 2 Mobile sealing of the piston with O rings:

1- O ring, 2- guide ring, 3- anti-extrusion ring

Figure 3 Mobile sealing of the piston with sockets 2-guide ring, 3- anti-extrusion ring, 5- sealing socket.

By installing the seal into its place with an elastic deformation by compression, there is created locally an initial sealing pressing. When there occurs pressure of the sealed environment automatically increases the force by which the cylinder socket presses against the cylinder liner. As one can see, the initial pressing, when assembling, has less weight compared to pressing resulting from hydraulic environment pressure. Uniformly distributed radial force [3] for sealing the piston is calculated using the formula:

53


F = πDpl

D – piston diameter;

p – pressure within the piston;

l – sealing length

Figure 4 1-initial pressing, at mounting, of the socket in the

absence of pressure

Figure 5 1- final pressing of the socket in the

presence of static pressure

Fluid losses during displacement of rod and piston are in fact leakages through the clearance existing between the sealing and surrounding parts. In the specialty literature they are also known as losses in flow of the cylinder as this amount of oil reaches the tank circuit.

Flow loss consists of the amount of oil which during the input stroke of the rod in the cylinder no longer returns to the area under pressure. Size of these losses depends on several parameters: oil viscosity, roughness, deviations in shape of the part that the socket is in contact with, materials, state of wear of the seal, constructive solution, size etc. The sealing through contact process is related to the presence of a clearance needing to be closed, characterized by the pressure gradient dp/dl, parameter which determines the pressure p curve shape along the length l of the socket.

54


At piston displacement the socket is deformed during the hydrodynamic process of occurrence of a clearance. During displacement the initial pressing from the seal mounting does not change. Socket deformation is minor compared with the average size of the clearance, the graphic of hydrodynamic pressure variation within the clearance along the sealing length of the piston socket looking, hypothetically, as shown in Figure 6, [4].

When switching from non-operative mode to reciprocating motion mode in the presence of viscous oil the two areas are completely separate, hydraulic film whose thickness changes will result in increased oil loss. Fluid film thickness on the work surface should be as small, but sufficient to produce lubrication of seal.

Figure 6 Variation of hydrodynamic pressure within the sealing

clearance of the socket

3. Quantifying flow losses (leakage) at hydraulic cylinders

Measurement of fluid loss in hydraulic cylinders is performed to assess the functional status of the seal after installation but also after certain periods of working, at endurance test for determining the lifetime of the product. This check is also performed periodically within the program of hydraulic system maintenance and compulsorily after repairing of cylinders.

3.1. Quantifying flow losses at rod socket

In accordance to the provisions of STAS 8535-83, measurements are performed at a pressure of 1.25...1.5 rP (rP – rated pressure) after performing five double strokes - the testing diagram is shown in Figure 7.

55


The safety valve sV is adjusted at the pressure of 1.25 ... 1.5 rP, and then there are performed a minimum of five double strokes by switching on and switching off the electromagnet of distributor ED. Way throttle valves WT are adjusted so as to achieve the load opposite to cylinder displacement. This operation aims to achieve a maximum thickness of oil film on the surfaces moving relative to the stationary ones (piston - liner, rod – guide).

Measurement of losses at the rod is performed during the external tightness check on a certain amount of time, permissible values being specified in manufacturer standards and very rarely in supplier catalogues. Measurements are usually performed at an oil viscosity of 33±3 mm2/sec, corresponding to a temperature of 50±5ºC.

Measurement diagram of flow losses to the rod is shown in Figure 8.

After pulling back the piston in position in Figure 8, pressure adjusted by the stand valve is maintained constant over a certain period of time 10...15 min.

Figure 7 HC – hydraulic cylinder to be tested, HP- hydraulic

pump, sV- safety valve, M-manometer, ED- electrically actuated distribuitor, WT – way throttel

Losses of fluid through the piston sealing area reach the tank via distributor DE, while losses through the rod sealing clearance, which are of interest for these checks will be collected in calibrated vessel cV. Regarding fluid loss amounts there are diverse opinions. According to some authors, on the rod is allowed forming of a film of oil without forming droplets. Another opinion is that the volume of the film which is formed should not exceed a volume of 0.5 cm3 to 50 double strokes, a double stroke having a length of 0.5m at the maximum working pressure. There is promulgated an even more tolerant amount of fluid loss at a 30mm diameter rod: up to 5 drops / min (0..25 cm³/min) at the pressure of 110 bar. This value is too large and should not be taken into account.

Figure 8 HC – hydraulic cylinder in drawn back position, ED-

electrically actuated distribuitor, cV - calibrated vessel

56


On the other hand, the notion of admissible fluid loss is still insufficiently clarified. For example, in the specialized literature they speak of zero losses, although these losses can mean a few cm ³/h, without the occurrence of drops. An acceptable amount of flow losses to rod, measured as illustrated in Figure 8, is up to 6.5 mg/min to seal in new condition. During operation of the cylinder losses get larger and time when the product must be replaced is chosen by the user.

In conclusion, rod seal of a cylinder is chosen depending on the purpose, requirements and safety that it must comply with inside the hydraulic system, e.g. at the cylinder rod from aircraft outside losses are not allowed. An indicator of a seal performance is the number of kilometers covered with a certain pressure not exceeding the amount of loss provided in manufacturer standards.

3.2. Quantifying flow losses at piston sockets

Internal flow losses through the piston sealed clearance are an important parameter which is measured during type tests, on a regular basis but also during predictive maintenance program. Checking, according to the standard in force, is performed in five fixed, intermediate, equally spaced positions of the cylinder stroke at a pressure of 1.25...1.5 rP after performing minimum five strokes on a stand, Figure 7. Measurement of internal flow losses is performed according to the diagrams in Figures 8 and 9.

The amount of liquid that is collected in the calibrated vessel cV relates to a minute, but for more accurate measurement duration should be about 10 ... 15 min. Usually this check is performed at both ends of the stroke - Figures 9 and 10.

Figure 10

Figure 9 HC – hydraulic cylinder in advanced head stroke position, ED- electrically actuated distribuitor, cV - calibrated vessel, Man-

manometer

Measurement of internal losses in the other three fixed points of the rod complicates the stand structure, for this reason the check is performed only at the end of the stroke. Currently, most manufacturers of cylinders buy pipe for cylinder liner from specialized manufacturers, so that there are guaranteed deviations from cylindricity, roundness, straightness and material. During the experiments, due to the high safety coefficient, 3.5, no residual deformities were observed as convex or concave shape. Moreover, in calculation formula of cylinder liner thickness the liner length is not taken into consideration and thus checking in intermediate positions is not justified.

57


Measurement of internal losses in several points is performed, usually, at cylinders on mining machinery which have undergone hydraulic shocks and whose liner was deformed as a result of massive rock fall.

In terms of internal losses, in this area also technical information are few and elusive, not being indicated the sampling parameters: pressure, socket diameter, material, oil temperature, etc.

In figure 11 is shown a diagram of the internal losses depending on piston diameter at hydraulic cylinders of 160 bar used in machinery – tools, [5]. Permissible values are higher compared to those of rods, their growth is not visible outside the hydraulic system, and the negative effect is quantified by lowering rod speed compared to initial adjustment. Mean value recommended by the manufacturers of hydraulic cylinders, measured under the above conditions, is 6 mg/min in the case of a socket in new condition.

It can be seen that the losses increase proportionally with socket circumference and they have the lower limit value 1 cm ³/min which is still a high value. Flow losses diminish the more dynamic viscosity is higher, it being inversely proportional to them. Theoretically, the variation of viscosity is dependent on oil temperature and actually it materializes in the amount of leakage. Determination of fluid loss is based primarily on the results of experimental tests, even more as the friction regime of sealing surfaces is most commonly semi-fluid and exceptionally fluid.

Figure 11

Variation of flow losses depending on the outside diameter of sockets

to a pressure of approximately 160 bar

4. Internal losses at other hydraulic devices

Permissible values of fluid losses at hydraulic cylinders are much smaller than at other devices. For instance, in Figure 12 are shown permissible losses to hydraulic slide valve manifolds at a pressure of 160 bar. Values are high because their operation is conditioned by the existence of radial clearances, unlike mobile sealing elements where there is a pressing force at mounting which increases in the presence of pressure inside the cylinder. Variation of losses on the levels of dimensional parameters rD6, rD10, rD20, rD32 depends on hydraulic diagrams of manifolds and the overlap between the slide valve and the body. For a cylinder that has a piston diameter of 90 mm and moves at 0.020 m / s is required a flow rate of 800 l / min, corresponding to a manifold rD 32. Comparing the graphs in Figures 11 and 12 it can be noted that permissible losses at the mobile piston seal are of 6 cm³/min, while at the manifold they are of 600 cm3/min – that is 100 times higher.

58


Figure 12 Flow losses at 160 bar at slide valve manifolds

At axial piston pumps permissible flow losses are caused by leakage through diametrical and frontal clearances. In Figure 13 is presented the variation of flow losses related to flow. These values increase during operation, the user choosing the moment when he/she considers the pump should be replaced [6].

Figure 13 Flow losses at 160 bar at axial piston pumps

59


5. Conclusions Evaluation of flow losses in hydraulic equipment as accurate as possible is required for efficient sizing of hydraulic electric pumps. Pump flow that is taken into account is the useful flow of hydraulic system to which are added flow losses of the equipment, and in this sense, we can calculate the volumetric efficiency of the hydraulic system. Also, increasing over time of oil losses due to wear is an important parameter in determining the lifetime of hydraulic equipment, of sealing elements and in establishing the maintenance program.

REFERENCES

[1] C. Cristescu, P.Drumea ,D I.Guta, C.Dumitrescu, P.Krevey ‚’’Theoretical and experimental research regarding the dynamic behaviour of linear hydraulic motors’’, Journal : Hidraulica no. 1 - 2 / 2011, ISSN 1453 – 7303 [2] Assofluid, ’’Hydraulics in industrial and mobile application’’, Milano, September, 2007 [3] A.Fatu, et al., “Evaluation of the elastomer hyper elastic behavior a U-cup hydraulic rod seal” Hidraulica

no.3, October 2010, ISSN 1453-7303 [4] C. Cristescu, P.Drumea, “Mathematical modeling and numerical simulation of the tribologic behaviour of mobile translation sealings subjected at high pressures”, Hidraulica no.2 (22) September 2008, ISSN 1453-7303 [5] A.Oprean, “Hidraulica masinilor unelte”, Didactic and Pedagogical Publishing House, Bucharest,1965 [6] T.C. Popescu, I. Lepadatu, D.D. Ion Guta, Experimental research activies regarding the reduction of

energy compsuntion at endurance test stands of rotary voliumetric machines, Proceedings of International Scientific Technical Conference Hydraulics and Pneumatics 2009, Wroclaw, 7-9 October, ISBN 978-83-87982-34-8

60


DOUBLY FEED INDUCTION GENERATOR FOR BIOMASS COMBINED HEAT AND POWER SYSTEMS

Curac IOAN1, Craciun BOGDAN IONUT2 Banyai DANIEL VASILE 3 1 Technical University Of Cluj Napoca, Department of Electrical Machines and Drives, [email protected] 2 Aalborg University Department of Energy Technology, [email protected] 3 Technical University Of Cluj Napoca, Department of Mechanical Engineering,

[email protected]

Abstract: Due to concerns regarding environmental problems this paper present the potential of biomass CHP (Combined Heat and Power) systems to improve ancillary services in distributed generation systems. Thermal power SG (synchronous generator), advantages and disadvantages comparison with DFIG (Doubly Feed Induction Generator) are also shortly described. Keywords: biomass, combined heat and power, distributed generation, ancillary services, grid codes, DFIG 1. Introduction

A lot of literature in Power Systems study cogeneration biomass power plant grid couplet with SG and them grid his control. That why the aim of this paper is to propose to open a detailed study of DFIG in state of SG for biomass power plant with steam turbine and capacity between 1-5 MWe . This capacity is suitable with DG from rural area and for sawmills. International Council on Large Electricity Systems (CIGRE) defines DG unit as a generation unit that is not centrally planned, not centrally dispatched, usually connected to the distribution network and smaller than 50-100 MW [3], [4]. There are a wide variety of potential benefits to distributed energy systems both to the consumer and the electrical supplier that allow for both greater electrical flexibility and energy security [1], [2]. A system with appropriate levels of security and power quality is not necessarily to run in an optimal manner. For example, reactive power injected at terminals of a transmission line can increase the active power transit capacity. Similarly, if some transmission capacities are reserved to allow the supply of ancillary services, less power for energy can be transmitted. Moreover, the provision of ancillary services plays a role in amount of losses and impacts the aging of infrastructures. More generally, the consequences on the various resources of the power system have to be taken into account while using ancillary services in order to use the resources of the system in an optimal manner.[5]. Ancillary services are defined as services provided in addition to real power generation. They include, amongst others, reactive power control, provision of spinning reserve, frequency control, and power quality improvement.[6]. 2. BIOMASS COMBINED HEAT AND POWER (CHP) PLANTS

2.1 Biomass cogeneration Waste wood biomass conversion uses basic two categories of technologies, one is thermochemical that use high temperatures to convert feedstock to energy However, the technologies have potential to produce electricity, heat, bio products, and fuels. The other

61


technologies are biochemical and use biological agents to convert biomass feedstock to clean energy. In addition, this technology has the potential to produce electricity, heat, bio products, and fuels [7]. Biomass combustion is the main technology route for bioenergy, responsible for over 90 percent of the global contribution to bioenergy. The selection and design of any biomass combustion system is mainly determined by the characteristics of the fuel to be used, local environmental legislation, the costs and performance of the equipment necessary or available as well as the energy and capacity needed (heat, electricity)[9]. For example biomass cogeneration systems used in sawmills are indirect fired and compound from biomass boiler condensing steam turbine with extraction and synchronous generator as shown in Fig. 1.

Fig. 1. Combined Heat and Power Diagram – Indirect Fired [8]

Biomass used is a result from wood industrialization has humidity between 50-60% and is burned in grate-fired or circulating boilers. Steam produced is expanding in turbine witch produced mechanical power for synchronous generator, then pass in condenser and boiler thru feed water pump. Low-pressure steam produced is used for drying chambers, space heating or cooling systems.

62


Fig. 2. Power-to-heat ratio function of the plant size of biomass-fuelled CHP plants in Finland and

Sweden with 1–20MW [9].

2.2 Steam turbine Steam turbines are the most common technology used in power plants and industries. Depending upon the exit pressure of the steam, steam turbines fall into two types: backpressure turbines and condensing turbines. Backpressure turbines operate with an exit pressure at least equal to atmospheric pressure, and are suitable for some sites with a steam demand of intermediate pressure. Condensing turbines have the advantage of changing electrical and thermal power independently and they work with an exit pressure lower than atmospheric pressure [10]. Governing systems for steam turbine is containing three basic functions: normal speed load control, over speed control, and over speed trip. In addition, the turbine controls include a number of other functions such as start-up/shutdown controls an auxiliary pressure control [12] 2.3 Synchronous generator In grid-tied operation the voltage characteristic is given by the mains grid. The SG has to be synchronized to the grid’s voltage with regard to its voltage magnitude, frequency, phase sequence, and phase shift by use of the above-described control capabilities. The rotor is then forced by the stator field to rotate with the network frequency [11] 2.4 Doubly Fed Induction Generator The stator of a doubly fed induction generator (DFIG) is connected to the grid directly, while the rotor of the generator is connected to the grid by electronic converters through slip rings, as shown in Fig. 3.The generator can deliver energy to the grid at both supersynchronous and subsynchronous speeds. Thing that can help biomass cogeneration power plants to compensate different humidity and low quality fuel[13].

63


Fig. 3 Exemplary of DFIG grid connection

3. Conclusions Since Power Electronics develop powerful models of inverters for wind turbines new possibilities and opportunities in research of cogeneration power plant for a large integration in DG has been opened. REFERENCES

[1] Benjamin Kroposki, Senior Member, IEEE, Christopher Pink, Member, IEEE, Richard DeBlasio, Senior Member, IEEE, Holly Thomas, Member, IEEE, Marcelo Simoes, Senior Member, IEEE, and P.K. Sen, Senior Member, IEEEE. , Benefits of Power Electronic Interfaces for Distributed Energy Systems., Power Engineering Society General Meeting, 2006. IEEE

[2] G. Pepermans, J. Driesen, D. Haeseldonckx, R. Belmans, W. D’haeseleer, “Distributed generation: definition, benefits and issues”, Energy Policy, vol. 33, pp. 787-798, 2005.

[3] [3] Angelo L'Abbate, Gianluca Fulli, Fred Starr, Stathis D. Peteves, Distributed Power Generation in Europe:technical issues for further integration, EUR 23234 EN – 2007

[4] Conseil International des Grands Réseaux Electrique(CIGRE) WG 37-23, Impact of increasing contribution of dispersed generation on the power system, Final Report, 2003

[5] Yann Rebours, "A Comprehensive Assessment of Markets for Frequency and Voltage Control Ancillary Services," Ph.D. thesis, Faculty of Engineering and Physical Sciences., University of Manchester, 2008

[6] Johannes MORREN, "Grid support by power electronic converters of Distributed Generation units", Ph.D. thesis, Technical University of Delft 2006

[7] Biomass Conversion Emerging technologies Feedstock and Products, EPA/600/R-07/144 [8] Curac IOAN, Craciun Bogdan Ionut Creta Ioan, State Of The Art Biomass Combined Heat And Power

Technology, Proceedings of 2012 International Conference of Hydraulics and Pneumatics, HERVEX, 7-9 November, Calimanesti-Caciulata, Romania, pp. 411-417

[9] Sjaak van Loo and Jaap Koppejan, The Handbook of Biomass Combustion and Co-firing, UK and USA in 2008.

[10] D.W. Wu, R.Z. Wang, Combined cooling, heating and power: A review, Progress in Energy and Combustion Science 32 (2006) 459–495

[11] Martin Braun, Provision of Ancillary Services by Distributed Generators, Ph.D. thesis, Kassel university, 2009

[12] P . KUNDUR, Power Systems Contol and Stability,United States 1994

64


A NEW MODEL OF PNEUMATIC TRANSDUCER USED IN THE DRYING STAGE OF THE CERAMIC PRODUCTS OBTAINING

Murad Erol1, Dumitrescu Catalin2, Haraga Georgeta1, Dumitrescu Liliana2

1. Universitatea POLITEHNICA Bucuresti 2. INOE 2000 -IHP Abstract For obtaining high-quality ceramic products, the final stage of the production process, the drying, must be realised in good conditions, leading to very low final losses caused by the cracks and cleaves which occur during the burning in the oven. An important parameter of the drying process is the dimensional contraction of the bodies, during drying. For the online control of the drying process and depending on the contraction evolution, it was created a pneumatic transducer, with which can be measured the linear contraction, on a witness body, with a max error of 1%. The transducer performs a conversion linear contraction → pressure, due to the fact that the air presure is not influenced by temperature and the elastic element of the pneumatic cylinder is made from steel, with a slight variation of the elastic properties up to 150oC. The measurement pressure is linearly converted in a tension which is input signal in the PLC which controlls the drying installation. In order to obtain a very low consumption of energy of the measurement system, the transducer works in a sampled operational mode, which may provide a high measurement precision.

Key words: drying, pneumatic transducer, contraction, ceramics

1. Introduction Drying is a very important operation within the process of manufacture of ceramic products and construction materials. A high quality drying leads to very low final losses caused by the cracks and cleaves which occur during the burning in the oven, as last stage. The drying time for the average size bricks are relatively high, 36 - 48 h. In all the drying installations it is used a drying agent obtained by mixing combustion gases and atmospheric air. The famous brands (Lingle, Ceric, Rietter, Keller, Fuschs, etc.) deliver technological installations with automatic control systems, which use process models pre-selected depending on the type of brick to be dried and the chemical, mineralogic and physic ceramic properties of the raw clayey materials. For optimizing the drying process, it must be compensated entirely the influence of the perturbations caused by the composition and granulometry of the bodies subjected to the drying process. The online measurement of the discharged water mass and of the linear contraction of the bodies allows the achievement of high quality products, with minimum energetic consumption and an increase of the real manufacture capacity. In figure 1 is shown the graphic of the relative variation of humidity W and contraction C of the bricks, during the the drying process.

65


Fig. 1. The relative variation of humidity and contraction at drying of bricks

The relative linear contraction of the ceramic bodies, in the studied case of the bricks, during drying is in the range 6...12% with typical values 6...8%. The measurement of the contraction is made by measuring the variation of the distance between the two anchors introduced in the raw brick before starting the drying process. The common values of the initial distance L0 for the mechanical measurement devices are L0 ∈{100, 150, 200} mm. During the drying process the body contracts itself and it is measured a closeness of the anchors with ∆L. The relative contraction coefficient ε is calculated with the relation:

(% )1 0 00LL∆

⋅=ε (1)

Due to the fact that during the drying process temperature varies between 20 and 150 ºC, the anchors and the support mechanism are made of invar. The mechanical devices in use is ponderous and has no output signal for the modern automatic adjustment systems. There are also used resistive displacement transducers with temperature compensation, which are expensive and less reliable.[11]

2. Pneumatic transducer for measuring contraction

In figure 2 is shown the functional scheme of the pneumatic transducer for measuring the linear contraction of the bricks during the drying process and in figure 3, the functional scheme of the control block.

Fig. 2 Functional scheme of the pneumatic transducer for measuring contraction

66


The absolute contraction is represented by the variation of the distance ∆L between the anchor (1) fixed on the body (2) of the transducer and anchor (3) which is fixed on the rod (4) that slides in the guide (5) on which is the pawl (6). The rod (4) is connected to the support (7) on which is jointed a lever (9) which is pushed continuously by the spring (11) to the fix contact (8) which is at 2s distance of the fix contact (10). The lever (9) is also in contact with the rod (12) of the pneumatic cylinder with simple action (13) that has a membrane (16) which presses on the rigid centre (15) that leans on the spiral spring (14). For having the center of gravity of the device between the 2 anchors, it is mounted a load (17) for balance. For simplifying the construction it was adopted a scheme with 2 contacts K1 and K2 serially linked to R1, with signal Imas in a current intensity on a single conductor connected by means of R2 at the plus of the supply voltage stabilized Ust. The command signal in tension Ucd is compared on 2 comparators C1 and C2 with two tensions 0,25 Ust at C1 and 0,75Ust at C2. The output signals of the comparators CS for writing and CR for deleting are applied to a bistable BIST1 whose output u1 commands the distributor D1 through which is introduced compressed air in the cylinder (13) through the pneumatic resistance RP. For discharging the air from the cylinder (13) it is commanded through the output command signals CRp and CSp the bistable BIST2 with which is commanded the distributor D2. The measurement pressure pmas is converted in tension signal with a convertor P/U from which gets out the signal Umas.

Fig. 3 The scheme of the command block

Initially in the raw brick are plunged the anchors (1) and (3) positioned at the nominal distance L0. If in the pneumatic cylinder (13) there is no pressure, the spring (14) retracts the rod (12) of the piston and the spring (11) presses on the comparison lever (9) that will rely on the contact K1. When the contact K1 is shut in the command block is generated a signal u1=1 of opening the distributor D1; it starts the pressure growth pmas, the rigid centre (15) compresses the spring (14) and displaces the rod (12) until it reaches the lever (9) that it will rotate around the joint until it leans on the contact K2. The command block generates a shut signal u1 = 0 for D1 and pmas stabilizes at a new value pmas0 that is the value from which is started the measurement. From the converter p/U results Umas0 memorized as origin for measuring contraction. In the drying process cause of the brick contraction the anchor (3) and rod (4) will displace towards the anchor (1), the spring (11) releases and under the action of the spring (11) the comparison lever (9) rotates itself until the contact K1 shuts. It is generated the signal CS which applied at BIST1 makes that u1 = 1 and D1 it opens. It starts the pressure pmas growth from the cylinder which leads to the compression of the spring (14) and the displacement of the rod (12) that rotates the

67


lever (9) until the contact K2 shuts, which leads to the generation of a signal CR which applied at BIST1 makes u1= 0 and shuts D1. For a displacement with 2s between the contacts K1 and K2 the rod (12) displaces with Δharc:

⋅⋅

=+

⋅⋅=∆p g

a rc is

dddsh 22

21

2 (2)

where ipg is the transfer factor of the lever 9. The variation Δpmas of the measurement pressure for a sampling step is:

m e f

a rc

p gm e f

a rca rcm a s S

Ki

sS

Khip ⋅⋅

=∆

=∆2][ (3)

Because the value of variation ΔL[i] of the absolute contraction for sampling is:

.2][ c o n s ti

siLp g

=⋅

=∆ (4)

It results that:

.][ c o n s tSKL

SKhip

m e f

a rc

m e f

a rca rcm a s =⋅∆=

∆=∆ (5)

If it adopts constructively the value L0 = 100 mm and the max. contraction is εmax ≤ 10%, it results that the max. value of the displacement to be measured is ΔLmax = 10 mm. For being efficient in controlling the drying process it is required that the device to belong to the CP1 class of precision, meaning that the sampling error ΔL[i] is: .1.00 1.01 0][ m a x m mm mLiL e st =⋅=⋅∆=∆ ε (6) For s = 0,25 mm it results that the transfer factor ipg must have the value:

51,02 5,02

][2

=⋅

=∆⋅

=iLsip g (7)

For the measurement accuracy and for decreasing the weight of the pneumatic cylinder it is limited the total variation of the measurement pressure at Δpmax = 0,4 bar, which leads to:

25

m a x

m a x /1 040 1,0

4 0 0 0m

mNmP a

Lp

SK

m e f

a rc ⋅=⋅

=∆∆

= (8)

With the relation (8) may be dimensioned the pneumatic cylinder in constructive correlation with the assembly of the measurement device.[11] The transition from 20 to 150 ºC may generate dilatations of the cylinder (13), of the rigid centre (14), as well as modifications of the elastic characteristic of the spring (14). These may influence the precision in measuring the contraction of the brick, during the drying process. The material from which will be made the cylinder, the rigid centre and the spring is superinvar type (58% Fe + 42% Ni) which has a dilatation coefficient of α1 = 4 · 10-6 until 300 ºC and it maintains its elastic characteristic until 150 ºC. Therefore the effective average dyameter of the goffered membrane will have the real value of :

)1()( 10 TDTD m e fm e f ∆+⋅= α (9) where: ΔT = Tmas – 20 is the temperature difference The error εD caused by the temperature variation will have the value:

( ) ( ) 5262m a x1

2

0

m a x 0 2 7,411 3 041111)( −− =−⋅+=−∆+=−

= EET

DTD

m e f

m e fD αε (10)

From the relation (10) it results that the error εD caused by the thermal dilatation of the pneumatic cylinder is very low and it may be ignored in the measurement of the contraction.

68


3. Conclusions

For measuring an important parameter of the drying process of the raw ceramic products, the linear contraction, parameter that must be measured in an environment with a relatively high temperature 120..150 ºC, it is proposed the use of an unconventional pneumatic transducer by means of which may be performed precise measurements, which has a plain structure and is much cheaper than the similar electronic variants. The pneumatic transducer for measuring online the contraction of the raw bricks during drying has more precision at measurement, more than 1% for an initial distance between the measurement anchors of 100 mm. It is conceived for a safe and easy use, it couples with the outside by means of a pneumatic pipe made of teflon with the dyameter of 4 mm and an electric conductor of 1 mm2 has at wires a PLC compatible electric signal. Were used the concepts specific for low cost automatisation – which led to the achievement of a plain and precise device, much more cheaper than other variants in use. The metallic materials used for the pneumatic cylinder, spring and membrane have the dilatation coefficients very small which ensures a very slight variation of the effective average dyameter of the goffered membranes below 4 · 10-5, which leads to a very high measurement precision.

References

[1]. Berling B., Heinrich B., Thrun W., Vogt W., Kaspers/Küfner Messen- Steuern- Regeln: Elemente

der Automatisierungstechnik, Springer DE, 2005 [2]. Douglas M. Considine, Process/Industrial Instruments and Controls Handbook, 4th Edition, ,

McGraw-Hill, 1993 [3]. Hasatani M., Itaya Y., Muroie K., Contraction Characteristics of molded ceramics during drying, I.J.

Drying Technology, Volume 11, Issue 4, 1993 [4]. Hitoshi Takeda, CIA - Low Cost Intelligent Automation: Produktivita ̈tsvorteile durch

Einfachautomatisierung, Landsberg am Lech : Mi-Fachverl. Redline, 2006. [5]. Murad E., The measurement of the parameters of the drying process of ceramic products with

unconventional pneumatic transducers, Simpozion HERVEX 2007, Călimăneşti, 14-16 noiembrie, 2007

[6]. Murad E., Chercheş T., Unconventional pneumatic transducers with low energetic consumption for measuring the forces from the agricultural inbstallations and in food industry HERVEX 2008, Călimăneşti 15-17 noiembrie 2008

[7]. Murad E., Dumitrescu C., Haraga G., Dumitrescu L., Pneumatic metering systems for amount of water extracted in convectiv drying processes, International Scientific Conference Conference - DTMM, Iaşi, 14 -16 mai 2010

[8]. Murad E., Dumitrescu C., Haraga G., Dumitrescu L. Force pneumatic transducers with low energy consumption in stochastic measurement operations; Simpozion HERVEX, 2010, Călimăneşti, 10-12 noiembrie 2010

[9]. Murad E., Dumitrescu C., Haraga G., Dumitrescu L., Pneumatic transducers for measuring the speed of drying ceramic materials, SINUC 2010, Al XVI-a Simpozion National de Utilaje pentru Construcţii, Bucureşti, 16-17 decembrie 2010

[10]. Radcenco V., Alexandrescu N., Ionescu E., Calculation and design of the pneumatic elements and schemes, Editura Tehnică, Bucureşti. 1985

[11]. Comes M,Drumea A.,Mirea A., Matache G. – Intelligent servohydraulic device for the control of motion – 24th International Spring Seminar on Electronics Technology: Concurrent engineering electronic packaging, conferince proceedings ISSE 2001, 05-09 may 2001, pag.282-285

[12]. * * * Displacement Mesuremment, Linear and Angular, CRC Press LLC, 1999

69


THEORETICAL CONSIDERATIONS REGARDING THE MECHANISM FOR ADJUSTING THE CAPACITY OF THE PUMS WITH RADIAL

PISTONS

Lepadatu Ioan1, Dumitrescu Catalin1

1 INOE 2000 -IHP Bucuresti, e-mail: [email protected] ; [email protected]

Abstract The pumps with radial pistons -ppr- have been little studied in the scientific engineering

environment from Romania. Due to the fact that these pumps are very important for the field of

hydraulic and pneumatic drives, IHP has set as one of its objectives, the exhaustive study of this

issue and the finding of some innovative solutions which to lead to higher performances and a

diversification of the applications in which are used pumps with radial pistons.

This article presents the main theoretical elements of the dynamics of the mechanism for adjusting

the capacity of the pumps with radial pistons.

1. Introduction

The variation of the flow of the pumps with radial pistons is performed by a positioning

servomechanism which modifies the capacity (geometrical volume) of the pump [1], [4]. For

explaining explicitly its operation below is briefly described the construction and operation of the

pumps with radial pistons.

The pumps with radial pistons-fig.1 consist of the stator ring 1 in whose interior is wheeling the

rotor 2 that has, radially positioned, a certain number of cylindrical cavities 3 in which may displace

a corresponding number of small pistons 4. The fluid is aspired from the circuit by the aspiration Ca

in the aspiration chamber A from where by pumping is carried in the discharge chamber R,

respectively in the discharge circuit Cr.

During the wheeling of the rotor the piston heads press on the interior wall of the stator ring, due to

the centrifugal force and a guiding system. The axis I – I of the rotor is displaced with distance e

facing the axis II – II of the stator. At a complete rotation of the rotor each small piston makes a

displacement on a radial direction towards the exterior, during the crossing of the arc length aa and

a displacement towards the interior during crosssing the arc length a’a. The radial pistons make

during the rotor wheeling a relatively alternative rectilinear move to the rotor.

During the displacement towards the exterior the piston cylinders are in connection with the

aspiration chamber A, performing the aspiration.

70



Fig. 1

During the displacement towards the interior the cylinders are in connection with the discharge

chamber R, performing the discharge. The pumping becomes possible due to the relatively

excentric position of the rotor towards the stator ring. The flow of the pumps with radial pistons is

proportional with a constructive constant C and with two variable parameters, eccentricity e and

turation n of the pump.

neCQ ⋅⋅=

The result „C · e” is called capacity or geometric vollume and it is expressed Vg = C · e.

The variation of capacity it is obtained by the variation of eccentricity between the two set limits + e

şi – e.

By the eccentricity variation varies the relative position of the rotor to the stator ring and as a

consequence the displacement speed of the small pistons in the cylinders. At null eccentricity the

flow is also null.

2. The forces developed by the pistons of the pumps with radial pistons The symbols used and their meaning

Fp axial force given by the piston

FM the Fp projection on the normal n at the sliding ring in the contact point

Fx, Fy the projections of force FM on the system of axis X O Y

ρ instant radius units the rotation center O1 with the point of contact on the sliding ring; e the

eccentricity of the pump;

β the angle between the direction of the piston axis and the normal at the cirumference of the

sliding ring in the point of contact of the piston axis; R the radius of the sliding ring; dp the

dyameter of the piston; α the current angle of rotation of the piston K – 1; z the number of pistons;

71


na the number of active pistons in discharge; lp the length of the piston; mb the minimum contact

length rotor piston; L the hypothetical contact length sliding ring – case; j the gap case-sliding ring;

( )αb the instant contact width sliding ring-case; b the average contact length rotor-piston; η

dynamic viscosity

⋅ smkg032,0 ; a,b the halfaxis of the ellipse; 2 ϕ the angle between 2 pistons;

RE the exterior radius of the stator ring; RM the average radius of the stator ring; E elasticity

module

⋅⋅ 2

11101,2sm

kg ; x the halfwidth of the segment of intersection ellipse circle; F the

resultant of forces; f distortion from the circular shape under the action of the force (hypothesis

rigid case); tH, th tolerances at adjustment corresponding to the dimension RE; m total mass; TC

time constant of the mechanical system; KE the slope of the characteristic ( )ifp =∆ ; a0,

a1 modules area mass; I inertial moment

Fig. 2

a) The kinematics of the system

ρ== MORMO 12 ; ; It is applied the theorem of sinuses and it is obtained:

( )αβπρ

αβ −−==

sinsinsinRe

; ( )[ ] ( ) ( )αβαβπαβπ +=+−=+− sinsincossin

αβαβ

sinsin;sinsin R

eRe==

72


At the axial components of the forces Fp with which the pistons operate on the sliding ring of the

pump – force Fx oposses force Fx developed by the linear hydraulic motor – with piston of higher

diameter of the system, (fig. 7) force which positions the ring in any point of the eccentricity – e.

β= cos F F pM The force of one piston: ( ) ( )β+αβ=β+α= cos cos F cos F F pM1x

where: ( ) ααααβαβαβα sin sin R

e - sin Re - 1 cos sin sin - cos cos cos 2

2

2

==+

Developing in binominal series and taking into account that ( ) 1sin max =α , and the relation

110−≅Re ; αααα 2

2

24

4

42

2

221

22

2

sin 2Re- 1 sin

Re

81 - sin e

21 - 1 sin

Re - 1 ≅⋅=

R

; 010 44

4

≅≅ −

Re

b) The precise formula of the unitary forces on the two directions x and y, is:

( ) ( )( ) ( )

++=

+=+==

βαββα

βαββαβ

sincossin

coscoscoscos

1

1

pMy

pMxpM

FFF

FFFFF

if ,0≅β it results:

c) The approximate formula of the unitary forces of a piston become:

Axial force : αcos1px FF = (in opposition with the force of the positioning mechanism);

Normal force : αcos1py FF = (normal force which determines the friction of the sliding ring).

d) The force in the piston The force developed by the pistons of the pump is given by the pressure of the hydraulic oil

from which is decreased the force caused by friction between the pistons and the rotor of the

pump:

;24

2

−−= V

jbpjdp

dF p

pp

ηππ

αωραρ sin: vitezacos: eVsieRCu +==−≅

Fig. 3

73


From fig. 3 it results:

( )( ) ( )

( ) ( ) ( )

+=+

−+=

−+=

πραρα

αρα

πρ

min

min

bbrbl

rbl

p

p

( ) ( ) ( ) ( )( )

−=+=

→−+=ααρ

πραρπρα

cosmin eReR

bb

( ) ( ) ( )( ) αα

ααcos

cos

min

min

eebbeReRbb

++=−−++=

( ) ( ) ( )[ ]

( )[ ] ebbDeciebeb

eebdxbb

+=+=+=

=++=−

= ∫

minminmin

0min0

:;1

sin10

1

ππ

ααπ

απ

α ππ

Replacing the expression of speed in the formula of the force developed by the pistons of the

pump it is obtained:

αωηππ

αωηπππ

sin22

sin24

2

jdeb

pjdd

Fej

dbp

djp

dF ppp

pppp

p +

−==+−⋅=

3. The force required at the positioning mechanism The positioning mechanism must conquer the sum of the forces Fx which represent the horizontal

components of the force of the active pistons (being under pressure in the discharge chamber) :

Noting with RFx this resultant su mis obtained:

( ) ( ) ( )

( ) ( )

−=

−−−=

=

−−−==

++==

∑ ∑

∑∑ ∑

= =

== =

απ

π

αππαπα

παπαπαα

sinsin

coscos2cos21sinsin21coscos

21sinsin21coscos21coscos

1 1

11 1

z

zz

Fz

Kz

KF

zK

zKF

zkFFFR

p

na

K

na

K

p

na

K

p

na

K

na

K

ppx

a) The normal force which determines the friction of the slide

The unitary force on axis y has the value ( )

−+=

zKFF py

πα 21sin: 1

The resultant of all pistons from the discharge chamber pressure chamber ÷ RFy is:

74


( ) ( ) =

π

−α+π

−α= ∑=

na

1Kpy z

21Ksincosz

21KcossinFFR

( ) ( ) ;z

21Ksincosz

21KcossinFna

1K

na

1Kp

π−α+

π−α= ∑ ∑

= =

First and second sum, respectively S1 and

S2 being

απ

π

+απ

=π

π

=π

= cos

zsin

zcos

sinz

2cosRF :finalIn

zsin

zcos

S;z

2cos S y21

b) sliding ring - housing friction In correspondence with the notations from fig. 4

If :j

LbKv 2η

= is coefficient of viscous friction,

Then existing two forces of the kind it results :

ring theof speed theis x where,xj

LbFf η

=

Fig. 4

c) The length of the contact surface: L L is deduced from the distortion of the ring under the action of all forces by determining the points

of intersection between the initial circle and the ellipse resulting due to the charge with the

respective forces.

The equations of the ellipse and circle from fig. 5 are :

( ) ( )

=+

=−

++

222

2

2

2

2

1

RyxfR

yfR

x

EE

As ;z

22 π=ϕ z-number of pistons

Deformation

ϕϕϕ

−ϕ

−ϕ

−= 2

3M

sincos

sin12

EI3RF

f:

In which the resistence module:

Fig. 5

75


−

−+

= hhRhRRRbI MM 2

2ln2 with expressions

−=

+=

−

→

RRh2

RRR

widthb

E

EM

The maximum force is the sum of the resultants:

( ) ( )2max

2maxmax yx RFRFF +=

For the position from fig. 6.7 may be written :

( )( )

maxmax

max

maxmax

sin α

α

FF

RFRF

tgy

y

=

=

zπϕ =

It is solved the system (the intersection between circle

and ellipse) and are obtained :

−=+=

=+−

−=RbRa

careinby

ayRyRx 1; 2

2

2

2222

22

2

2

11

1

ab

bR

y−

−+=

baRb

>>

Fig. 6

It results that the length L is: yL 2≤

d) The gap: j corresponds to the average game of the fit 77

hH

( )2

hH ttj

−−= hH tt ; tolerances at fit

May be calculated now:

⋅=η

η=

smkg0342,0

b

j,L

cu,j

LbK V

determined

the ring width

(dynamic viscosity)

In the end the equation of the friction force: xKF Vf ⋅= , becomes definite entirely.

76


4. The balance of the forces from the positioning mechanism In the correspondence with the notations from fig. 7 may be written:

( ) pappARFFFF xefi −∆−=+++ , where "A" is the area of the bigger piston and a" of the

small piston, p∆ being the pressure drop on the valve :

The inertial force:

( ) xMmMF ippi ⋅++= , with

Mp: the mass of the bigger

piston;

Mp : the mass of the smaller

piston;

Mi: the mass of the sliding ring

The elastic force:

Fe = (K1 – K2) x,

K1 and K2 – the rigidity of the

arcs

The viscous friction force:

Ff = KV x

Fig. 7

The resultant of the axial forces RFx is:

−⋅⋅= απ

π

απϕ

πsin

sin

coscos2cos

2

z

zz

pd

FR px where: tωα =

If it is noted : ( )

−= απ

π

απϕ

πsin

sin

coscos2cos

2

z

zz

dtf p , results: RFx = p f (t)

In the end the equation of balance has the following structure:

( ) ( ) ( )[ ]( ) ( )

=∆

=α=∆==

∆⋅−−−=−++

p41p

slope sticcharacteriK ;Ktgarc;iKp00x0x

pAptfaAxKKxKxm

EEE

21V

77


5. Remarks and conclusions The determination of the constructive operational parameters and the dimensioning of the

servomechanism which adjusts the capacity of the pumps with radial pistons lead us to complex

mathematical relations and at non linear diferential equations whose solving is made using

numerical calculation systems.

The electrohydraulic adjustment systems [2] are very complex systems where take place

phenomena associated to the flowing of fluids from the field of volumetric hydraulic machines and

also phenomena specific for the processes of automatic adjustment. Due to the complexity of

these phenomena the finding of most adequate solutions in designing and realizing the mit is made

iteratively. Reaching such performances [3] implies the use of the methods of mathematical

modelling and numerical simulation of these systems.

Reference:

[1] T.C Popescu, D.D. Ion Guta, I. Lepadatu & P. Drumea, „Experimental Research on Servomechanisms

That Adjust Capacity of Radial Piston Pumps”, Proceedings of The 21th International Conference on

Hydraulics and Pneumatics, June 1-3, 2011, Ostrava, Czech Republic, ISBN 978-80-248-2430-7, 37-44 pp.

[2] T.C. Popescu, D.D. Ion Guţă, C. Calinoiu, “Modern instruments for analysis of hydrostatic transmissions“,

U.P.B. Sci. Bull., Series D, Vol 72, Iss 4, 2010, 201-210 pp.

[3] D.D. Ion Guta, T.C. Popescu, C. Dumitrescu, „Optimization of hydrostatic transmissions by means of

virtual instrumentation technique”, Proceedings of SPIE Vol. 7821, Advanced Topics in Optoelectronics,

Microelectronics, and Nanotechnologies V, 19 November 2010, DOI: 10.1117/12.881904

[4] I. Lepadatu, „Experimental Research focused the mechatroic positioning systems for regulating th

geometrical volume of the pumps with radial pistons”, UPB Scientific Bulletin, Series D, vol. 72, Iss. 4, 2010,

ISSN 1454 – 2358, 161 – 174 pp.

78


CAVITATION EROSION RESISTANCE FOR A SET OF STAINLESS

STEELS HAVING 10 % NICKEL AND VARIABLE CHROMIUM CONCENTRATIONS

Ilare BORDEAȘU1, Mircea Octavian POPOVICIU2

1 „Polytechnic“ University of Timisoara, Mihai Viteazul No.1, 300222, Timisoara, Romania, E-mail: [email protected] 2 Academy of Romanian Scientists, Timisoara Branch, Mihai Viteazul No.1, 300222 Timisoara, E-mail:

[email protected]

ABSTRACT: Regardless if the erosion phenomenon takes place in a laboratory facility or in an industrial device cavitation erosion intensity depends on two different factors: the quality of the steel and the intensity of the cavitation. Researches to obtain better materials are done every time in laboratory devices in which the cavitation intensity is very great and the research time is reduced. In most cases, the intensity of cavitation in industrial devices is smaller. The present laboratory researches upon eight stainless steels with great content of austenite are important because such materials are used to repair by welding the affected details. The chemical compositions were established as follows: the Nickel content approximately the same 10%, two contents of carbon 0.1% and 0.036% and eight different Chromium contents between 6 to 24 %. The laboratory facility is a device with piezoelectric crystals respecting the ASTM G32-2010 Standard. The laboratory results show that all the tested steels have very good cavitation erosion resistance; the best obtained result is for the steel having 6% Chromium and 0.1% Carbon with the structure having 32% martensite and 68% austenite. It is interesting to note that this result is better than that obtained for steels with greater content of martensite. Key words: stainless steel, cavitation erosion, microstructure, vibratory test facility 1. Introduction The great majority of the modern hydraulic machineries have the runners or the blades made from stainless steels with reduced carbon content (under 0.1%; the reason is to have good weld ability, without heat treatments), low nickel content (about 5%, the reason is the cost reduction) and a high content of Chromium (about 13%). The material structure is composed mostly by martensite giving high mechanical characteristics and also high cavitation erosion resistance. The repair works are done using electrodes with austenitic or austenite-ferrite structure. The problem is to choose electrodes depositing a material with higher cavitation erosion resistance than the genuine one. The cost of the material has not great importance because the used quantity is relatively small. The present research is directed towards such materials with improved cavitation erosion resistance having high content of Chromium and Nickel and austenite structure. If such materials will have also low costs, in the future, it will be possible to use them also for manufacturing the whole runner. 2. Tested materials The eight materials tested in the present research have a constant nickel content (approximately 10%), and variable chromium and carbon content. From the point of view of carbon content there are divided in two groups: four of them have 0.1% C and the other four 0.036% C. The steels from the first group have the following chromium content 6%, 10%, 18% and 24%. The steels from the second group have the following chromium content 13%, 14%, 16% and 18%. The cavitation erosion specimens were manufactured from small cast samples subjected to heat treatments. The heat treatment consisted in: homogenization annealing and solution quenching

79




(with air cooling for steels with martensite and ferrite structures or water cooling for steel with austenite and ferrite structures). In Table 1 are presented the mechanical characteristics and in Table 2 the micro structural constitutions determined from the Schäffler diagram on the ground of Chromium (Cre) and Nickel (Nie) equivalents [6]. Because the evaluation of the cavitation erosion resistance is done by comparisons with the steel OH13NDL with martensitic structure [3], [7], (a steel largely used for manufacturing hydraulic equipment in Romania) in both there are given also the characteristics of this stainless steel. For identification of the tested steels were utilized the principal chemical constituents (nickel, chromium and carbon) and the figures representing the concentration of those three elements. The content was symbolized as follows: for nickel Ni10; for chromium Cr6 to Cr24 (signifying 6 to 24%), for carbon C1 (signifying 0.1% C) or C036 (signifying 0.036%).

Table 1 Mechanical properties [8]

Steel Carbon content

% Rm

[N/mm2] Rp0,2

[N/mm2] HB

Ni10Cr6C1

≈ 0.1

1550 1120 489 Ni10Cr10C 1450 1020 447 Ni10Cr18C1 1335 934 372 Ni10Cr24C1 1280 901 307 OH12NDL 650 400 225

Ni10Cr13C036

≈ 0.036

856 618 276 Ni10Cr14C036 341 240 346 Ni10Cr16C036 996 700 309 Ni10Cr18C036 527 369 375

Table 2 Microstructural constitution [8]

Steel Cre [%]

Nie [%]

Structural Constituents

Ni10Cr6C1 11,924 15,173 32% M+68%A Ni10Cr10C 14,919 14,854 100%A Ni10Cr18C1 22,414 14,138 98% A+2%F Ni10Cr24C1 30,362 15,101 81%A+19%F

Ni10Cr13C036 13,209 11,454 55% M+45%A Ni10Cr14C036 15,022 11,4935 30% M+70%A Ni10Cr16C036 17,824 11,515 100% A Ni10Cr18C036 19,610 11,508 93% A+7%F

OH12NDL 13.2 4.45 88%M+12%F A, austenite, M-martensite, F-feritte

3. Test facilities and testing method The specimens were tested in a vibratory device with pieyoelectric crystals, realized in the Cavitation Laboratory of Timisoara Polytechnic Univerisity [8]. The facility parameters are: the generator power 500 W, the vibration frequency 20 kHz, the double amplitude 50 µm, the specimen diameter 15.8 mm, all parameters respect the ASTM G32-2010 Standard [2]. As testing

80


liquid was choosen the drinking water from the urban water-supply network and not the distilled water commonly recommended [3]. The motive was the fact that this water has physico-chemical properties closer to that of the river water where hydraulic machineries runns.

Test facility 1-Horn, 2-Electronic system, 3-Temperature

control system, liquid vessel and cooling coil, 5- Ventilation system

Diagram of test facility 1-Piezoelectric transducer, 2-Ultrasonic

generator, 2- Cooling system, 3-Liquid temperature control

Fig. 1. Vibratory device

In conformity with the procedures apllied in our laboratory [8] the total cavitation exposure was 165 minutes, divided in 12 intervals (one of 5, one of 10 and the rest of 15 minutes). To put into evidence the behavior in which the exposed area, respectively the material stucture, resisted to cavitation, after the total exposure time (165 minutes) the attacked areas were examined with optical microscopes (magnification x4, x10, x20, x40x and x80) and scaning electron microscopes (magnification x500). For a better examination, the eroded areas were attacked with nitromuratic acid (1/4 HNO3 – 3/4 HCl and 1-2 glicerine droplets) and a compound formed from 1/10 HNO3 and 9/10 water. The metalographic analyzes were realized at Bucharest Polytechnic University at the Center for Special Materials Survey (CEMS). 4. Test results. Discussions In figure 2 are presented images of the eroded areas and their structure after 165 minutes of cavitation exposure obtained with an „OPTICA” microscope and the electronic „Philips XL30 ESEM” microscope. In order to analyze the cavitation structural degradations, fig.2, poz.1, the attacked specimens were axial seectioned, metalographical prepared and studied with a SEM microscope. The following conclusions were obtained:

1. Ni10Cr6C1 shows a mxit aspect with very fine caverns uniformly distributed on the surface, with intergranular propagation of cracks. The fractures have fragile aspect.

2. Ni10Cr10C1 presents caverns with great dimensions, over 200 µm, inter-granular cracks and cleavage planes. The fractures present a fragile character and are propagated through slipping lines.

3. Ni10Cr18C1 and Ni10Cr24C1 show caverns with great dimensions, over 200 µm and mix propagation of the fracturing front through inter-granular cracks and cleavage planes. The fracture has a fragile character.

4. Ni10Cr13C036 and Ni10Cr14C036 present aspects of fragile rupture with fine and very fine caverns. The fracture propagate through inter-granular cracks and cleavage plans.

81


5. Ni10Cr16C036 has a mix aspect with very fine caverns, uniformly distributed on the

surface. There were observed cleavage zones and inter-granular cracks with radial propagation. The fracture has a fragile character with intergranular propagation.

6. Ni10Cr18C036 shows great caverns. The fracture has a fragile character with intergranular and cleavage propagation. There were observed secondary cracks, cleavage planes and the fracture propagate along sliping lines.

a-Ni10Cr6C1

b-Ni10Cr10C

c-Ni10Cr18C1

d-Ni10Cr24C1

e-Ni10Cr13C036

f-Ni10Cr14C036

g-Ni10Cr16C036

h-Ni10Cr18C036

Fig.2 Images of the structures and the erosions produced on the exposed areas (after 165 minutes of cavitation exposure) (1 – eroded microstructure obtained with a scanning electronic

82


microscope (SEM), (x500); 2- erosion in a cross section normal to the eroded area with the

maximum depth penetration erosion put into evidence, (x4); 3-steel microstructure before the exposure (x500))

The quantitavive differences between the cavitation erosion resistance of various steels can be apprecaited better by comparing MDER(t) (the mean depth erosions rate curves) of the researched steels with the curve of the standard steel OH12NDL, fig.3, [3,4,7], on the ground of the MDE (mean depth erosions), fig. 4 or the maximum measured depth of the erosion measured in the axaial cross-section, fig. 5.

Fig. 3 Mean depth erosion rate against exposure time

The evolution of the curves in fig. 3 prezent zones of gradually increasing of MDER till 40 to 90 minutes of exposure. For the resistant steels this time is smaller than that for the weacker materials. After reaching the maximum rate this value remains approximatelly constant. Such an evolution characterises materials with high cavitation erosion resistance [3] [10]. With the exception of Ni10Cr18C036, which is a little weaker, all the tested steels have better cavitation erosion resistance than the standard steel OH12NDL. As a consequence, from the point of view of erosion all the tested steels can be used either for manufacturing or for repair works of details subjected to cavitation. From the studied materials, Ni10Cr6C1 has the most favorable behavior. The steels Ni10Cr10C1, Ni10Cr18C1, Ni10Cr24C1 even if are a little weaker than Ni10Cr6C1 remain steels with excellent cavitation erosion behavior. The differences between them appear in the first period of exposure and are without importance [1], [3]. The superior behavior of Ni10Cr6C1 which has only 32% martensite in comparison with Ni10Cr13C036 having 55% martensite can be explained by the increased percentage of carbon, which increases the hardness of the material [9].

83


Fig. 4 Cavitation resistance comparisons after the computed mean depth erosion with the microstructure put into evidence

Fig. 5 Comparisons of different penetration depths: „EPmax maximum measured erosion penetration”, „MDE computed mean depth erosion” iar in ordonata „Penetration depth PD” The hystogram in fig.4 show that after an exposure of 165 minutes all steels with 0.1% C present mean depth erosions smaller than the steels with 0.036, regardless of the microstructural constitution. We appreciate that this situation is principally determined by the unstable austenite, wich under the bubble implosions impact is localy transformed into martensite. The conclusion results from the comparisons of the excellent resistance steel Ni10Cr16C1 (100% austeniitic structure, but unstable) with the lower cavitation erosion resistance steel Ni10Cr16C036 (100% austeniitic structure, but stable). This fact shows the beneficial effect of the increased carbon content, even if the structure is the same. Our conclusion is that the carbon content must be reduced but only to a value giving an acceptable weldability but maintaining the unstable austenite. The future researches must be foccused on this condition.The increase of the ferrite content for the steels with 0.1 C worsen the resitance to cavitation erosion but this decrease is not in direct proportion with the ferrite increase.

84


Interesting conclusions appear also for the comparisons between the excelent cavitation erosion steel N10Cr24C1 (81% A and 19% F) and N10Cr18C036 (93% A and 7% F) steel with a wicker cavitation erosion behavior, even if the last has a smaller content of ferrite but has not sufficient carbon. The hystogram in fig.5 show a great diffeence between the value of the mean depth erosion “MDE” computed from the mass los of the specimen during the entire exposure (165 minutes) [5], and the greatest depth of caverns “EPmax” mesured in the axial cross section. We consider that the value chose in ASTM G32 Sandard, namely “MDE” is the correct one because it take into account the whole eroded mass. This value and must be compulsory adopted for the evaluation of the various material resistance to cavitation erosion. We also note that the value “EPmax” is relatively difficult to appreciate, because it has great variation for different axial cross sections. 5. Conclusions 1. In comparison with the standard material OH12NDL all the researched steels present better cavitation erosion resistance, so they can be used for repairing or even manufacturing blades and runners for hydraulic machines 2. The best cavitation erosion resistance was obtained for the stainless steel having 0.1% C and 6% Cr and a structure with 32% martensite and 68% austenite. 3. From steels having 0.036% C the specimen with 13% Cr having the structure composed by 45% austenite and 55% martensite present the most reduced cavitation erosion resistance. 4. The microstructure has a great influence upon the cavitation erosion resistance. Especially the presence of martensite improves the cavitation resistance. 5. An increased content of carbon content also improves the behavior of the steels to cavitation erosion. All researched steels with 0.036% have smaller cavitation erosion resistance than those with 0.1% C. 6. The Chromium content has an important effect in establishing the proportion between the micro structural constituents and in the same time upon the mechanical properties and the cavitation erosion resistance. The increase of the Chromium content reduces the erosion resistance because the ferrite zone is amplified. 7. The austenite increases the erosion resistance because during the cavitation attack the hardness is increased, or even martensite is formed by bubble implosions. 8. The mean depth of erosion, computed in conformity with the G32-2010 Standard is an excellent indicator for cavitation erosion comparisons between various materials. 10. The maximum penetration depth of the eroded area is not recommended for establishing the cavitation erosion behavior of different materials.

REFERENCES [1] Anton I., Cavitatia, Vol I, Editura Academiei RSR, Bucuresti, 1984. [2] *** Standard method of vibratory cavitation erosion test, ASTM, Standard G32-2010 [3] BORDEASU, I.: Eroziunea cavitaţională a materialelor, Editura Politehnica Timişoara, 2006, 208p [4] JURCHELA A.D., BORDEASU I., MITELEA I., KARABENCIOV A., Considerations on the Effects of Carbon Content

on the Cavitation Erosion Resistance of Stainless Steels with Controled Content of Chromium and Carbon, 21st International Conference on Metallurgy and Materials- 25th, 2012, Brno, Czech Republic, pp.718

[5] JURCHELA, A.D., BORDEAȘU I., MITELEA I., KARABENCIOV A., Considerations on the Effects of Carbon Content on the Cavitation Erosion Resistance of Stainless Steels with Controled Content of Chromium and Carbon. METAL 2012, 21st International Conference on Metallurgy and Materials, May 23-25, 2012, Brno, Czech Republic, pp.718.

[6] BORDEASU I., MITELEA, I., KATONA, S.E. Considerations regarding the behavior of some austenitic stainless steels to cavitation erosion, METAL 2012, 21th International Conference on Metallurgy and Materials, May 23-25, 2012, Brno, Czech Republic, pp.730.

[7] BORDEASU I., MITELEA, I., POPOVICIU, M.O., CHIRITA, C. Method for classifying stainless steels upon cavitation resistance, METAL 2012, 21th International Conference on Metallurgy and Materials, May 18-20, 2011, Brno, Czech Republic, pp.626

[8] KARABENCIOV A., Cercetări asupra eroziunii produse prin cavitaţie vibratorie la oţelurile inoxidabile cu conţinut constant în nichel şi variabil de crom, Teza de doctorat, Timișoara, 2013, pp.188

[9] Mitelea I., - Studiul metalelor, Litografia Institutului Politehnic”Traian Vuia” Timisoara, 1983

[10] FRANC, J.P., MICHEL, J.M. (2004) Fundamentals of Cavitation, Kluwer Academic Publishers, P.O.Box, 322, 3300 AH Dordrecht,The Netherlands

85

http://www.astm.org/Standards/G32.htm


VIRTUAL INSTRUMENT FOR PLOTTING SERVO-VALVES CHARACTERISTICS AFTER SIGNIFICANT MAINTENANCE OPERATIONS

Radu RĂDOI1, Iulian DUȚU2

1 Hydraulics and Pneumatics Research Institute, [email protected] 2 Hydraulics and Pneumatics Research Institute, [email protected]

Abstract: Maintenance using virtual instrumentation for diagnosis has entered the field of servo-hydraulics few years ago, offering new accurate methods and concepts. The authors have developed a virtual instrument for plotting servo-valves characteristics, achieving larger flexibility and versatility compared to classic methods, thus shortening the time allocated for reconfiguration and recalibration of devices and equipments mounted on a specific test stand. Proposed virtual instrument acquires, conditions, processes and stores data automatically, using structured database model, thus reducing the intervention of the human operator, increasing overall accuracy. The virtual instrument that the authors propose has large integration capabilities into informatics systems and can be set to work with other types of hydraulic equipment besides servo-valves. Keywords: servo-valve, virtual instrument, test stand

1. Introduction

In present, static and dynamic testing of servo-valves it’s done by using complex testing stands and installations with large number of electro-hydraulic equipment connected together thus maximizing the risk of being influenced by perturbations such as electromagnetic interferences, noises, unwanted vibrations and electrostatic discharges in the measurement circuits, all conducting to wrong results or damage of ESD sensitive equipment. The authors have developed and tested in laboratory conditions a virtual instrument for plotting the functional characteristic of servo-valves (hereinafter referred as VI) using a simplified hydraulic system.

Studying the basics of virtual instrumentation field, it can be seen that a VI has two main components: hardware and software, comprising switches, software buttons, knobs, sliders, digital indicators and so on, having a front panel that can be modified very easy. Virtual instrumentation design implies having a personal computer with specialized extension boards (such as data acquisition boards) and a software environment that together simulates the features and the operation of one device or measurement system. Transducers, analog to digital converters and the usage of data conditioning electronic circuits is mandatory. It can be said that one difference between virtual instrumentation and classic testing is that all control functions are automated and centralized using a personal computer with specific software.

2. Failure Causes in Servo-Valves

When a malfunction occurs in a hydraulic system and a servo-valve is the main cause of it, the operator must unmount it from the installation and put it on a test stand using either classic testing or modern technologies. For reducing downtime of the hydraulic system the operator can replace the defective servo-valve with another one with same functional parameters - if available. After mounting the defective servo-valve on the test stand the operator must perform certain tests in order to identify the cause of failure. It has been documented that common symptoms of defective servo-valves are:

- nozzles clogged, when the fluid flows only in one way, when applying an electric command; - broken coil, when the servo-valve does not respond to any electric command; - shifted null point, when there is flow without an electric command;

86


- asymmetry, having unequal flow values at equal electrical command for both polarities; - high wear of spool and sleeve, when high flow that cannot be canceled by adjustments

occurs in null point; - large hysteresis when reversing electrical control because of the friction between the spool

and sleeve due to residues present in hydraulic oil. Servo-valves must be diagnosed only by qualified personnel.

Fig. 1 – Electro-hydraulic stand for optimization of servo-valves

Maximum operating pressure of the test stand is 315 [bar] and the nominal flow is 50 [l/min]. The servo-valve that the authors used has the following characteristics:

- valve size: 04; - nominal flow: 10 [l/min]; - nominal pressure: 315 [bar]; - electrical command current: ± 20 [mA].

The VI described in this article is using a PC with two data acquisition boards and a control program for testing stand’s functions along with automated data acquisition, numerical filtration and storage of test data. It can be noticed that the human operator has a reduced interaction with the stand having a positive impact on system’s repeatability and accuracy along with shortening time needed for reconfiguration of physical connections between hydraulic and electronic devices. The VI is flexible and supports fast recalibration and reconfiguration, having the possibility to define and store certain test configurations.

The VI can perform different servo-valves tests; most important ones are step response and sinusoidal signal characteristics. Key parameter values for plotting step response characteristic are given by rise time, stabilization time and overshoot value. When plotting the sinusoidal signal characteristic it is necessary to draw the dependencies between: amplitude and frequency, phase and frequency and to plot Bode diagrams.

Another module of the VI generates test signals needed for both types of characteristics while other module acquires experimental data, in count of 25000 samples per channel. Test signal’s parameters and number of samples can be modified using VI’s configuration module. Experimental data are processed automatically and then displayed on a graphical control placed on the front panel of the VI.

Interacting with the VI is simple and intuitive: after displaying the main panel it will be initialized the data acquisition boards – including proper calibration and scaling - followed by setting all program constants. Selected type of test will begin when the human operator clicks on the START button – running signal generation, data acquisition, numeric filtering, data processing and storage on a column-separated values .DAT file, for later processing.

The VI will plot the tested servo-valve’s characteristic checking the following:

87


- hysteresis, defined as the difference between the command size, usually a current within

±20 [mA] supplied by an electronic module with input signal of ±10 V, required to achieve a flow rate variation in upwards with the input level , from minimum to maximum, and the control level necessary to obtain same flow rate at a command size variation in downward from maximum to minimum;

- linearity, determined by the maximum value of difference between command level from real diagram and that obtained on theoretical diagram (drawn between extreme points from the hysteresis diagram);

- repeatability, maximum difference between values obtained at the same level of electric command value;

- sensitivity, defined by the ratio between output value variation and the corresponding variation of input value, in case of a linear static characteristic;

- the characteristic of flow rate – command level from input (usually a current within ±20 [mA] supplied by an electronic module with input signal of ±10 V) at a constant pressure drop;

- the pressure-flow characteristic given at a constant electric command; - minimum operating pressure, which is the lowest value at which the flow can be adjusted

on the entire operating range. 3. DAQ Structure

Proposed data acquisition structure is simple and classical having simultaneous parameters acquisition, comprising the following:

- specific sensors and transducers (force, flow, pressure); - measurement amplifiers for system’s transducers; - data acquisition board (technical characteristics are given below); - virtual instrument; - electric power source and voltage stabilizer.

During the test process, the data acquisition structure will perform the following: - acquisition of parameters values; - conditioning and processing acquired data; - storage of data; - graphical interface with human operator; - graphical display of parameter variation or user defined characteristics.

Fig. 2 – VI’s main panel

The data acquisition board used is National Instruments USB-6218, with the following technical features:

- 32 single-ended or 16 differential analog inputs; - 16-bit resolution;

88


- 2 x 16-bit analog outputs; - sampling rate of 260kS/s; - input voltage range: -10…10V; - 8 TTL input channels and 8 TTL output channels; - on-board sample memory: 4095 samples; - digital trigger. The human operator can perform specific static and dynamic tests on servo-valves - most

significant are step response and dynamic response at sinusoidal signal input. The authors have developed the virtual instrument using LabView, because of its easy

integration with measurement and control processes. The VI can perform static and dynamic tests with a minimum need of human operator

intervention. Experimental data for the selected test are acquired continuously after pressing the START button on main panel of the VI. Data are displayed on a graphic control.

Fig. 3 – Adjustment characteristic of a servo-valve

4. Conclusions

The plotting of functional characteristics of servo-valves using presented VI allows operators to evaluate the functioning state after maintenance operations. Due to its modularity, the VI is easy to be interfaced with common electro-hydraulic systems.

Supply power for electronic modules can be taken from main electrical installation of the hydraulic system or test stand, current consumption being low.

By combining electronic modules and informatics technologies it can be simplified the implementation and reconfiguration of servo-hydraulic systems, using specific transducers, data acquisition boards, and virtual instrumentation.

REFERENCES

[1] D.I. Guta, C. Dumitrescu, I.Lepadatu, C.Cristescu: Experimental identification of electrohydraulic servomechanisms with virtual instruments technique, HIDRAULICA no.3/2010, pp.49-56.

[2] A. Drumea, P. Svasta, Microcontrolere pe 8 biti utilizabile în comanda şi controlul dispozitivelor hidraulice, Salonul National de Hidraulică şi Pneumatică HERVEX2002, 13-16 noiembrie 2002, Calimanesti-Caciulata, pp. 124-128.

[3] Guillon, M., Commande et asservissement hydrauliques et electrohydrauliques, Editions Lavoisier, Paris, 1996.

[4] P. Drumea, M. Blejan, A. Mirea, I. Ilie, G.Matache, Virtual Instrument Designed For Dynamic Tests Of Electro-Hydraulic Devices, ISSE 2007, pp.293-297.

[5] Information on http://www.ni.com.

89


THE INFLUENCE OF ANGLE OF TILT OF THE SEPARATORS AND THE

AIR COURSE VELOCITY ABOUT QUALITATIVE COEFFICIENT AND THE EXPLOATATION AT THE CLEANING AND SORTING OF THE CORN

PULSES Constantin POPA1, Mihaela-Florentina DUȚU2, Iulian DUŢU3

1 University POLITEHNICA of Bucharest, Faculty of Biotechnical Engineering, [email protected] 2 University POLITEHNICA of Bucharest, Faculty of Biotechnical Engineering,

[email protected] 3 Hydraulics and Pneumatics Research Institute, [email protected]

Abstract: This work aims at supplying optimum ways of peeling (cleaning) and sorting the wheat seeds. There has been used an experimental a stand made of two bodies out of which the higher one having adjustable angle of tilt (1…8°), and the lower one having a fixed angle of slope (15°). The machine also has a centrifugal ventilator which we have measured five values of the air-blast velocity like: 2, 3; 3; 5; 6 and 7 m/s. Keywords: tilt angle, air velocity, cleaning and sorting of corn 1. Introduction

The experimental stand used at the researches is destine to cleaning and sorting after the geometrical size (width and depth) and after the aerodynamic proprieties of the cereals, technique plants, obtained from the combines or threshers. The cleaning and the sorting are done until at the condition degree foresee in the standards in vigor for the products – goods.

2. Material and method

The experimental researches are done for corn of autumn at the three debits of feeding for the experimental stand, three air course velocity, regulate by the cooling machine and for four values of the angle of tiltβ, of the inferior separators of the superior framework with separators. The three air course velocities are: v1=5 m/s; v2=6 m/s; v3=7 m/s. The four values of the β angle are: β1=2°; β2=4°; β3=6°; β4=8°. At the experimental stand (fig.1) are used: the caliber, the measuring cord, the precision balance, balance type suitcase, chronometer, dynamometer balance, cup anemometer.

2.1. The establishing of the separation degree of the big impurities p from the initial mass of pulses in function of the inferior separator of superior framework

The separation degree of the big impurities and breaches p, represent the percentage of pulses and the big impurities calculate in report with initial mass of pulses (%).

The determinations are done at the feeding debit q = 1kg/s and 51 == vva m/s, for angle of tilt α=4°of a superior framework separator and for the angle of tilt β of the inferior separator of the same framework: 2, 4, 6, 8°. Example: α =40 and β = 20 - are gathered the big impurities; - are gathered in bags and then are weighed the quantity of the big impurities p; - are calculated in percentage the medium values for p;

90


- for β = 4, 6, 80 are proceeded similar; - are marked the graph p = f(β), for α = 4° and Vair=5 m/s.

For each of the air course velocity values Vair are marked the adequate graphs. Each of the previous determination presented are done for corn in three repetitions. 2.2. The establishing of the separation coefficient ε of the little pulses from the initial mass of pulses in function of the inclination of the inferior separator of superior framework separators The separation coefficient of little pulses ε from the initial mass of pulses, represent the report between the quantity of the little pulses effective separate from the separator of the length L and the quantity of pulses who are may separated through the separator (the last represented the quantity of the little pulses who are find in the initial material) [2], [3], [5], [6].

At the each test are timed the necessary time for effectuate this, with a view of determination of the work capacity. Then are determinate the work capacity of the machine with the relation:

tmQ 3600= (1)

where: Q – work capacity of the machine (kg/h); m – quantity of the conditioned barley (kg); t – necessary time for each experimental determination (s).

Then is determinate the separation coefficient of the little pulses ε from the initial mass of pulses, with relation:

LQcb ⋅⋅

+=

1

1ε (2)

where: ε – separation coefficient of the little pulses from the initial mass of pulses (%); b – separator width (mm); c – content of the little pulses from initial mass, at each determination (kg); L – separator length (mm).

Are marked the graphs ε = f (β). Each from the previous presented determinations is done for corn in three repetitions. 2.3. The establishing of the separation degrees of the little impurities c and the big impurities p of the separation coefficient of the little pulses from the initial mass of pulses ε in function of the air course velocity at the three debits of feeding The separation degree of the little impurities c, represent the pulses and the little impurities percentage calculate in report with initial mass of pulses (%) [1], [4]. The determinations are done for each from the five air course velocity values. at each of the three feeding debits considered. Are marked graphs: c=f(Vair) (figures 7, 8, 9, 10 – superior curves); p=f(Vair) (figures 11, 12, 13, 14); ε=f(Vair) (figures 7, 8, 9, 10 – inferior curves) for each of three feeding debits for corn. The separation coefficient of little pulses from initial mass ε is establish with relation (2) after the established the separation degrees for the little impurities c and respectively for the big impurities p and after the preliminary established the work capacity of machine Q with relation (2). The angles of tilt values of the two separators, α and β are combine between them: α= 4° and β = 2°, 4°, 6°, 8°. Also the determinations are done in three repetitions.

91


Fig.1.Correlation between the angle of tilt of an inferior separator of a superior framework β, for a angle of tilt of a superior separator of this framework of the separation of the little pulses from initial mass of pulses ε at 5=airV m/s Fig.2. Correlation between the angle of tilt of an inferior separator of a superior framework β, for a angle of tilt of a superior separator of this framework of the separation of the little pulses from initial mass of pulses ε at 6=airV m/s Fig.3. Correlation between the angle of tilt of an inferior separator of a superior framework β, for a angle of tilt of a superior separator of this framework of the separation of the little pulses from initial mass of pulses ε at 7=airV m/s

y = - 0,0025 x 2 0,001 x + 0,99 + R 2 0,9933 =

0,8

0,85 0,9

0,95 1

2 4 6 8 β (°)

ε(%)

y = - 0,0025 x 2 0,012 x + 0,975 + R 2 0,9947 =

0,9 0,92 0,94 0,96 0,98

1

2 4 6 8 β (°)

ε(%)

y = - 0,0062 x 2 0,0305 x + 0,945 + R 2 0,9341 =

0,75 0,8

0,85 0,9

0,95 1

2 4 6 8 β (°)

ε (%)

92


Fig.4.Correlation between the angle of tilt of an inferior separator of a superior framework β, for a angle of tilt of a superior separator of the same framework 04=α and the separation degree of the big impurities p at 5=airV m/s Fig.5.Correlation between the angle of tilt of an inferior separator of a superior framework β, for a angle of tilt of a superior separator of this framework 04=α and the separation degree of the big impurities p at 6=airV m/s

Fig.6.Correlation between the angle of tilt of an inferior separator of a superior framework β, for a angle of tilt of a superior separator of this framework 04=α and the separation degree of the big impurities p at 7=airV m/s

y = - 0,0006 x 2 1,4067 x + 3,6525 + R 2 0,991 =

5 7,5 10

12,5 15

2 4 6 8 β (°)

p (%)

y = 0,01 x 2 1,108 x + 6,7 + R 2 0,9892 =

5

10

15

20

2 4 6 8 β (°)

p (%)

y = - 0,0212 x 2 1,2635 x + 8,905 + R 2 0,9997 =

10

12,5

15

17,5

20

2 4 6 8 β (°)

p (%)

93


Fig.7.Correlation between air course velocity airV and separation degree of the little impurities c, respectively the separation pulses from pulses initial mass ε for the debit of 1kg/s, the separation inclination 04=α and inferior separator inclination 02=β Fig.8. Correlation between air course velocity airV and separation degree of the little impurities c, respectively the separation pulses from pulses initial mass ε for the debit of 1kg/s, the separation inclination 04=α and inferior separator inclination 04=β

Fig.9. Correlation between air course velocity airV and separation degree of the little impurities c, respectively the separation pulses from pulses initial mass ε for the debit of 1kg/s, the separation inclination 04=α and inferior separator inclination 06=β

y = 0,0018 x 2 0,02 x + 0,7647 + R 2 0,988 =

y = - 0,0021 x 2 0,0035 x + 1,8493 - R 2 0,8742 =

0,5

1,5

2,5

2 3 4 5 6 7 Vair (m/s)

c, ε (%)

y = - 0,0096 x 2 0,1329 x + 0,5297 + R 2 0,9857 =

y = 0,0379 x 2 0,5039 x + 3,2985 - R 2 0,9883 =

0

1

2

3

2 3 4 5 6 7 Vair (m/s)

c, ε (%)

y = - 0,0047 x 2 0,0687 x + 0,7416 + R 2 0,957 =

y = 0,0034 x 2 0,0506 x + 1,8947 - R 2 0,9572 =

0,5 1

1,5 2

2,5

2 3 4 5 6 7 Vair (m/s)

c, ε (%)

94


Fig.10. Correlation between air course velocity airV and separation degree of the little impurities c, respectively the separation pulses from pulses initial mass ε for the debit of 1kg/s, the separation inclination 04=α and inferior separator inclination 08=β Fig.11. Correlation between air course velocity and the separation degree of big impurities p for debit of 1 kg/s, the superior separator inclination of superior framework 04=α and inferior separator inclination of superior framework 02=β

Fig.12. Correlation between air course velocity and the separation degree of big impurities p for debit of 1 kg/s, the superior separator inclination of superior framework 04=α and inferior separator inclination of superior framework 04=β

y = 0,0107 x 2 0,059 x + 0,8486 - R 2 0,906 =

y = - 0,0151 x 2 0,0747 x + 1,8871 + R 2 0,872 =

0

1

2

3

2 3 4 5 6 7 Vair (m/s)

c, ε (%)

y = 0,1974 x 2 0,0712 x + 1,034 + R 2 0,9978 =

0

5

10

15

2 3 4 5 6 7 Vair (m/s)

p (%)

y = - 0,031 x 2 2,4758 x - 2,092 + R 2 0,9989 =

3 6 9

12 15

2 3 4 5 6 7 Vair (m/s)

p (%)

95


Fig.13. Correlation between air course velocity and the separation degree of big impurities p for debit of 1 kg/s, the superior separator inclination of superior framework 04=α and inferior separator inclination of superior framework 06=β Fig.14. Correlation between air course velocity and the separation degree of big impurities p for debit of 1 kg/s, the superior separator inclination of superior framework 04=α and inferior separator inclination of superior framework 08=β

3. Conclusions

1. Parameters p, c, ε vary with the separators inclination and the air course velocities concordant the graphs from figures 2, 3, 4, 5, 6, 7, 8, 9, 10, 11, 12, 13, 14, 15, but the variation mode is better appreciate of the integral rational function of second degree of shape ax2+bx+c.

2. The correlation coefficient 2R have the values very high, what demonstrate that the two curve the real one, and the theoretic one are identically or very close.

REFERENCES

[1] Căsăndroiu T. - „Utilaje pentru prelucrarea primară şi păstrarea produselor agricole”, Curs – vol 1, U.P. Bucureşti, 1993. [2] Dinu I. - „Curs de mecanică”, Editura Printech, Bucureşti, 1999. [3] Méchtcherski I.V. – „Recueil de problèmes de mécanique rationelle”; Editions Mir, Moscou, 1973. [4] Scripnic V., Babiciu P. – „Maşini agricole”; Editura Ceres, Bucureşti, 1979. [5] Targ S.M. - „Eléments de mécanique rationelle”; Editions Mir, Moscou, 1975. [6] Voinea R., Stroe I. – „Technical Mechanics”; U.P. Bucharest, 2000.

y = - 0,0228 x 2 2,4896 x - 0,5583 + R 2 0,9957 =

4

9

14

19

2 3 4 5 6 7 Vair (m/s)

p (%)

y = - 0,2741 x 2 5,0883 x - 4,3449 + R 2 0,9846 =

5

10

15

20

2 3 4 5 6 7 Vair (m/s)

p (%)

96

Hidraulica magazine 1 2013

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