Hidraulica Magazine 3 2013

No. 3/2013

ISSN 1453 – 7303 “HIDRAULICA” (No. 3/2013) Magazine of Hydraulics, Pneumatics, Tribology, Ecology, Sensorics, Mechatronics

CONTENTS

• EDITORIAL Petrin DRUMEA

5 - 6

• ELECTROHYDRAULIC SERVOMECHANISM FOR DRIVING THE REELS OF COIL WINDING MACHINES SPECIFIC TO WIRE ROLLING MILLS

Teodor Costinel POPESCU, Marian BLEJAN, Ioan LEPĂDATU

7 - 14

• RANS SIMULATION OF COMBINED FLOW AND HEAT TRANSFER THROUGH OPEN-CELL ALUMINUM FOAM HEAT SIN

Petre OPRIŢOIU

15 - 25

• ANALYSIS SOLAR RADIATION Carmen Otilia RUSĂNESCU, Marin RUSĂNESCU, Dorel STOICA

26 - 31

• REPAIRING AND TESTING OF THE HYDRAULIC SERVO VALVES Radu RADOI, Ioan BALAN, Iulian DUTU

32 - 37

• RESEARCH ON VARIATION OF DISPLACEMENTS, VELOCITIES AND ACCELERATIONS AT A SITE SELECTOR BLOCKS (FANNER) GRAIN

Dorel STOICA, Carmen Otilia RUSANESCU

38 - 42

• NUMERICAL SIMULATION OF FLOW IN ANASTOMOTIC COMPLEX AFTER GASTRIC RESECTION WITH GASTROJEJUNAL ANASTOMOSIS

O. VAIDA, Liviu VAIDA, A. ANDERCOU

43 - 52

• TESTING OF LINEAR PNEUMATIC ACTUATORS WITH HYDRAULIC LOAD Gabriela MATACHE, Stefan ALEXANDRESCU, Gheorghe SOVAIALA, Ioan PAVEL, Iulian-Cezar GIRLEANU

53 - 56

• ANALYSIS OF PRESSURE IN BUCHAREST BETWEEN 2009-2012 Carmen Otilia RUSĂNESCU, Gigel PARASCHIV, Gheorghe VOICU, Dorel STOICA

57 - 63

• EXPERIMENTAL TESTING OF A LOW SPEED HYDRAULIC MOTOR WITH AXIAL PISTONS

Laura GRAMA, Daniel BANYAI, Liviu VAIDA

64 - 73

• SONIC EFFECTS OF A UNCONVENTIONAL HEAT INSTALLATIONS Carmen BAL, Nicolaie BAL, Lucian MARCU, Carmen Ioana IUHOS

74 - 79

• ANALYTICAL MODEL OF THE CONNECTION PIPES OF THE ALTERNATING FLOW DRIVEN HYDRAULIC SYSTEMS

Ioan-Lucian MARCU, Daniel-Vasile BANYAI

80 - 85

3


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:

ISSN 1453 – 7303; ISSN – L 1453 – 7303

4


EDITORIAL CAUZE SI ARGUMENTE

In ultima perioada constatam o stagnare in dezvoltarea hidraulicii, la nivel national, intr-o masura ingrijoratoare pentru cei care stiu ce a fost si mai ales inteleg utilitatea acestui tip de actionare. Argumentele de tipul scaderii numarului de firme productive in domeniu, al scaderii numarului de specialisti si altele asemenea ar putea explica in mare masura situatia, totusi mai sunt si alte elemente de luat in calcul. Eu cred ca un element esential il reprezinta nivelul profesional din ce in ce mai scazut al specialistilor in actionari hidraulice indiferent de pregatirea lor scolara (academica). Ruptura creata intre nivelul prea ridicat si mai ales teoretic al pregatirii universitare si realitatea industriala, lipsa unei pregatiri coerente la nivelul muncitorilor si

Dr.ing. Petrin DRUMEA DIRECTOR INOE 2000 – IHP

tehnicienilor, precum si inexistenta unei calificari recunoscute pentru specialistii in hidraulica, au codus la ideea ca in tara noastra nu mai putem face prea multe in domeniu. Aceasta deficienta poate fi inlaturata cu ceva eforturi prin intensificarea perfectionarii profesionale in conformitate cu cerintele CETOP. O alta cauza ar fi ca prea multi conducatori de firme considera ca hidraulica este o “chestie” la indemana oricui si ca nu merita sa angajezi un specialist, care costa, pentru rezolvarea problemelor, cand se poate ocupa de acest tip de activitati, pe bani putini, orice persoana care are tupeu si ceva pregatire mecanica. Sigur ca in final lucrurile nu merg, sau merg prost, dar se poate da vina pe hidraulica spunand ca aceasta e imposibil de intretinut si ca oricum “merge si asa.” Nu este de neglijat nici situatia precara a dotarii unitatilor de mentenanta si de reparatii, ceea ce conduce la o depistare greoaie a defectiunilor si mai ales la o reparatie si o reglare finala neconforma cu cerintele functionale ale fiselor tehnice ale produselor. Probabil ca ar fi extrem de util sa apara in tara cateva firme de engineering capabile sa proiecteze modernizeze si intretina sistemele modern de actionare hidraulica. Este destul de clar ca astazi din cauza deficientelor prezentate si a inca multor altora s-a ajuns la rezultatul cel mai grav reprezentat de o crestere alarmanta a pierderilor energetice in cadrul utilajelor complexe, care nu se datoreaza utilizarii sistemului hidraulic ci interventiilor incorecte asupra acestuia. Lupta specialistilor pe plan international de a reduce pierderile energetice si a mentine actionarea hidraulica in contact cu dezvoltarea impetuoasa a actionarilor electrice si a actionarilor mecanice va trebui sa gaseasca si in Romania un raspuns favorabil. Agresivitatea nespecialistilor care se implica in domeniul actionarilor hidraulice ar putea fi diminuata doar de coeziunea putinilor specialisti care inca mai lucreaza in tara si care ar putea sa intervina pentru marginalizarea celor neaveniti.

5


EDITORIAL CAUSES AND ARGUMENTS We have found lately a certain stagnation in the development of hydraulics at national level, in such a degree that worries the people who know what it meant before and understand the utility of this type of drive. Such arguments as the decrease in number of the manufacturing companies from the field and of the specialists and many other of the same kind, could explain in a good measure the situation, however there are some other elements to be taken into account as well. I consider that an essential element is represented by the lower professional level of the specialists in hydraulic drives, regardless to their professional education. The break created between the elevate level of the academic education and the industrial reality,

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

the lack of a coherent training for workers and technicians as well as the inexistent qualification acknowledged for the specialists in hydraulics, led to the conclusion that in our country there is nothing more to be done in the field. This deficiency may be removed with certain efforts, by the intensification of the professional training in accordance with the CETOP requirements. Another cause may be that too many company managers consider that hydraulics is something out of anyone s reach and it is not necessary to hire a specialist who costs for solving the problems, when anyone can in fact deal with such matters and perform the related activities if has some guts and some mechanical training. Of course that in the end things don t work or go wrong but we can blame it on hydraulics saying that it is hard to provide its maintenance and anyhow it goes anyway. It must not be ignored either the precarious situation which regards the equipping level of the maintenance and repair units, fact that leads to a difficult screening of the flaws and to some repairs and final adjustment which are not at all in accordance with the operational requirements comprised by the technical sheets of the products. It might be extremely useful to be founded some engineering firms in the country, capable to design, modernize and maintain the modern hydraulic drive systems. It is clear enough that nowadays, cause of the deficiencies presented above and of many others it was reached the most severe result, represented by an alarming increase of the energetic losses at complex equipment, which is not caused by the use of the hydraulic system itself but by the wrong interventions performed at it. The fight of the specialists on international scale against the energetic losses and for maintaining the hydraulic drive in touch with the impeuous development of of the electric and mechanical drives must find a favorable answer in Romania as well. The aggressiveness shown by the non experts who involve in the field of hydraulic drives might be diminished only by the cohesion of the few specialists who are still working here and might intervene for the marginalization of the undesired ones.

6


ELECTROHYDRAULIC SERVOMECHANISM FOR DRIVING THE REELS OF COIL WINDING MACHINES SPECIFIC TO WIRE ROLLING MILLS

Ph.D. Teodor Costinel POPESCU1, Ph.D. Marian BLEJAN2, Ph.D. Ioan LEPĂDATU3 National Institute for Optoelectronics, INOE 2000-IHP Bucharest,

e-mail address: [email protected] ; 2 [email protected] ; [email protected]

Abstract: The drive system of a winding machine, mounted on the production line of a wire rolling mill, must perform at the same time driving in a rotary motion of the drum on which the wire is wound and linear displacement of the wire reeling device. On the dynamic performance of the operation of this system depends largely the quality of wire winding in the coils delivered to the beneficiaries, reducing losses caused by re-melting the wire improperly wound and productive capacity of the mill. This system is usually of two types: electromechanical, based on variable speed electric motor and ball valve screw or electro-mechano-hydraulic, without control loop, based on variable speed electric motor, speed reduction gear and hydraulic drive system on-off type. Typically, the drum which the wire is wrapped around is operated electromechanically for both types of systems, and the reeling device, which moves linearly, along the drum, the wire coming out of the mill, is driven by a ball screw, for the first type, or hydraulically, without control loop, for the second type. The authors of this material have developed and put into operation an electro hydraulic servomechanism for driving the reeling head that allows controlling its position and speed by means of an electro hydraulic control loop.

Keywords: reeling device, coil winding machine, electrohydraulic servomechanism

1. Introduction

The production practice of wire manufacturers, such as SC ALRO SA Slatina, shows that the drive systems of drums cause no problems in operation and exploitation; they achieve both a satisfactory uniformity of the rotation, for each layer of wire coiled, and a corresponding reduction in rotational speed at the beginning of each new layer of wire that will be wrapped. The two known types of drive systems of the reeling head however cause problems, regarding correlation with the rotational speed of the drum, for the first type, and problems of reliability, for the second type. In this context came the demand from SC ALRO SA, submitted to INOE 2000-IHP Bucharest, to develop a new solution for driving, operation and control of the reeling head, with operating parameters better than the existing solutions [1]. Achieving this new solution for control of the reeling head considered the operating conditions of winding machines, specific to wire rolling mills existing in Foundry Division at ALRO Slatina.

2. The operating conditions of winding machines specific to wire rolling mills

The winding machines, specific to wire rolling mills at ALRO Slatina, operate within foundry divisions where aluminum bars and wire are manufactured. The flow of production of these divisions is continuous, almost totally automated and it includes preparation of raw material in furnaces or electrolysis baths, feeding bar rolling mills and wire rolling mills, storage of finished products, namely wire bars and coils. A wire winding machine is the terminus point of a production line of aluminum wire. It contains two drums, operated successively, by means of a variable speed electric motor and a reduction gear with two output shafts. The wire coming out of the mill is directed along a groove towards the evenly rotating drum. The constant value of the speed depends on the diameter of wire that will be wrapped around the drum, namely: 9.5; 12; 15; 19.3 mm.

7

mailto:[email protected]




Fig.1. Wire production flow (back to front):

furnace, rolling mill, gutter, winding machine with two drums and two reels.

The technology of winding includes the next steps: a) Before winding wire around one of the two drums, it begins to rotate uniformly accelerated,

until it reaches the constant speed specific to the diameter of wire coming out of the mill. Simultaneously, the device for wire reeling on the drum, electro hydraulically actuated, moves to the right end and waits for the wire from the rolling mill.

Fig.2. Two drum coil winding machine idle: both reels stationed at the flange on the left of drums; right,

foreground –gutter through which the wire comes from the rolling mill; right, background - gutters through which the wire is directed to the drums, where it is taken up by the reel.

8


Fig.3. Reeling head, driven by a hydraulic cylinder with fixed bilateral rod and mobile liner.

b) When the wire coming out of the rolling mill, directed along the gutter by the guide in the reeling head, comes into contact with the drum, there is actuated a device for clamping the end of the wire on the drum. The reeling head is stationary during winding the first spire, then it moves to the left. It contains two limiters, one of which controls the duration of standing and the other controls change in direction of reeling head motion (at the end of stroke). Movement of the reeling device is correlated with the drum rotational speed as follows: during one rotation of the drum, the reel moves uniformly and continuously, along a distance equal to the diameter of the wire and 0.7...0.8 mm additionally. x

c) On completion of winding of the first layer of wire around the drum, there is ordered reversing of reel movement, by activating the other proportional electromagnet of the control directional valve which actuates the cylinder of the reel. Simultaneously, the drum rotational speed decreases, so that the tangential speed of the wire on the drum to be steady.

d) At each change of layer, the winding diameter is changed and, from the automation panel of electric motor operation, lower speed of the drum is ordered, so that the wire speed remains constant. Thus, for the first layer of wire wrapped with 9.5 mm diameter, the displacement speed of the reel is 1.8 m / min, and the drum rotational speed is 180 rev/min.

e) During the winding the drum is weighed; when the scale shows 2.1 ... 2.2 tons, order is made for the drive of the second drum, respectively for the displacement of the second reel to the standby position.

f) The wire on the wound drum is cut by means of a guillotine, then it is directed to the second drum, on which winding continues.

g) The duration of wire winding around the drum is approx. 60 minutes. In this period of winding time is sufficient for manually binding the wound coil, manual removal of a flange of the drum, by unscrewing a large nut, extraction of the coil from the drum, using a hydraulic device, its moving by means of the running bridge to the storage location, reassembling of the drum flange and preparing it for a new winding cycle.

h) Working is non-stop, in three shifts, and the number of wire rolling mills, each serviced by two drums, and the wire diameter are set according to the demands.

i) In case of failure in operation, poor winding quality and at the end of each coil, the wire is cut and directed to a chopper, until the fault is rectified, or winding is switched to the other drum.

9


j) The wire wastes resulting from chopper are melted down and sent back into the

manufacturing cycle.

Fig.4. Wire coil bound, taken from the drum and stored for delivery.

3. The electro hydraulic servomechanism for control of the reeling head

In the division at ALRO, each reel of a winding machine, which services a wire rolling mill, is driven by a hydraulic cylinder with bilateral fixed rod and mobile liner, powered by an adjustable pump, directly driven by the speed reducer. Along the supply lines of cylinder there are mounted adjustment valves and throttles, from which is made the setting of cylinder speed in both directions of travel. For each diameter of wire is made an adjustment of the pump. The actuating directional control valve of the cylinder is on/off type, not proportional. In these circumstances, one loses much time for adjustments, and the quality of wound coils is poor, which results in multiple wastes (which are re-melted). To test the proposed solution, in December 2012, at ALRO, INOE 2000-IHP Bucharest mounted on a drum of a winding machines an own system for operation of the reel, type electro hydraulic servomechanism, and on the other was kept the existing home solution. This system worked and compared with the classical solution proved to be more reliable (higher quality in winding wire, less wastes). From the original solution, existing at ALRO, there has been preserved only the hydraulic cylinder with fixed bilateral rod and mobile liner. The structure of this electro hydraulic servomechanism for control of the reeling head includes: The pumping unit, which has a modular design and comprises: an oil tank (with lid, filling opening and vent, level gauge and drain plug) on the lid of which is bolted a plate of hydraulic mechanical joints, inside which, by means of a grip mechanical coupling, is made coupling between the shaft of a fixed gear pump, mounted submerged in the tank, and the shaft of an electric motor, of constant speed, fitted to the top of the plate. Also inside this plate there are embedded a check valve, acting as anti-cavitation for the pump, a pressure control valve with direct control and a 2/2 normally closed directional control valve, which protects the pump at starting, not allowing it to start under load. In the proximity of the plate of mechanical hydraulic joints there is mounted an adapter plate, which distances from the electric motor the other modules of the pumping unit, respectively the plate incorporating the return filter and the mounting plate of the proportional directional control valve Dn6. On the mounting plate of the 4/3 proportional directional control valve, Parker production, there is screwed a glycerine pressure gauge with measuring range 0-100 bar.

10


Fig.5. The existing pumping unit and the shoe and drum brake of the reducer: there have been blocked the

pipes by which this group supplies the hydraulic cylinder of the reel; a speed transducer was mounted on the drum shaft of the reducer.

Fig.6. The modular pumping unit

fitted with proportional directional control valve (IHP).

Fig.7. Drive cylinder of the reel:

position and speed cable transducer (IHP).

The hydraulic cylinder with bilateral rod and double acting (existing subassembly) is connected by means of two 12x1 pipes to the consumers A and B on the mounting plate of the proportional directional control valve. Thermostatic control of hydraulic oil in the installation is provided by a system comprising the following components: a dedicated module of the electric and automation panel; a resistance for heating the oil, immersed in oil and attached to the front wall of the tank; a winding tube of copper 14x1 pipe, oil immersed, with the ends attached to the tank cover and a normally closed electro valve on the inlet of the cooling water in the winding tube; a temperature sensor mounted immersed in oil and attached to the tank cover. Automating the operation of the reel is made by the following procedure:

- on the existing hydraulic cylinder, actuating the reel, is mounted a position transducer, which allows precise monitoring and control of travel speed of hydraulic cylinder liner and instantaneous position of the liner, throughout the active displacement stroke; position and velocity control is carried out by means of an electro hydraulic control loop;

11


- drive of the hydraulic cylinder is carried out through the hydraulic proportional directional

control valve, which actuates the hydraulic cylinder in a displacement closed loop; - travel speed of the reeling head is v [mm/s] = d [mm] * n [rev/s] , where d is diameter of the

wire which is wound around the drum, and n is rotational speed of the winding machine drum (in the period required for winding one spire around the drum, the reeling head, thus the cylinder liner, performs a stroke equal to the wire diameter);

- the information on the drum rotational speed is taken from a variable reluctance speed transducer, mounted on the shaft of winding machine reducer;

- there has been used an TWIDO (Schneider Electric) programmable logic controller (PLC) equipped with an LCD console for monitoring and introduction of operating parameters of the reeling head (for example, the diameter of the wire to be wound around the drum). On customer request, the programmable logic controller (PLC) can be provided with a data communication line, type RS232/485, CANopen or Ethernet, with MODBUS communication protocol (RTU or ASCII).

The electric and automation board includes: the electrical power components, namely the contactors and the safety devices specific to the electric motor actuating the fixed flow pump and to the hydraulic oil thermostatic control system; ON / OFF switch; HAZARD WARNING mushroom push button; optical indicators; temperature controller; TWIDO programmable logic controller.

Fig.8. The electric and

automation board (IHP).

Fig.9. Wire connections at transducers, made on the inner surface of the board

case (IHP).

Fig.10. The outside of the board case (IHP): wire diameter selection keys; reel manual displacement control; operating mode switching control;

start; stop; signaling lamps.

4. Schematic diagram and technical characteristics of the product

The schematic diagram, fig.11, refers only to the operation of a single reel, but it can be also extended to two reels, keeping the same hydraulic unit to power both hydraulic cylinders, as they work almost continuously in sequence and only a short sequence at the same time (at the end of winding around a drum and preparation of winding around the other drum). This schematic diagram can be materialized for all 7 double winding machines of wire rolling mills at ALRO and is suitable for both the introduction of the reel operating parameters and the monitoring of the production of wire. The structure of the diagram in Figure 11 is as follows: 1- oil tank, fitted with tight cover, level indicator, filling and ventilation opening, drain plug, with Vmax.= 30 l and Veffective=22 l; 2- oil heating resistor, with N=1330 W; 3- winding tube for cooling the water inside the copper 14x1 pipe; 4- normally closed electro valve, G1/2”, ∆pmax= 6 bar, supply voltage 24 V DC; 5- temperature sensor Pt 100; 6- gear pump, capacity 3.65 cm3/rev, maximum pressure 250 bar, inlet filter; 7- electric motor, 220 V, single phase, 0.55 kW, 1400 rev/min; 8- grip mechanical coupling; 9- pressure control valve 0...60 bar; 10- check valve, opening pressure 1 bar; 11- 2/2 hydraulic distributor, normally closed, operated electrically and manually, which protects the pump at starting. 12- intermediate plate with role in fastening and mounting the pumping unit; 13- return filter, filtration fineness 10µm, equipped with bypass valve; 14- pressure gauge wit glycerine, measurement range 0...100 bar; 15- proportional hydraulic directional control valve,

12


4/3, rD6, closed center; 16- (existing) hydraulic cylinder, bilateral fixed rod and mobile liner, Øpiston = 65 mm; Ørod = 48 mm; Stroke = 870 mm; 17- cable position transducer, for monitoring and control of hydraulic cylinder speed; 18- speed transducer of winding machine drum; 19- (existing) end of stroke signaling devices, controlling the change of the direction of hydraulic cylinder displacement.

Fig.11. Schematic diagram of the electro hydraulic servomechanism for driving the reeling head

13


5. Description and operation of the electro hydraulic servomechanism driving the reeling head

At the level of the pumping unit, mechanical energy from the electric motor shaft (the wedge of speed and torque) is converted into hydraulic energy (the wedge of flow and pressure) materialized by the oil flow, routed from the pump discharge pipe to the hydraulic cylinder. The amount of oil pressure, depending on the load of the hydraulic cylinder, can be adjusted by manual operation of the pressure valve. When starting the pumping unit there is operated automatically (or manually), for 4-5 seconds, the hydraulic 2/2 directional control valve, normally closed, which closes the flow path to the hydraulic cylinder and allows free discharge (no load) of the pump to the tank. At the level of hydraulic proportional 4/3 directional control valve, closed center, on the center position, without drive, the holes P (pressure, from pump), T (tank), A and B (consumers, that are connected to the hydraulic cylinder) are closed (not communicating with each other). When supplying one of the two coils of directional control valve, its slide valve moves, left or right, with a stroke proportional to the intensity of the supply current of the coil, and connects P to A, respectively B to T, for a direction of travel of the hydraulic cylinder or P to B, respectively A to T, for the opposite direction of travel. The distribution system of the hydraulic oil, to the hydraulic cylinder, ensures: proportionality between actuation of proportional directional control valve in current and speed of movement of the hydraulic cylinder, change of the direction of movement of the hydraulic cylinder, starting and stopping of the hydraulic cylinder. The closed loop of electro hydraulic control, achieved by means of the speed transducer, displacement transducer and the PLC, controls the travel speed of the hydraulic cylinder by maintaining the linear static characteristic current-flow of the proportional directional control valve in a constant and sufficiently small deviation range. At the level of the hydraulic oil thermostatic control system, when the temperature sensor detects values lower than 38oC, there is ordered starting of supply of the electrical resistance, and at values of the oil temperature higher than 42oC there is ordered opening of the electro valve for access of cooling water into the winding tube of copper pipe.

Conclusions

• Tests conducted during putting the product into service, as well as its subsequent exploitation, over a period of 3 consecutive months, have clearly demonstrated the advantages of this type of drive, command and control of the reeling head, compared to the existing solution: higher quality of wire winding around the drum; decrease of scraps of badly wrapped wire, that is re-melted; thermostatic control of the hydraulic oil temperature; the possibility to set the diameter of wire that is to be wrapped; on customer request, the possibility to monitor, record and control the production of wire through a data communication line.

• The only drawback of the product is related to random blocking of the slide valve of proportional directional control valve, due to impurities from the environment, which can not be prevented from entering the hydraulic oil. The solution for removing this obstacle is to take three steps: installing a filter on the hole P of the proportional directional control valve; an increase of 10 bar of the maximum pressure adjusted (from 30 bar to 40 bar); choosing a higher gear electric driving motor.

REFERENCES

[1] INOE 2000-IHP Bucharest, S.C. ALRO S.A. Slatina, “Centrală Hidraulică pentru Bobinatorul Laminorului de sârmă“ (Hydraulic plant intended for the bobbin winder of a wire rolling mill), Contract No. 4600007204/30.08.2012, Construction documentation, 2012

14


RANS SIMULATION OF COMBINED FLOW AND HEAT TRANSFER THROUGH OPEN-CELL ALUMINUM FOAM HEAT SINK

assistant professor Petre OPRIŢOIU*

*Technical University of Cluj-Napoca, Departament of Infrastructure, str. Observatorului, nr.72-74, 400363, Cluj-Napoca, Romania, E-mail: [email protected]

Abstract Frequently, advanced electronics, optics, nuclear equipment and high frequency microwaves

systems require cooling of some devices at a heat flux of about 5-30 MW/m2. To meet this demand the porous medium of the heat exchangers has to be compressed thus the spherical particles are distorted and agglomerated.

The aim of this research is to study by simulation the effect of open cell aluminum foam on the heat transfer and pressure drop in cooling devices at the high heat fluxes. Heat transfer and pressure drop in an open foam heat exchanger, made of aluminum of different porosity (ε) and porous density (PPI), cooled by water were investigated numerically using CFD code Fluent and the results are presented.

Maximum fluid flow velocity used was 1.2m/s. The permeability (K) and form coefficient (cF) varied from 2.52×10-10 m2 and 1276 m-1 to 3.44×10-9 m2 and 4731 m-1, respectively. It was determined that the flow rate range influenced these calculated parameters.

Heat flux (q) up to 1.38 MW/m2 was removed by using porous sample with porosity 60% (ε=0.608) and average pore diameter (dp) 2.3mm. Under this simulating condition, the difference between the temperature of the wall and the bulk water did not exceed 63⁰C.

An estimate of heat sink efficiency using compressed aluminum foam for cooling high-power electronic devices was done. The results obtained in this study are relevant to engineering applications employing metal foams ranging from convection heat sinks to filters and flow straightening devices.

Keywords: simulation, pressure drop, flow characteristics, heat transfer performance.

1. Introduction

The problem of dissipating high heat fluxes has received much attention due to its importance in applications such as open-cell aluminum foam heat exchanger, cooling of electronic equipment.

The most effective way of cooling is pumping liquid inside these devices through michrochannels of porous medium. The effects of fluid velocity, particle diameter, type of porous medium and fluid properties on the heat enhancement were investigated.

However the estimated maximal values of the heat flux dissipated does not exceed 1.38MW/m2, at pressure drop 8.22bar and at a velocity 1.2m/s in a single phase water flow. Most of the theoretical models and numerical simulations used pore diameter (dp) as one of the basic parameters for calculating both heat transfer and friction in the porous medium of cooling systems.

Experimental studies on the effect of compression and pore size variations on the liquid flow characteristics and heat transfer has been performed by Boomsma and Poulikakos [1].

They showed the compressed open-cell aluminum foam heat exchangers had thermal resistance (Rt) that were two or three times lower than the best commercially available heat exchanger, with the same pumping power (Ẇ).

15


2. Theory

Compression in the y direction is parallel to the airflow. Compression in the x direction is transverse to the airflow. The highest heat transfer coefficient (αeff.) was obtained with compression in the x or y directions. However, the highest pressure drop (ΔP) was obtained with compression in the x direction. High-pressure efficiency was obtained with both y, and equal x and y, compression. Compression in the y direction yields high heat transfer and moderate pressure drop. Equal x and y compression results in lower heat transfer and lower pressure drop, compared with y compression [2].

In this case a model for numerical simulation and practical calculation needs somewhat other approaches and require an extension of experimental base to provide the necessary background.

Open-cell aluminum foam was used as a porous medium in the model of heat sink. Applied porous media is manufactured from special grades of atomized metal powders. The initial powder particle size controls the pore size and distribution when sintered to a specified density.

The permeability (K) is related to the pore size (dp) and pore distribution (PPI). Material properties such as thermal conductivity, thermal expansion and density are highly dependent on the porosity and generally decrease as porosity increases.

Table 1 gives an overview of the physical properties of all foams which were tested by simulation. 40 PPI foam, of two different initial porosities, one of 92% and the other of 95% were compressed by various factors from two to eight.

The first two digits of the foam’s name designate the porosity of the foam in pre-compressed form. The second pair of numbers of the foam name signify the compression factor.

Table1. Compressed foam physical data(A) and uncompressed foam physical data(B)[1].

Foam Compression Name Expected porosity [%]

Effective porosity [%]

Panel A

5% 2 95–02 90.0 88.2 4 95–04 80.0 80.5 6 95–06 70.0 68.9 8 95–08 60.0 60.8 8% 2 92–02 84.0 87.4 3 92–03 76.0 82.5 6 92–06 52.0 66.9

Panel B

Foam Pore diameter [mm]

Ligament diameter

[mm]

Specific surface area [m2/m3]

Effective porosity [%]

10 PPI 6.9 0.40 820 92.1 20 PPI 3.6 0.35 1700 92.0 40 PPI 2.3 0.20 2700 92.8

To measure the actual values of the porosity, each compressed foam block was weighed,

each compressed foam block was weighed, and based on the external measurements, an effective porosity was calculated and compared to an expected final porosity based on the foam’s initial solid fraction initial and compression factor.

The expected porosity was based on the simple physical relation for a change in volume, where M is the compression factor (ratio of the original uncompressed foam block height to the final compressed height) and ε is the void fraction of the material(0<ε≤1).

16


εcompressd = 1– M(1 – εuncompressed) (1)

Figure 1 shows the expected porosity of the compressed foams blocks as lines with the actual porosity measurements represented as points.

Fig. 1. The expected compressed metal foams porosities based on the precompression porosity and nominal compression factor compared against measured values [1].

The porosity of each block was calculated by dividing its weight by the volume, as measured by the external dimensions, and then comparing this value to the density of the solid material, aluminum 6101.

The surface aria to the volume ratio (specific surface aria) is also tabulated for the uncompressed metal foam blocks in Table 1. This specific surface area data were provided by the foam manufacturer [3].

Porous media model is nothing more than an added momentum sink in governing momentum equations. Since the volume blockage that is present physically is represented in the model, FLUENT uses and reports a superficial velocity inside the porous medium, based on the volumetric flow rate, to ensure continuity of the velocity vectors across the porous medium interface. The following properties are required [4]:

(a) Porosity (ε); (b) Viscous resistance (1/K), for aluminum foam:

pdK ×=ε

π321 (2)

where K - permeability,[m2] and dp - pore diameter,[m];

(c) Inertial resistance (cF)

LKLP

cF ρ

µ

2v

v2

−∆

= (3).

17


The theoretical pressure drop per unit length for porous media was predicted following

Forchheimer equation (1901):

2vv ⋅+⋅=∆ ρµ

Kc

KLP F

(4)

where ∆p/L - pressure drop per unit length, [Pa/m]; μ - fluid viscosity, [kg/m∙s];K - permeability, [m2]; v - velocity, [m/s]; cF - innertial coefficient, [m-1]; ρ - fluid density, [kg/m3].

3. Results and discussion

3.1 Heat transfer performance

The final overall dimensions of the compressed foam blocks used in pressure-drop and heat transfer simulations were 240mm×100mm×100mm, with the cross-sectional area normal to the flow direction measuring 240mm×100mm.

To make them functional heat exchanger, each foam was brazed in a central position to an adjoining heat spreader plate made by solid aluminum.

A typical flow and heat transfer configuration is shown in fig. 2. A heat source is bonded or joined to a thin conductive substrate on which a block of open-cell aluminum foam of length L and thickness W is attached.

The foam is then placed in a channel, and cooling fluid of velocity u0 at a temperature T∞ is pumped through the open celled material, thereby affecting heat transfer from the hot source to the cooling fluid.

Fig. 2. Schematic of geometrical model as used in simulation [5] and contours of fluxes.

The heat transfer rate to the coolant q is defined by the following energy balance in eq. (5)

)( inout TTcmq −=

(5)

where Tin, Tout – the inlet and outlet temperature of the liquid, [K]; ṁ - mass flow rate, [kg/s] and c - specific heat, [j/kg∙s]. Fig. 3 shows the value of the heat flux q, [MW/m2] remove by the heat

18


exchanger of 10 cm thickness, depending on the velocity of the coolant. We can see that dissipation of heat flux up to 1.38MW/m2 was obtained by simulation.

Fig. 3. Heat flux plotted against flow velocity

Under this simulating condition, the difference between the temperature of the wall and the bulk water did not exceed 63⁰C. A practical measure of the performance of a heat exchanging device is the dimensionless Nusselt number (Nu) as given in eq. (6).

f

lpor dNu

λα ⋅

= (6)

where αpor, [W/m2K] is the convection heat transfer coefficient, which caracterizes the heat transfer between a solid and a fluid; λf, [W/mK] is thermal conductivity of the coolant and dl, [m] is ligament diameter. The Nusselt numbers were calculated at various coolant flow rates and plotted against the coolant flow speed in fig. 4.

Fig. 4. Nusselt number plotted against flow velocity.

The heat transfer from the foam to the fluid will increase as either the porosity decreased (thus increasing surface area for heat transfer) or as the relative density (ρ%) increases (thus increasing heat conduction through the ligaments) or as the velocity of the fluid increases.

The localized heat transfer coefficient Nusselt (Nu) will vary with velocity, even at the low Reynolds number. This is primarily due to a physical phenomena caused by the tortuous nature of the porous flow.

00.20.40.60.8

11.21.41.6

0 0.5 1 1.5

q [M

W/m

2 ]

v [m/s]

q ε=0.608

q ε=0.669

q ε=0.805

q ε=0.825

q ε=0.874

q ε=0.882

0

200

400

600

800

1000

1200

0 0.5 1 1.5

Nu

[-]

v [m/s]

Nu, ε=0.669

Nu, ε=0.825

Nu, ε=0.874

Nu, ε=0.608

Nu, ε=0.805

Nu, ε=0.882

19


At the openings between the cells, where the fluid passes from one cell to another, fluid

passes through a contraction and expansion. At the back side of the opening (expansion side) the fluid develops eddies and vorticies. These vorticies will affect the mixing and local boundary layers, thereby affecting localized heat transfer from the cell walls to the fluid.

Heat transfer are commonly characterized by the Colburn factor (j), which gives a heat transfer performance estimate comparing the convection coefficient to the required flow rate of a heat exchanger. The Colburn factor is given in eq. (7) [7].

3/2ν

v

=

acj por

ρα

(7)

where ν is the kinematic viscosity and a is the fluid thermal diffusivity. Fig. 5 show the value of the Colburn factor depending on the Reynolds number.

Fig. 5. Colburn factor plotted against permeability-based Reynolds number

The values of Colburn factor are highest for foam 95-08 with ε=0.608, at very low values of ReK. The heat sink 95-02, with ε=0.882 shows lower magnitude of Colburn factor, but its observed for extended range of ReK. Clearly that the metal foam with verry high initial porosity, even compressed, yields low Colburn values.

In any heat exchanger design, the heat convection performance of the heat exchanger must be weighed against the energy required to operate the system, which is the pumping power in this configuration. The required pumping power Ẇ, [w] was calculated for the aluminum foam heat exchanger at various coolant flow velocities, according to eq. (8).

vQ⋅∆= PW (8)

where ΔP is the pressure drop across the aluminum foam heat exchager and Qv, [m3/s] is the volumetric flow rate of the coolant passing through the heat exchanger. Also, a common means to measure the heat convection effectiveness is the thermal resistance Rt, [K/w] as shown in eq. (9).

)( inout

inplt TTcm

TTqTR

−−

=∆

=

(9)

where Tpl – average wall temperature. Lower thermal resistance facilitates the heat flow through heat exchanger (fig. 6).

0

500

1000

1500

0 5 10 15 20

j [-]

Re[-]

j 92-06

j 92-03

j 92-02

j 95-08

j 95-04

j 95-02

20


Fig. 6. Plot of the required pumping power against the corresponding thermal resistance.

In fig. 6 the optimal design is that which minimize the distance from the point to the origin of the plot. This point was obtain by foam 92-06 with porosity ε=0.669 and for a thermal resistance of 0.061K/W. The worst performance was generated by 95-02 with porosity ε=0.882 and for a thermal resistance of 0.199K/W. The metal foam heat exchangers decreased thermal resistance by nearly half when compared to currently used heat exchangers designed for the same application.

3.2 Pressure and flow characteristics

The amount of work required to pump the coolant through a heat exchanger is a critical heat exchanger design parameter. The parameters used to describe the pressure drop characteristics of the foam heat exchangers are the permeability and form coefficient which are defined in equation (4).

All data were calculated and reported on a Darcian flow velocity basis. This velocity accounts only for the channel dimensions, its independent of the porosity of the test material, and is practical for comparison against other sets of porous media. The range for the water velocities were from 0.25 to 1.2m/s and for pressure changes were from 0.113 to 8.226 bar.

Figure 7 shows the pressure drops simulating data and the fitted curves in graphical form for the compressed blocks based on velocity.

Fig. 7. Pressure drop versus fluid velocity for compressed foams and contours of static pressure.

0100020003000400050006000700080009000

0 0.05 0.1 0.15 0.2 0.25

Ẇ [W

]

Rt [K/W]

Ẇ, ε=0.669

Ẇ, ε=0.825

Ẇ, ε=0.874

Ẇ, ε=0.608

Ẇ, ε=0.805

Ẇ, ε=0.882

0

2

4

6

8

10

0 0.5 1 1.5

ΔP [b

ar]

V [m/s]

92-06, ε=0.669 92-03, ε=0.825

92-02, ε=0.874 95-08, ε=0.608

95-04, ε=0.805 95-02, ε=0.882

21


As seen in fig. 7 compression has profound effect on the pressure-drop behavior of

compressed foam. As expected, those foams which possess the highest solid fraction 95-08 (lowest ε) as seen in Table 1, generated the largest pressure drop. The foam which produced the lowest pressure drop was foam 95-02, which was also the most porous of the samples.

For a more general base of comparison, the hydraulic characteristics of the heat exchangers can be viewed using non-dimensional flow factors, like Reynolds number based on permeability (ReK).

The characteristic length is replaced by the square root of permeability, as shown in eq. (10), where ρ is the density of the fluid, v is the Darcian flow velocity, and μ is the dynamic viscosity of the fluid.

µρ KvRe =K (10)

The other commonly used non-dimensional flow describing factor is the Fanning friction

factor (f) which is given in (eq.11). This provide information concerning the required pressure drop (ΔP) across a heat exchanger and come into use when the heat transfer performance to cost ratio is considered.

∆=

2v

42ρ

hDL

Pf (11)

In eq. (11), the hydraulic diameter (Dh) is described by eq. (12):

p

Ch L

AD 4= (12)

with AC being the cross-sectional area of the flow channel and Lp being the wetted perimeter of the flow channel. Fig. 8 plots the friction factor of eq.(11) against the velocity. Refering to the friction factor the pressure drop of the foam is dominated by the form coefficient of eq. (4).

Fig. 8. Friction factor plotted against darcian velocity and contours of skin friction coefficient.

0

0.5

1

1.5

2

2.5

3

3.5

0 0.5 1 1.5

f [-]

v [m/s]

f, ε=0.669

f, ε=0.825

f, ε=0.874

f, ε=0.608

f, ε=0.805

f, ε=0.882

22


Permeability and form coefficient were calculated for each block using the entire flow rate

range tested for each foam block (Table 2).

Table 2 Flow characteristics and associated pressure drop

Both foam samples series which were 95% and 92% porous before compression show similar flow behavior with the respect to the changes in the compression factor. For the 95% original porosity series, increasing the compression factor from two to four reduced the permeability from 3.44E-09m2 to 6.87E-10m2.

The other series of compressed foam blocks showed approximately the same sensitivity between the compression factor and the change in permeability.

Fig. 9 shows a plot of the permeability based on the measured porosity of the compressed metal foam samples. There is no difference made in the plotting data points between foams of 95% and 92% pre-compression porosity; all are placed on the same scale by their measured porosity in compressed form.

Fig. 9. The permeability of compressed foams is plotted against the values of measured porosity.

The form coefficient also varied with the compression of the metal foam blocks and the differing pre-compression porosities, ultimately being controlled by the porosity of the compressed metal foam. The form coefficient of the foams increased monotonically with decreasing porosity (fig. 10).

Foam ε[-] K[m2] cF[1/m] ΔP[bar]95-08 0.608 2.52E-10 4731 1.149-8.22692-06 0.669 3.95E-10 3399 0.180-2.80095-04 0.805 6.87E-10 2957 0.461-3.90592-03 0.825 8.26E-10 2820 0.393-3.50392-02 0.874 3.08E-09 1472 0.128-1.45795-02 0.882 3.44E-09 1276 0.113-1.274

0.00E+00

5.00E-10

1.00E-09

1.50E-09

2.00E-09

2.50E-09

3.00E-09

3.50E-09

4.00E-09

0 0.2 0.4 0.6 0.8 1

K[m

²]

ε[─]

95-08

92-06

95-04

92-03

92-02

95-02

23


Fig. 10. The form coefficient of compressed foams is plotted against the darcian velocity. The three aluminum foam blocks which were tested were of nearly the same porosity. The

only difference between the samples was the average pore diameter. Referring to Table1, the porosities of these aluminum foam blocks ranged from 92% to 92.8%, and the pore diameter varied from an average of 6.9mm to 2.3mm. The difference in pore diameter appeared to dramatically affect the permeability and the form coefficient of the foams.

Decreasing the pore diameter (dp), decreased the permeability and increased the form coefficient. The 10 PPI foam (95-02), which had a pore size of 6.9mm, generated the least flow resistance. In contrast, the 40 PPI foam (92-06) with a pore size of 2.3mm, had the greatest flow resistance. The increase of flow resistance directly relates to the effective surface length as explained by Lage [6], which relates an increase in drag to the increase in the specific surface area.

That values for the permeability and form coefficient of the porous medium depend upon the flow velocity range over which are they are calculated [7]. The permeability and form coefficient were calculated for each foam by varying the flow velocity range which the terms were calculated to investigate this dependence [1]. Fig. 11 plots the permeability based on an increasing flow speed. The permeabilities of the three uncompressed foams are nearly constant.

Fig. 11. Permeability plotted against darcian velocity

0

1000

2000

3000

4000

5000

0 0.5 1 1.5

c F[m

-1]

v [m/s]

cF 92-06

cF 92-03

cF 92-02

cF 95-08

cF 95-04

cF 95-02

0.00E+005.00E-101.00E-091.50E-092.00E-092.50E-093.00E-093.50E-094.00E-09

0 0.5 1 1.5

K [m

²]

v [m/s]

K 92-06

K 92-03

K 92-02

K 95-08

K 95-04

K 95-02

24


4. Conclusions about flow and heat transfer through open-cell foam

Open-cell aluminum foam compressed by various factors and then fashioned into heat exchangers for cooling electronic devices can dissipate large amounts of heats. Various heat exchanger evaluation methods were applied to the data, which included the hydraulic characterization, the heat transfer performance and an efficiency study to determine the most efficient metal foam heat exchanger configuration.

The compressed aluminum foam performed well not only in the heat transfer enhancement, but they also made a significant improvement in the efficiency over several commercially available heat exchangers which operate under nearly identical condition [8].

Open-cell aluminum foam were numerically tested to evaluate their hydraulic characteristics using water. The modules consisted of open-cell aluminum foams of various porosities and pore diameters.

The characterization procedure involved solving for two terms, the permeability and the form coefficient. These two factors accurately described the pressure-drop versus flow velocity behavior in porous media in general and were shown to be applicable to high porosity metal foams. From these simulations several conclusions can be drawn:

- The structural differences in the precompressed form did not a noticeable effect on the permeability. When comparing compressed foams with varying degrees of compression and initial porosities, the post-compression porosity governs the permeability and the resulting pressure drop.

- Increasing the compression factor decreased the permeability by regular amounts, which were nearly equal for of the two foam series (92 and 95%).

- The permeability of the compressed foams became more sensitive to changes in the porosity as the porosity increased.

- Holding the porosity constant and decreasing the pore diameter increased the flow resistance by reducing the permeability and increasing the form coefficient. This increase is attributed to the higher specific surface area generated by the smaller pore size.

- Using different flow velocity regimes resulted in various permeability and form coefficient values. Whenever the permeability and the associated form coefficient for a high-porosity porous medium are stated, the flow velocity range over which these terms are calculated must also be specified for accuracy.

REFERENCES

[1] Boomsma K. and Poulikakos D., 2002, The Effects of compresion and Pore Size variations on the liquid Flow characteristics in metal foams, ASME J. Fluids Eng., 124, pp. 263-272. [2] John F. Klein, Noe Arcas, George W. Gilchrist, William L. Shields, Jr., Richard Yurman, and James B. Whiteside, 2005, Thermal Management of Airborne Early Warning and Electronic Warfare Systems Using Foam Metal Fins, Technology Review Journal. [3] Duocel Aluminum Foam Data Sheet, 1999, ERG Material and Aerospace, Oakland. [4] Opritoiu, P., 2007, Fluid flow and pressure drop simulation in aluminium foam heat exchanger, Acta Tehnica Napocensis, nr.50. [5] Nihad, D., Ruben, P., and Alvarez, H., 2006, Heat transfer analysis in metal foams with low-conductivity fluids, ASME J. of Heat Transfer, 128, pp. 784-789. [6] Lage J.L., 1998, The Fundamental Theory of flow through permeable media from Darcy to turbulence, Transport Phenomena in Porous Media, Elsevier Science, Oxford, pp.1-30. [7] Antohe, B. V., Lage, J. L., Price, D. C., and Weber, R. M., 1997, Experimental Determination of Permeability and Inertia Coefficients of Mechanically Compressed Aluminium Porous Matrices, ASME J. Fluids Eng., 119, pp. 404-412. [8] Boomsma, K., Poulikakos, D., Zwick, F., 2003, Metal foams as compact high performance heat exchangers. Swiss Federal Institute of Technology, Zurich, Switzerland.

25


ANALYSIS SOLAR RADIATION

Rusănescu Carmen Otilia1, Rusănescu Marin2, Dorel STOICA1

1 University Polytechnic Bucharest, Biotechnical Faculty of Engineering, [email protected] 2 Valplast Industrie Bucharest Abstract: In this paper is monitored the intensity of solar radiation by the meteorological station in the year 2012 and is calculated: intensity of diffuse radiation, direct radiation intensity, declination angle, hour angle . The the weather station type: AWS / EV is a product born from the need to frequently monitor the environment variables. The usage of the suitable mathematical algorithm makes possible to accurately follow the movement of the sun. Keywords: global, direct and diffuse radiation, declination angle hour angle.

1. Introduction Sun is the largest object in the solar system containing 98% of its mass. He is a ball of incandescent gas mass from which we get heat and light. It has a diameter of 1.391 million kilometers which means it is 109 times greater than Earth Radiation, the most important agent of heat in the atmosphere plays a major role in the processes that occur at medium and large scale. Radiation appears as a genetic element of the climate on a global scale [3]. Even under clear sky radiation that reaches the earth's surface in all directions from the diffusion phenomena, known as diffuse radiation, is 5 ... 15% of the flux of solar radiation that reaches the Earth's surface without being affected by this phenomenon, known as direct radiation. Together, direct and diffuse radiation, the so-called total radiation.

2. Materials and methods

Determination of the Sun-Earth angle (angle of declination, zenith angle, solar azimuth) makes it possible to determine the position of collector of solar radiation from the sun so that its efficiency is maximized. Based on mathematical algorithm, we determine the values of these angles for the period January to June 2012, and the minimum and maximum period.

The efficiency of a solar collector (the heating panel or PV) can be significantly increased if the collector is located under the sun so that the angle of incidence (angle between the radius of the sun and the line perpendicular to the collector) becomes zero or very small. Implementation of this requirement involves modeling the Sun-Earth angle, which must be accurate, relatively simple.

Global solar radiation intensity G horizontally was monitored weather station: AWS / EV Biotechnical Faculty of Engineering, Geco MICROS SIAP program version 2.3.2 software automatically records the following parameters: air temperature, wind direction and speed, atmospheric humidity, solar radiation, rainfall.

The weather station is wireless transmission range up to 300 m and the set of sensors integrated pillar of 1.77m and tripod for. [2]. Solar radiation sensor is manufactured in accordance with international specifications WMO (World Meteorological Organization).

It consists of a transducer which is heated in proportion to the incident solar radiation, absorbed by a special layer of black paint of the measuring surface of the heat. Double layer shielding of special optical glass to optimize the characteristics of the measurement under different environmental conditions.

26



This transducer is included in the family of smart sensors, as it is equipped with a

microprocessor that performs multiple functions: checking the operation right, data preprocessing, A / D conversion to electrical signals, etc..

These features will ensure excellent accuracy and high reliability of data. The protection is made of aluminum alloy corrosion, shield UV-resistant plastic material with a low thermal capacity. Internal circuits are protected from atmospheric discharges and polarity reversal. This is an analog sensor output signal between 0 and 2 VDC. Privacy Framework is a aluminum alloy corrosion, UV resistant plastic with low thermal capacity. Internal circuits are protected from atmospheric changes and polarity inversion. This is an analog sensor output signal having a range from 0 V to 2 V [6]. Measurement from 0-1300 W/m2. Sensitivity of 1.5 mV / W / m 2 Accuracy + / - 10 W / m2 Resolution + / - 0.5 W/m2. Linearity: + / - 1% Operating Temperature -30 to 600C Output signal: 0 V (0 W/m2) at 2 V (1300 W/m2) Sensor connector 4 pin female Mounting: with support (mast), the position is important because it must be pointed south.

3. Results and discussion

Based on recorded global radiation intensity, we calculate the direct and diffuse components of solar radiation. Based on 24 hour weather station record of 24 in 2012, we assumed diffuse radiation intensity equal to one fifth of global radiation intensity and the intensity of direct radiation is the difference between global and diffuse [8]. According to equation (1), D - is the intensity of scattered radiation; G - global or total radiation intensity B - Direct radiation intensity

DGBGD −== ,5

(1)

hour angle: determines the position of the Sun in the sky at a given moment. Is 0 when the sun passes the local meridian corresponding point of the sensor location. This angle is positive to the east (to the east) and negative to the west (at dusk). Within an hour the sun across the sky at an angle of 15 °, and the position of the clock (T) is determined by the relationship:

( )T−⋅= 1215ω (2) If you are known angle of declination, latitude and hour angle can be determined by calculating the position of the Sun in the sky the sun height angle and solar azimuth angle, applying the above calculation formulas. The angle between the direction to the sun the place of capture and the equatorial plane is called declination δ. Relations for calculating the angle of declination δ are:

+

⋅⋅=365

284360sin45,23 nδ [1] (3)

27


−

=365

)80(360sin45,23 nδ [8] (4)

where n is the day of the year in which the measurements were taken. In Figure 1 is the angle of declination.

Figure 1 Representation of establishing the position angle sun in the sky [4]

Based on recorded global radiation intensity, we calculated the direct and diffuse components of solar radiation. Figure 2 shows the proportion of diffuse radiation intensities and intensity of direct radiation in global radiation. It is interesting to note that the intensity of diffuse radiation has a high intensity compared with direct radiation. Based on 24 hour weather station record of 24 in 2012, I assumed diffuse radiation intensity equal to one fifth of global radiation intensity and the intensity of direct radiation is the difference between global and diffuse. In Figure 3, we present the variation of global radiation, diffuse and direct weather station recorded in May 2012. Based on the measured and calculated values of global radiation, diffuse, direct from January 2012 to June 2012 we plotted Figure 4. It is noted that the high value of global radiation was recorded on May 8 at 14, having a value of 879, the 10th of June at 14 was recorded the highest value of global radiation during January 2012-June 2012 which is the 899 [W / m²]. Some authors have performed statistical analyzes on different materials. [2,5].

Figure 2 The variation in global radiation, direct and diffuse for January-June 2012

28


0100200300400500600700800900

5/1/2012

5/3/2012

5/5/2012

5/7/2012

5/9/2012

5/11/2012

5/13/2012

5/15/2012

5/17/2012

5/19/2012

5/21/2012

5/23/2012

5/25/2012

5/27/2012

5/29/2012

5/31/2012Sol

ar r

adia

tion

inte

nsity

[W

/m2]

G

D

B

Figure 3 Intensity variation of global radiation, direct and diffuse recorded by the weather station in May 2012 [8]

Figure 4 Correlation between components: global direct diffuse solar radiation during January-June 2012 [8]

If you are known angle of declination, latitude and hour angle can be determined by calculating the position of the Sun in the sky the sun height angle and solar azimuth angle, applying the above calculation formulas. According to figures 5 and 6 the angle of declination is dependent on the day they were made in solar radiation measurements. In Figure 5 are the values of the angle of declination in May 2012 and in Figure 6 are the minimum and maximum declination angle based on statistical analysis in 2012. From Figure 7 is observed, as shown in the literature [1] that the values are positive hour angle morning and afternoon negative.

29


-4

-3

-2

-1

0

1

2

3

5/1/2012

5/3/2012

5/5/2012

5/7/2012

5/9/2012

5/11/2012

5/13/2012

5/15/2012

5/17/2012

5/19/2012

5/21/2012

5/23/2012

5/25/2012

5/27/2012

5/29/2012

5/31/2012

May, 2012

ω

Figure 5 Declination angle variation by day of n calculated in May 2012 [8]

Figure 6 Variation of minimum and maximum angle of declination in January-June 2012

30


Figure 7 The variation angle zone in May 2012 [8]

Conclusions Using mathematical algorithms presented in this paper to determine the Sun-Earth angle (angle of declination, zenith angle, hour angle), makes it possible to determine the position of collector of solar radiation from the sun so that its efficiency is maximized. Based on mathematical algorithm, we determined the values of these angles for 2012 and we represented graphically angles for July month was recorded maximum solar radiation and minimum and maximum values for the entire year. The results of this study to monitor solar radiation, allowing interpretations that may be used to establish local potential use of solar energy. To complete this study, it takes more time to monitor solar radiation. Solar energy is the gateway to a new era, with its use in heating, resulting in reduction of environmental pollution. The efficiency of a solar collector (the heating panel or PV) can be significantly increased if the collector is located under the sun so that the angle of incidence (angle between the radius of the sun and the line perpendicular to the collector) becomes zero or very small. Implementation of this requirement involves modeling the Sun-Earth angle, which must be accurate, relatively simple.

REFERENCES

[1] Goswami, D.J., Kreith, K., Kreider, J.F.: Principiile ingineriei solare, Philadelphia, PA, George H. Buchanan Co., 1999. [2]Irina Istrate, Diana Cocârță, Silvia Neamțu, Talida Cirlioru - The assessment of an alternative treatment for PCB polluted soils in the romanian context - bench scale tests, 2012, Water air and soil pollution, vol 224 (4),DOI: 10.1007/s11270-013-1516-2, ISSN: 0049-6979 (print version), ISSN: 1573-2932; [3].Cristian Oprea – Radiatia solara, Aspecte teoretice si practice Bucuresti 2005 [4]. Daniel Tudor COTFAS, Petru A. COTFAS, Corina COJOCARIU, Lucian COSTINESCU, Cornel SAMOILA,Solar Tracker cu algoritm matematic - Conferinţa Naţională de Instrumentaţie virtuală, Ediţia a V-a, Bucureşti, 20 mai 2008 [5] D. Stoica, G Stanciu: “Influence the degree of sorting the separation process a conical sieve” Digest Journal of Nanomaterials and Biostructures Vol. 8, No. 2, April - June 2013, p. 513 - 518 [6].Manual de instalare al Statiei Meteo AWWS/EV firma Elettronica Veneta; [7]. Messenger, R., Ventre, J.: Sistemul fotovoltaic de inginerie, Boca Raton, London, New York, Washington, CRC Press, 2000. [8]Rusanescu Carmen Otilia – Indrumar de meteorologie si climatologie, Editura Matrix Rom, Bucuresti, 2013

31


REPAIRING AND TESTING OF THE HYDRAULIC SERVO VALVES Radu RADOI1, Ioan BALAN2, Iulian DUTU3

1 INOE 2000 - IHP, [email protected] 2 INOE 2000 - IHP, [email protected] 3 INOE 2000 - IHP, [email protected]

Abstract: (Arial, 11pt, Italic, Justify) The equipment which must work at high values of speed and response frequencies and

reach optimum performance need to have their hydraulic installations equipped with servo valves. The costs of production of the servo valves are very high, due to their mechanical complexity and they are not compatible at all with the contaminated fluids, cause of the small gaps between the parts in motion and of the very small apertures of nozzles. If these installations are not properly maintained servo valves can be damaged very quickly and can only be remedied in centers (laboratories) by specialized personnel. The paper presents ways of identifying a faulty servo valves, symptoms and possible causes and their testing.

Keywords: hydraulic, servovalve, testing, fault

1. Introduction

The equipment which must work at high values of speed and response frequencies and reach optimum performance need to have their hydraulic installations equipped with servo valves. The modern servo valves are very reliable components, due to the major improvements applied to them in the course of time. The costs of production of the servo valves are very high, due to their mechanical complexity and they are not compatible at all with the contaminated fluids, cause of the small lost motions between the parts in motion and of the very small apertures. These very performant equipment are therefore suitable for sophisticated systems and require a rigurous maintenance.

These prerequisites regarding precision and dynamics of the hydraulic drives lead to the use of hydraulic devices of control with high performances. The dynamic performances of the servo valves with direct command, are limited at 80 Hz and of those with piloting stage at about 100 Hz. Beside the influence of the hydrodinamic forces, the performances of the hydraulic devices are much influenced by the dynamics of the electromechanic actuators, which drive, directly or indirectly the slide.

The close loop drives have been utilized a lot lately. In a drive of this kind with close loop or with servo control, the servo valves are the key elements, with the greatest impact upon the static and dynamic features of the drive system.

The servo valves must provide high fluid flows and dynamic features, high rigidity of the system at load and very slight deviations of position of the operating element, from the prescribed position.

For reaching this must be accomplished the following: - High degree of accuracy at the execution of the bushand servo valve slide, for obtaining

a symmetric and linear signal flow function - Reduction of the mass of the parts in motion

32


- Reduction of the interference forces which have a negative impact upon the slide

dynamics friction forces, hydrostatic forces, hydrodynamic forces and impulse forces - Reduction of power losses of the piloted servo valves and minimization of the control

chambers volumes - Improvement of the dynamics of the electromechanic convertor.

2. The description of the servo valves

Due to the afferent performances, the most widely spred servo valves are those with two stages see fig.1 These type of servo valves have beside the compulsory pilot, a stage of hydraulic amplification, directional type. The pilot which comprises the couple motor and the nozzle flap amplifier may be considered and sometimes it is even utilized as a servo valve with an amplification stage. This is a classic directional valve with marks executed in a higher precision class. The symbol of this type of servo valve may be seen in fig.2

1 - torque motor; 2 - hydraulic amplifier; 3 - control spool; 4 - coil; 5 - armature; 6 - torque tube; 7 - flapper plate; 8 - control orifices; 9 - feedback spring.

Fig. 1 Rexroth servo valve type 4WS2EM

Fig. 2 Servo valve symbol

The servo valves are commanded with electric signal applied to the spools of the couple motor. The two spools may be connected serially or in parallel. With a few exceptions, the servo valves use only c.c.signal. The hydraulic installations which enclose servo valves must take into account some very strict conditions:

• The installed flow of the pump must be 10% higher than the highest flow supplied by the servo valve

33


• The servo valves must be protected with a filter of 5 + 15 um This filter is recommended to

be put as an individual protection for each servo valve. It is recommended that the safety filter of the servo valve to be deprived of by pass valve

• The pressure valve which maintains pressure almost constant on the input in the valve must be chosen in such a waz that for the range of the flows discharged in the basin during work the variation of the adjusted pressure to be below 10 bar

• It is recommended that the hydraulic cylinders which equip the hydraulic installation to be with bilateral rod of the same dyameter. This ensures a good stability allowing also an easier and more accurate calculation of the hydraulic system and also bring simplification of the electronic control system.

• It is recommended that servo valves to be placed as close as possible to hydraulic motors they serve. Is indicated as links of servo valve with hydraulic motor to be rigid (pipes) or not to vary their volume at pressure variations.

• It is recommended that between pumping group and servo valve be mounted a hydropneumatic accumulator whose size is determined by the pump flow and the operating mode of the system. It is necessary to provide manual controls to bring hydraulic motors in the original starting position.

• It is recommended that the hydraulic system be equipped with oil cooling - heating system to ensure a temperature of 45 to 55 °C during operation.

3. Finding faults in installations with servo valves

If it is found that the system that comprises a servo valves is not working, for testing can be replaced with another spare servo valves, and if you do not have a backup servovalve will proceed as follows:

• Shall be measured the command signal of servovalve thus checking the electronic system operation:

- If electronic block does not give electronic signal to needed parameters search fault here and fix - If electronic block give control signal and if one can not vary manual the signal it must be used an external signal generator. - If the load does not move (at external command signal) put gauges on the supply and return ports of servo valve. If the supply pressure is good that means the pressure source is good and then continue searching fault.

• Place pressure gauges on motor ports. Vary the control signal and check output pressure at motor ports.

- If we have not a gradual increase in pressure means that the servovalve is damaged.

Faulty servovalves must be removed from a system only after the place around them it is cleaned. For servovalve removal first disconnects electrical connector, remove the screws with which it is fixed and mount the replacement servo valve whose protective cover was removed previously. Finally, the protective cover is mounted on demounted servo valve, to prevent dirt entering in the joints and losing O-rings.

To see if servovalves can be repaired these should be sent to specialized laboratories operated by highly qualified personnel .

4. Repairing servovalves

The symptoms of main defect encountered in servovalves can be:

34


• The servovalve give flow only one way and the electric control gives no results (nozzle

clogged) • Servovalve not respond to commands (broken coil, clogged nozzles); • Presence of flow without giving an electric command, flow which decreases at electric

command in a certain sense (shifted null); • Unequal flow at equal control level for both polarities (asymmetry) • large hysteresis on reversing electrical control (friction between the spool and sleeve due to

impurities) • High flow at null that can not be canceled by adjutments (high wear of spool and sleeve

edges).

Removal and installation of servo valves should be done only by specialized personnel and only based on accurate and complete instructions. The place where is made removal and instalation must be perfect clean. Because of the size of the nozzles and small gaps between moving parts any impurity can cause blockage of servovalve. Dismantled parts shall be placed on a non-metallic surface. On removal is well to note (mark the relative positions of the parts). Spool must be handled with care not to damage the edges. He should not be placed on hard surfaces. It is advisable to avoid removing the nozzle if it leads to change their position and if does not have the possibility of resettlement at fixed odds. If they are removed have to avoid damaging them, especially if there does not exist conditions to recalibrate them. After cleaning and repair servovalve component parts are assembled at place, then appropriate adjustments will be made and will be drawn static characteristic.

5. Testing of servovalves

Testing of servovalves is made on specialized stands. Such a stand for servo valves testing (Fig. 3) consists of: a tank on which is fixed a plate for connecting servovalve in testing circuit, a pumping group, pressure and flow transducers, a servocontroler and a computer equipped with data acquisition board.

Testing is as follows:

1) Check that stand and equipment fitted on it corresponds with mounting scheme; 2) Adjust supply pressure to servovalve at nominal value 3) The input current is passed several times through the circuit 4) The testing application is opened on the computer and then start the testing, after that the program generating the increasing steps of control signal and flow given by the servovalve is recorded and after draw the characteristic diagram 5) Check that the machine pressure remains relatively constant throughout the current cycle 6) periodic signal applied continuously allows recording characteristics during a complete cycle

35


Fig. 3 Servo valve mounted on stand for testing

Periodic signal during a complete cycle ± Imax [mA] (maximum control current in both directions) applied to servo valve for plotting the characteristic is generated by a software application (Fig. 4). The application allows recording the data for further processing. The application generates the control signal and carries signals from the transducers through a data acquisition board.

.

Fig. 4 Panel of the testing application made in TestPoint medium

From a recorded chart can be seen if hysteresis is large, if the parameters stipulated in the data sheet (maximum output at maximum control signal) are touched and if the chart is asymmetric or if nulll is shifted.

6. Terms of putting into service the installation with a new instaled servo valve

Before mounting a servo valve in the system is necessary, especially in complex installations that had a very long operating since last oil change, to make a washing of a hydraulic installation. For this the servo valve is replaced with a special adapter plate for washing, on which is mounted an electrically controlled directional valve that can provide equivalent flow with that one given by the servo valve.

In some cases instead of servo valve is mounted a simple plate that directly connect the tank port with inlet pressure port. All pumps are started at low pressures (with safety valves opened, checking not to appear control signals to servo valves). Adjust safety valves at the values set by the designer.

36


Are given commands to the directional valve who replaced the servo valve to achieve the

rated machine cycle speed of hydraulic motors. Check at the same time outward leakage appearance. Aftyer a while it stops the operation of the installation and replace the washer plate with the new servo valve. Are gived to the servo valve electrical signals covering all the field of control. Give commands in automatic mode making the final adjustments at the control electronics as well as the hydraulic. It is well that all adjustments to be made at recommended hydraulic medium temperature and designer prescription which is in domain 45 to 50° C. The null of servo valve should be adjusted according to actual conditions of automatically cycle (supply pressure, back pressure and oil temperature). Verification of correct adjustment of null is made by cutting electric connection of servo valve, in which case if the electronics is well adjusted, operated element must remain in place. After 3 hours filter cartridges are replaced with new ones.

7. Conclusions

Servo valves are high performance equipment and are suitable for sophisticated installations and systems.

Servo valves are sensitive devices and their use must meet a number of strict conditions such as:

- proper filter of oil to keep them in working order; - respecting the installation instructions for servo valves can be achieved maximum of

performance. Repair of servo valves must be performed only by qualified personnel, otherwise these can

be irreversibly affected.

REFERENCES

[1] Cornel Velescu, Aparate si echipamente hidraulice proportionale, Mirton publishing house , Timisoara, 2003 [2] M. Comes, P. Drumea, A. Mirea, G. Matache, Intelligent servohydraulic device for the control of the motion – 24th International Spring Seminar on Electronics Technology ISSE 2001 [3] M. Comes, A. Drumea, A. Mirea, I Enache. , Electronic module for servohydraulic system with frequency control, MTM 2001, Romania [4] Dutu, R.I. Radoi, M. Blejan “Digital control module developed for a servohydraulic positioning system”

Caciulata, Romania; 7-9 November, 2011, “Proceedings - HERVEX”, ISSN 1454-8003; pp.381-385 [5] http://www.moog.com/literature/ICD/G761_CDS6673_D.pdf

37

http://www.moog.com/literature/ICD/G761_CDS6673_D.pdf


RESEARCH ON VARIATION OF DISPLACEMENTS, VELOCITIES AND ACCELERATIONS AT A SITE SELECTOR BLOCKS (FANNER) GRAIN

STOICA DOREL1 , RUSANESCU CARMEN OTILIA1

1 , UniversityPOLITEHNICA of Bucharest, Faculty of Biotechnical Engineering,

[email protected]

Abstract: The researches aimed to study and establish technologies vibratory phenomena involves the tools and machines for processing agricultural products type wheat. The data recorded in the experiments were statistically designed a series of graphs showing the correlation between the outcome and the functional characteristics accompanied by reports of correlation. Functions developed have allowed us to draw a number of conclusions nature of generalization. Keywords: displacement, velocity, acceleration, selection, seed.

1. Introduction

The material obtained after harvesting with combine in the form of a mixture of primary culture seeds, grains of other crops, seeds of weeds, and various impurities (scrap straw, chaff, dust, sand, etc.), and the seed culture besides the main cover normally developed seeds, seeds dry, shriveled or broken, and others. [1].

After the harvest, agricultural products (grains, fruits and vegetables) can not be directly used for various purposes such as: storage, consumption, industrialization, commercialization, seed material, since contain impurities (plant debris and other objects) and products injured. Products harvested before receiving a particular destination is necessary and required to undergo cleaning and sorting operations.

Through these operations [2], aims to increase the purity of the product, whilst achieving best storage conditions, and a reduction in the transport and storage.

For operations of cleaning and sorting of seeds obtained after harvest, usually with combinations, using special complex equipment specific to this domain [1,2,3].

In these machines, an important place flat site blocks whose operation is based on the vibratory movement of the working surfaces. These are used to perform the separation of mixtures of grains which differs by one of the two geometrical dimensions of their thickness or width.

Other authors in various works conducted various statistical analyzes, [5]. There are other methods and principles of separation of impurities in seed mass, which

may be based on other operating principles, such as using air currents, tables densimeters separation after elastic properties, etc..[4].

Studies on the angle of the dials was conducted by the authors of paper [6].

2. Material and methods

In order to conduct experimental research needed for the thesis using winter wheat (variety Flamura 85).

Experiments were performed in laboratory experimental bench, resulting in adaptation Mechanical Vânturătorii VM - 4 existing laboratory of Agricultural Machinery Agricultural Mechanization team of University of Agricultural Sciences and Veterinary Medicine Bucharest.

The experimental stand (Figure 1) is designed for cleaning and sorting by size and by the aerodynamic properties of seeds of cereals, pulses, industrial crops, grasses obtained from or combine threshers.

38


It is used for cleaning and grading seed crop following wheat, rye, barley, oats, rice, peas,

beans, lentils, flax, mustard, hemp, sunflower, millet, rape.

Figure 1. Stade experimental Part of the stand experimentall: 1 – food cart; 2 – Superior support frame site; 3 – First sieve the upper bearing pedestal; 4 – Second Screening upper bearing pedestal; 5 – fan; 6 – shutter to evacuate chaff; 7 – shutter to evacuate chaff; 8 – first sieve the lower bearing pedestal; 9 – the second screen of the lower bearing pedestal; 10 – framework; 11 – site backup box; 12 - small impurities trough drain and leaks; 13 – a seed discharge chute topping; 14 – Electric motor drive; 15 - Seed discharge chute second rate; 16 – chute to evacuate large impurities.

For experimental determinations was used as a chain composed of the following devices:

1) - like National Instruments data acquisition with the following features: 24-bit resolution, sampling rate of 50 kS / s analog input 4 channel simultaneous, dynamic range 102 dB input range + / - 5 V, USB 2.0 interface for PC connectivity 2) - four accelerometers Brüel & Kjær 4508B with magnetic fastening and metal clip, each connecting cable with the following characteristics: - Description: top connector TEDS, sensitivity 100 mV / g, frequency range limits 10% (± 1dB): 0.3 to 8000 Hz resonance frequency: 25 kHz, the residual noise level: 0.35 mg operating temperature range: -54 100 deg C measuring range: 70 g; maximum level of shock: 5000 g weight: 4.8 grams connector: 10-32 UNF, mounting: magnetic and metal clip, connection cable length acquisition board: 5 m interface cable connection to acquisition board: BNC. 3) - computer software Labview data acquisition and processing;

In order to achieve the four measurements used accelerometers were located in pairs diametrically opposite the center of the grid being able to determine vibrations both in the tangential direction and the radial direction.

Two accelerometers are able to determine the parameters of the vibrations in the radial direction, while the other two accelerometers are methods of determining the vibration parameters for the tangential direction.

39


Measurements were made both idle and for driving the load in two directions, both in the

radial direction and the tangential direction as. The four accelerometers were connected to the data acquisition board via a computer with

printer for plotting the acquired signals. Before each sample were adjusted accordingly cinematic parameters namely oscillating

sieve oscillation frequency or amplitude of oscillation. For driving the load and oilseed rape have been used, measurements are made at about the same flow conditions. It was tried to be kept constant in all experimental measurements, that is around 0.01 kg / s.

Acquisition time was several seconds (8-12 seconds), the signals are acquired by taking on only those corresponding to a relatively uniform movements work site.

Signal acquisition was made through LabVIEW, data acquisition was performed before program structure by which the purchase was made and signal processing.

For plotting the acquired signals can also use a printer that can be connected to laptop computer.

Variation of displacement, velocity and acceleration of the upper block of the site in relation to time under load and at idle is achieved in the graphs in Figures 2 ... 13.

Changes in driving higher building site in relation to the time course load for the angle of the upper sieve higher block site ( 3=α , 6=α , 9=α ), and the angle of the lower grid block below the site 5=β is shown in figures 2-5.

Superior building site speed variation against time to walk to load the hopper angle higher upper block site ( 3=α , 6=α , 9=α )and the angle of the lower grid block below the site

5=β is shown in Figures 6-8 and in Figure 9 is shown the upper block with site speed range at idle.

Fig. 2. Variation of displacement with time α=3 si β=5

0.4

0.5

0.6

0.7

0.8

50 53 56 59 62 65 68 71 74 77 80

time(s)

disp

lace

men

t (m

m)

xant

x

xurm


0.2

1.2

2.2

3.2

4.2

5.2

6.2

7.2

42 45 48 51 54 57 60 63

time(s)

disp

lace

men

t (m

m)

xurm

x


0.1

0.3

0.5

0.7

0.9

52 53 54 55 56 57 58 59 60 time(s)

disp

lace

men

t (m

m)

xurm

x

xant

Fig.5. Variation of displacement with time idling

1.4

1.5

1.6

1.7

1.8

50 51 52 53 54 55 56 57 58 59 60 time(s)

disp

lace

men

t (m

m)

16

18

20

22

24

26

28

30

ocity

(m/s

)

vant

v

vurm

12

14

16

18

20

42 45 48 51 54 57 60 63

Vel

ocity

(m/s

)

vurm

v

40


Superior building site acceleration variation with respect to time in the course load for the

angle of the upper sieve higher building site ( 3=α , 6=α , 9=α ) and the angle of the lower grid block below the site 5=β is shown in Figures 10-12 and Figure 13 presents the variation in acceleration higher block site at idle.

Fig.8. Velocity variation with time α=9 si β=5

7

14

21

28

52 53 54 55 56 57 58 59 60 time(s)

Vel

ocity

(m/s

)

vurm

v

vant

Fig.9. Speed variation with time idling

7.5

7.6

7.7

7.8

7.9

8

8.1

50 51 52 53 54 55 56 57 58 59 60 time(s)

Vel

ocity

(m/s

)

Fig.10. Acceleration variation with time α=3 si β=5

4.5

5

5.5

6

6.5

50 53 56 59 62 65 68 71 74 77 80

time(s)

Acc

eler

atio

n (m

/s2)

a

aurm

Fig. 11. Acceleration variation with time α=6 si β=5

4

5

6

7

42 45 48 51 54 57 60 63 time(s)

Acc

eler

atio

n (m

/s2)

aurm

a aant

Fig.12. Acceleration variation with time α=9 si β=5

1.3

2.3

3.3

4.3

5.3

6.3

52 53 54 55 56 57 58 59 60 time(s)

Acc

eler

atio

n (m

/s2)

aurm a

aant

Fig.13. Variation of acceleration with time idling

1.5

1.55

1.6

1.65

1.7

50 51 52 53 54 55 56 57 58 59 60

time(s)

acce

lera

tie(m

/s2)

41


Conclusions

The process of separating the holes of the seed hopper is achieved by providing a relative movement of the seed from the hopper. This relative movement to provide, by printing a grid linear oscillatory motion of a drive mechanism (connecting rod - crank, eccentric). Vibratory movement is applied in practice in various fields including agriculture And food industry, both for the transport of granular and powder products, or even in the form of pieces, as well as processes to achieve the separation, or to remove impurities from the mixture seed or plant products sort by size. Separators is done in the form of block printing mechanisms driven oscillating movement screening material found on the surface of separation. The mechanisms that generate oscillations can be: crank, with backdrop oscillating vibrating devices unbalanced unbalanced rotating or vibrating devices electromagnets etc.. From the dynamic point of view, the separator is a vibrator vibrating with one or more oscillating masses linked to the support and to each other by elastic elements (made of metal or rubber) and a drive system (drive) to ensure the generation of disruptive forces necessary for a stable oscillating. Working body vibration machine (the block making) has a generally translational movement, linear or circular, depending on the type of vibration generator. The vibration generators used to operate the oscillating blocks are present as well, the rotating unbalanced mass, which results in a force directed interference (one-way). REFERENCES

[1] Căsăndroiu T.,Voicu Gh. – Curba de separare a materialului pe lungimea sitei superioare la sistemul de curăţire al combinelor de cereale, Bucureşti, 1992; [2] Voicu Gh., Stoica D., Ungureanu N. - Influence of oscillation frequency of a sieve on the screening process for a conical sieve with oscillatory circular motion, lucrare publicata în Journal of Agricultural Science and Technology, ISSN 1939-1250, USA June. 2011, Volume 5, No.2 (Serial No.27) [3] Elfverson, C., Regnér, S., Comparative precision of grain sieving and pneumatic classification on a single kernel level, Applied Engineering in Agriculture, p.537-541, Vol. 16(5), 2000; [4] Stoica D., - Contribuţii la studiul fenomenelor vibratorii privind utilajele din domeniul prelucrării produselor agricole (teza de doctorat, septembrie 2011) [5] Rusanescu, Carmen-Otilia. "DETERMINATION OF SUGARS IN RED AND WHITE ROMANIAN WINE SAMPLES." METALURGIA INTERNATIONAL 18.4 (2013): 131-133. [6] Constantin POPA, Mihaela-Florentina DUȚU, Iulian DUŢU - 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, - Revista Hidraulica, 2013.

42


NUMERICAL SIMULATION OF FLOW IN ANASTOMOTIC COMPLEX AFTER GASTRIC RESECTION WITH GASTROJEJUNAL ANASTOMOSIS

O. Vaida¹, L. Vaida², A. Andercou³.

¹ Department of General Surgery, Municipal Hospital Dej, Romania ² Department of Mechanical Engineering of the Faculty of Mechanical Engineering of the Technical University of Cluj-Napoca ³ University of Medicine and Pharmacy „Iuliu Hatieganu”, II Surgical Clinic, Cluj-Napoca, Romania

Abstract Aim: To investigate the flow by numerical simulation in mechanical hydraulic models that approximates the stomach and gastric resection procedures Billroth I (BI) and Billroth II (BII). Method: Are used geometries axially symmetric non-deformable. Calculations obtain hydrodinamic sizes for all configurations of interest: normal geometry and geometries that simulate gastric resection procedures BI and BII. Results: By calculating the modulus of resistance, M and of time of emptying it has been found an acceptable flow for procedures BI (Péan), Y Roux (Y-R) and gastric resection Leger (GRL). In techniques BII, Reichel-Polya (R-P), Hoffmeister-Finsterer (H-F) the flow is far away from that of the normal geometry. Conclusions: Operations drastically alter the flow in accordance with the extent of anatomical changes.

Keywords: Numerical simulation, Gastric resection, Anastomosis.

Introduction Motor function of the stomach and small intestine is essential for digestion of food. It is met of smooth muscles of these digestive segments with the participation of intramural nerve plexus Meissner and Auerbach. Stomach muscles are divided into three layers: external, longitudinal, middle, circular and internal with oblique fibers. The intestine has one external muscular layer and another internal circular. Gastric and intestinal smooth muscle contraction is coordinated by a complex nervous and endocrine mechanism that controls integrated and other functions: secretory of digestion and absorption. Stomach motor function is initiated by smooth muscle cells, by the contraction what triggers „migratory motor complex” (MMC). This is carried out in four phases and repeat cyclically to 90 minutes, producing peristaltic waves (1). ,,Flow” to the stomach is materialized by the their sequence and has the following consequences: gastric filling, fragmentation and mixing foods with gastric juice and finally discharge of chim into the duodenum through the pyloric sphincter. Pyloric sphincter along with the lower esophageal sphincter are specialized formations who acts synergistically in regulating ,,impulses and outputs” from the stomach. Pyloric sphincter by adjusting his tonus, opposes of discharge into the duodenum to remaining unprocessed food particles in the stomach and stops duodenogastric reflux (2). To explore the digestive segments have at hand various parameters like: intraluminal pressure, time and velocity of the exhaust, resistance to flowing, etc. There are numerous studies that can investigate anatomically intact gastrointestinal tract, but there are few those explore the digestive tract after surgery. In these cases impediments of explore are due to anatomical changes achieved trough operation on the one hand and on the other through the difficulties to find appropriate method of investigation whose results reflect with accuracy the functionality. Known two main types of digestive reconstruction after gastrectomy: Billroth I (BI) with gastro-duodenal anastomosis and Billroth II (BII) with gastro-jejunal anastomosis.

43


Depending on the type of intestinal section for manufacturing anastomosis, processes BII can be: BII by longitudinal enterotomie (LE) and BII by transverse enterotomie (TE). The procedures BII most used are: Reichel-Polya (R-P) Hoffmeister-Finsterer (H-F) through LE and the fitting in Y à la Roux (Y-R) (fig.1).

A. B.

Fig. 1. Procedures BII through LE

A. Reichel-Polya. B. Hoffmeister-Finsterer . 1 - gastric stump; 2- duodenal stump; 3 - afferent loop; 4- efferent loop.

The method through TE was launched in 1950 by L. Leger gastric resection Leger (GRL) (fig 2).

A. B.

Fig. 2. The fitting Y-R and GRL throughTE

A. The fitting Y Roux: 1- gastro-jejunal anastomosis; 2- gastric stump; 3- jejuno-jejunal anastomosis B. Gastric resection Leger: 1- gastrojejunal anastomosis; 2-afferent loop; 3- efferent loop.

44


In terms of peristalsis the BII anastomosis are isoperistaltic when the stomach and jejunum have the same peristalsis direction and antiperistaltic when the stomach and afferent loop have the opposite direction of peristalsis (3). Through gastro-jejunal anastomosis is born gastro-jejunal anastomotic complex (AC). AC is a structure resulting by uniting of anastomotic partners responsible by postoperative disordes. AC is different in procedures BII especially according on the type of enterotomy used, LE or TE. Whatever type of enterotomy in gastrojejunal anastomosis AC has the following components: afferent loop (AL), anastomotic area or moth anastomosis and efferent loop (EL). After LE anastomotic area is wide, gastric slope being the whole gastric trance (R-P) or half of it (H-F) and intestinal slope is jejunum incised longitudinally on distance corresponding to trance. AL and EL communicate each with anastomotic space, trough two openings, for entry (AL) and output (EL) distant from each other to the small and to the great curvature of the gastric stump. After TE anastomosis is to the great curvature of the gastric stump, with diameter of 3-4 cm. AL and EL are in continuity and delimited endoluminal of a parietal spur. They have a common opening toward the gastric anastomotic slope. Gastrojejunal anastomosis by ET resembles with one end-to-end. Numerical simulation can explore the functionality of AC but has limits related rigidity mechanical-hidraulic models that approximates the stomach and surgical procedures. It is a research stage which can be followed by the development of experimental models or experience on animals if deemed necessary.

Working hypothesis For numerical simulation of flow are necessary hydraulic mechanical models which reproduce approximately the stomach and anastomotic montages achieved by the techniques of digestive reconstruction after gastrectomy. Flow from the stomach intact into the intestine is the reference point to assess AC with specific structure for each technique of gastric resection. Are to be sought the differences of flow which exist between techniques BI and BII and that differ by the type of anastomosis used, gastroduodenal (BI) or gastrojejunal (BII). Is important to mention that in the techniques BII duodenum is excluded from gastro-intestinal circuit, and that the AC architecture is amended depending by type of enterotomy (LE or TE) for anastomosis.

Material and method Simulation stomach evacuation From point of view hydrodynamic gastric evacuation can be represented like a non-permanent and multiphase movement which takes place in a complex field, deformable. The flow is generated by two factors:

- pressure difference between the stomach and duodenum; - peristaltic movements of the walls of these organs.

Between the stomach and duodenum is located pyloric sphincter. He has a smooth circular musculature with annular thickenings, which under physiological conditions works like a unidirectional mechanical valve (fig.3).

45


Fig.3.The stomach and duodenum (7-pyloric sphincter; arrows indicate the direction of the outlet of the

stomach)

Measurable parameters that have relevance for the hydrodynamic study are: - the pressure in the stomach: p1 - pressure difference Δp (fig.3)

Δp is the size of dynamic most important which is based mathematical modeling of the flow from the stomach to the duodenum. Local balance between Δp and elastic tension of pyloric musculature provide its normal operation. For a normal function manometric relative pressure pm registered at the opening of pylorus is between 80 - 120 mm H2O, respectively 784,532 - 1176,78 Pa (1 mmH2O = 80665 Pa), a value that is induced by the muscle tension of pylorus. Pressure values outside this range can mean disfunctions of the pylorus. For the hydrodynamic calculation, pressures values Δp and pm are the fundamental size without which cannot obtain relevant quantitative results, simulation being limited at a qualitative description of the phenomenon of gastro-duodenal flow. Was created a first 2 D model that can study gastric emptying and who later allow qualitative assessments about it, in case of surgical procedures that change constructive this model. Proposed model for the analysis of hydrodynamics in gastric emptying uses an axial-symmetric geometry non-deformable (fig.4). Stomach is considered a reservoir with variable section, and duodenum and small intestine, two rigid sections in series, while large intestine has the rol of an adjusting tap. The main objective of using this model is to get credible numerical solutions through movement created in a rigid area at the ends of which applies constant differences of pressure. Pylorus is considered a component of local resistance of the segment which simulates duodenum, in direct relation with the size pressure variation between stomach and duodenum (fig.4).

46


Fig. 4. Geometry used to simulate evacuation stomach

Numerical simulations describe completely, hydrodynamics movements analyzed for a Newtonian fluid (water). Calculations allow obtaining of hydrodynamic size, for all configurations of interest: normal geometry and the geometries that approximates gastric resections procedures. BI (Péan) and BII (R-P, H-F, Y-R, GRL). The analysis will be made for low Reynolds numbers of the movement because the evacuation of the stomach is a very slow process. It will analyze the escape of liquid for a given model aiming to determine the time of emptying, in which the level of water in gastric reservoir, reaches from baseline h1 at one final h2 where h2 < h1 (4, 6) Such analysis starts from the fact that the movement cannot be considered permanent because hydraulic parameters change in time. For solving, the movement is considered as a succession of permanent movements, that take place at time intervals elementary, and total time will be obtained by summing the elementary time. It is considered the general case of a reservoir which empties free through a pipe, whose hydraulic resistance module is M. It is to be determined separately at all the variants analyzed (fig.4). Flow is clearly made from reservoir toward evacuation. By choosing of a the reference plan in accordance with figure 4 gives:

22

11 h

gph

gp

+⋅

⟩+⋅ ρρ

(1)

Will fallow the time in which, free surface in the reservoir reaches from the value h1 at h2. The Energy Law (Bernoulli’s equation) between points A and B where occurs movement it can be written in accordance with the relationship from Figure 4 as follows:

( )2

21

22

11 1

2Q

hAgMh

gph

gp

⋅

⋅−=

+

⋅−

+

⋅α

ρρ (2)

Or:

( ) 2* QhMHg

p⋅=+

⋅∆ρ

(3)

Where: ( )21 ppp −=∆ is the static pressure difference between points A and B;

47


21 hhH −= = is the height of position between points A and B;

( )hM * = is the global module of hydraulic resistance which includes and the kinetic terms;

Q = the exhaust flow

Because h2 is constant in the case of elementary variation can write: dhdh =1 (4)

Taking account of the continuity ecuation (mass conservation law) results: ( ) dhhAdtQ ⋅−=⋅ 1 (5)

Elementary time variation is:

( )Q

dhhAdt ⋅−= 1 (6)

If gon replace in the relation 6 flow expression (Q) resulting of 3 is obtained:

( ) ( )

dhh

gp

hMhAdt ⋅

+⋅∆

⋅=

ρ

*1 (7)

By the integration result of the reservoir depletion time T between the initial difference H (existing at the moment t=0) and final difference (at the moment t=T):

( ) ( )

dhh

gp

hMhAT

H

⋅+

⋅∆

⋅= ∫

∗

0

1

ρ

(8)

For explaining overall module of resistance will apply energy law, between sections 1-1 and 2-2 of the considered system. Under this law one can write:

212

2222

11

211

22 −++

⋅+

⋅⋅

=+⋅

+⋅⋅

phhg

pgvh

gp

gv

ρα

ρα

(9)

Where: - α1 si α2 are coefficients by nonuniformity of speed (Coriolis coefficient). These coefficients take into account uneven distribution of speed in the normal section studied. For these coefficients will be adopted the value α1 = α2 = 2, appropriate of laminar flow in ducts with circular section:

- the terms gv⋅⋅

2

2α, represents the kinetic energy reported at weight

- the terms hg

p+

⋅ρ represents the potential energy reported at weight

- 21−ph represents loos of total hydraulic load between sections 1-1 and 2-2. To note

that these are dissipation of energy that can be found in the form of temperature increases in the fluid the moving. Means that loss of total hydraulic load is the ratio of the flux of mechanical energy dissipated between the two sections of the a current of fluids and the product ρgQ where p is density of the fluid, g is gravitational acceleration and Q is the volume flow. The loss of total hydraulic load

21−ph is determined by summing losses of load uniformly distributed, hl and of locale losses hloc. For a circular pipe of diameter d and length l along which there is a number n of irregularities (disturbing elements like: narrowing or widening of the section, elbows, bends etc.) the loss of hydraulic load it write:

∑=

+=−

n

iloclp i

hhh1

21 (10)

Taking account of relationships for loss uniformly distributed and for local losses which is expressed by the relations:

48


gv

dlhl ⋅⋅⋅=2

2

λ and g

vhloc ⋅=

2

2

ς (11)

where : λ is coefficient of the linear loss distributed; l – the length of pipe; d – diameter of the pipe; v – fluid velocity in the pipe; ζ – coefficient of local loss; and also the equation of continuity: AvQ ⋅= (12) Is obtained for value of miscarriages, expression

21

2

1 221 AgQ

dlh

n

iip ⋅⋅⋅

+⋅= ∑

=−

ςλ , (13)

and for value of resistance modulus

211 2

1Agd

lMn

ii ⋅⋅⋅

+⋅= ∑

=

ςλ (14)

Medical literature contains numerous dates about gastric resection, but in the majority of cases, the studies are limited at quantifying of the extirpations, and efficiency of the operation of extirpation. The main parameter that take it in consideration is the pressure drop in the area of intervention. For shaping stomach evacuation is necessary following average values of physical sizes which influencing it:

• the exhaust flow Q = 8,83 ⋅10-6 m3/s; • the normal capacity of the stomach V0 = 1,2 ⋅10-3 m3; • the average diameter of the duodenum dd = 3 ⋅10-2 m; • the average diameter of small intestine ds = 3 ⋅10-2 m; • the density of fluid chyme ρ = 1000 kg/m3; • dynamic viscosity of fluid chyme η ≈4 mPa ⋅ s;

• the average speed v = ≈⋅ 2

4dQ

π0.0125 m/s, which corresponds to a Reynolds

number of about 120; • stomach pressure p1 = (0,7 ÷ 1) ⋅103 Pa.

Results In the case of gastric resection shall be amended the conditions of flow downstream of the stomach. They influence resistance modulus M. In table 1 are presented his values, calculated for each type of gastric resection. Also are measured the emptying time t in all anastomotic montages for a volume equivalent to 1/3 of the volume of the stomach (gastric stump volume remaining after resection).

49


Table 1

Type of gastric resection The main scheme equivalent

Modulus of resistance

[s2/m5]

Time emptying [s]

Stomach in normal

operation

2.83*109 28.7

Gastric resection

Péan

1.9*107 2.35

Gastric resection

Reichel-Polya

1.1*107 1.72

50


Gastric resection

Hoffmeister-Finsterer

1.2*107 1.8

The fitting Y Roux

1.7*107 2.23

Gastric resection

Leger

2.71*107 2.81

Discussion

51


After gastrectomy digestive reconstruction with gastrojejunal anastomosis (GJA) gives rise

at anastomotic complex (AC). He is the new section digestive of flow resulted through surgical act, which changes substantially the area. The flow in AC can be explored by calculating the modulus of resistance M, a value that shall be amended, depending on the elements of local and linear geometry, with significance for energy loss at the fluid passing through the system (4). Towards normal anatomical configuration interfere local and linear resistance factors, that influence modulus M at all procedures of gastrectomy used. Here are few details that affect the flow in the system: suppression of pyloric sphincter, the exclusion from circuit of the duodenum; length of the afferent loop (AL): izo- or antiperistaltic montage; intestinal section type (longitudinal or transverse). Other factors of resistance may be: parietal spur, anastomotic loop bends, imposed by technique or accidental; the multitude layers of digestive suture; deficiencies by fixation of anastomosis at mesocolic breach; adherence syndrome postoperatively that can deform AC. Watching results from table 1 shows the most acceptable flowing in gastric resection BI (Péan), in technique on loop in Y Roux (Y-R) and after gastric resection Leger (GRL). In the procedures BI and Y-R, anastomosis of gastric stump with duodenum or jejunum is done end-to-end. In GRL although it is a suture gastrojejunal end-to-side due to cross-enterotomy behave as one end-to-end. Over time termino-terminal anastomoses can arrive sphincters comparable with the pylorus (5). The farthest flowing from normal anatomy is observed after techniques BII (R-P) and (H-F). Local resistance elements from these techniques, reflected by modulus M and the time of emptying are: long AL; hipofunctional segment between AL and EL, obliquity of mouth anastomosis; longitudinal section of the jejunal wall (interruption circular muscular fibers), sutures in two plans, antiperistaltic gastrojejunal montage etc.

Conclusions The surgical intervention alters substantially the flow regardless of the process used after gastrectomy by changing the local resistances. The flow is acceptable after techniques that change as little the digestive tract (section flow) and doesn’t affect its structure. Deficient flow regime even if the simulation was done on models with rigid walls, it may reflected the postoperative disorders.

References 1. Gheorghe C. – Fiziologia gastrică din Tratat de chirurgie vol.VIII, Partea I B - Chirurgie Generală sub red. Popescu I. – Editura Academiei Române – Bucureşti 2008; 1299-1301. 2. Hăulică I. Fiziologie Umană- ediţia a III-a Editura Medicală, Bucureşti 2009; 484-94. 3. Filopovic N, Cvetkovic A, Isailovic V, Matovic Z, Rosic M, Kojic M. – Computer simulation of flow and mixing at the duodenal stump after gastric resection – World J. Gastroenterol, 2009 Apr. 28; 15(160: 1990-8. 4. Opruţa D, Vaida L. – Dinamica fluidelor – Editura Mediamira, Cluj-Napoca 2004; 161-170. 5. Popovici Z, Borcean Gh. – Rezecţia gastrică tip Leger reprezintă oare un progres? Revista Română de Chirurgie nr. 2, martie-aprilie 1986. 6. . Idelcic I. I., Indrumător pentru calculul rezistenţelor hidraulice, Ed. Tehnică, Bucureşti, 1984

52


TESTING OF LINEAR PNEUMATIC ACTUATORS

WITH HYDRAULIC LOAD Gabriela MATACHE1, Stefan ALEXANDRESCU1, Gheorghe SOVAIALA1, Ioan PAVEL1, Iulian-Cezar GIRLEANU1

1INOE 2000– IHP Bucharest, e-mail:[email protected]

Abstract: Actuators are execution elements used in automated mechatronic drives, which generate the useful mechanical work, needed for the working machine, converting hydraulic power generated by positive displacement pumps (Qm x p) into mechanical power (Mm x ω – for rotary motors, respectively Fm x v- for linear motors). Testing of linear pneumatic actuators, also called cylinders, by way of the load achieved through a hydraulic system uses test equipment of original design that connects to a hydraulic supply station and to an adjustable pressure air supply unit. Test device users will be companies manufacturing pneumatic cylinders, those reconditioning, as well as training laboratories, top and medium level. Keywords: pneumatic actuators, testing, checking, verification.

1. Introduction

Pneumatic execution elements, currently called pneumatic actuators, are intended to produce useful mechanical work, needed for the working machine in achieving its function within the specific drive chain. In terms of energy transformation, input, output and status variables, actuators turn the pneumatic power (energy) supplied by the generating elements – Qm x p into mechanical power – Fm x v at linear actuators or Mm x ω at rotary actuators- figure 1.1 and 1.2.

Qm - air flow at the engine input (actuator) p – input pressure LA – linear actuator (cylinder) RA – rotary actuator (motor)

Fm – force produced by the linear actuator v – piston speed Mm – torque ω – angular speed

Fig.1.1. Block diagram of linear actuator (LA)

Fig.1.2. Block diagram of rotary actuator (RA)

As it can be seen, hydraulic actuators are classified, in relation to the physical nature of the primary movement performed, into:

- linear actuators; - rotary actuators.

This article refers to the testing methodology in static mode of linear pneumatic actuators, with wide application in automated drives. In standard constructive design, linear pneumatic actuators are composed of a piston with rod, sealed in a liner, and caps, moving under the action of pressure and airflow - figure1.3. Special constructive design cylinders, also known as pneumatic servo cylinders, have in their structure devices that are controlled and programmed by proportional pneumatic elements.

53


Fig.1.3. Linear pneumatic actuator (cylinder) 1- brake rod head, 2-damper, 3- threaded sector of rod, 4- seals fitted piston, 5- seal guidance

for high speeds, 6- cap, 7- brake throttle

2. Verifying the pneumatic parts

Regardless the design of cylinders, on these devices in order to assess their quality, on the stand there are performed, according to the international standards, the following verifications: type checking - at prototype, zero series and constructive changes; periodic checking - after a period of manufacturing or number of pieces; batch checking.

As part of the methods of verifying the quality, an important role is played by the functional checks performed on pneumatic devices, which are performed on test stands: functional diagram, pressure resistance, exterior and interior tightness, etc.

3. Describing the schematic diagram of the stand

Below is shown the diagram of a stand on which pneumatic cylinders are tested, the antagonistic resistance force is achieved hydraulically – figure 3.1. Adjustment of testing force is performed by two pressure valves in both directions of movement of the hydraulic cylinder rigidly connected to the pneumatic actuator being tested. On this stand the following checks can be performed:

- pneumostatic pressure resistance; - exterior and interior tightness; - operation in idle and load; - starting pressure; - minimum travel pressure of the piston; - minimum and maximum travel speed of the piston; - thrust and tensile forces; - plotting the characteristic curves; - operating time (endurance).

54


Fig.3.1. Pneumohydraulic schematic diagram for checking pneumatic linear actuators The test device reproduces under actual conditions operation of pneumatic cylinders on the working machines, where the parameters pressure (which determines force), respectively flow (which determines speed) vary depending on the working cyclogram. The main subassembly of the stand is the test device that comprises the hydraulic load cylinder CHs/HCl, the pneumatic actuator which is tested Alp/LAp, served by the pneumatic installation, measuring instruments and transducers: pressure (TRp1,TRp2), force TRf, stroke TRc/TRst, and speed transducers TRv/TRsp. The hydraulic load system is powered by a low pressure pump with adjustable flow rate PH/HP, whose pressure is adjusted by the valve SP1/V1. Pump flow supplies the two chambers of the cylinder load CHs/HCl and it is designed to eliminate cavitation in the system at piston displacement. Pressure, adjustable in steps, by the valves SP1/V1 and SP2/V2, generates the load (the antagonistic force) on pulling and pushing the cylinder rod CHs/HCl by the actuator being tested. Furthermore, it is illustrated how a pneumatic actuator is tested in operation at idle and under load:

55


• after setting the pneumatic cylinder on the stand, the device is connected to the hydraulic

pumping unit and the pneumatic installation; • there are performed 15 ... 20 idle strokes, constantly actuating the electromagnets E1 and

E2, and by turns E3 and E4; • there are conducted tests under load in pressure steps – E1 and E2 not actuated, air

pressure being adjusted in the range 0…10 bar by the air preparation unit FRU, and load pressure by valves SP1/V1 and SP2/V2. The pneumatic throttles DRp1/THp1 and DRp2/ THp2 are intended to adjust the speed of the pneumatic cylinder which is tested. For working pressures lower than the minimum pressure adjusted by SP1/V1 and SP2/V2 there are used throttles DRh1/THh1 and DRh2/THh2.

Electrical parameters transmitted by transducers to the data acquisition board, DAQ and computer are then printed on the printer I/P. When checking functionality of pneumatic actuators, information about pressure from pressure transducers TRp1 and TRp2 and about force from TRf allow plotting the diagrams F = f(p) in situations when the rod advances - fig.3.2 and when it draws back – fig.3.2, for various sizes of pistons.

Fig.3.2. Variation of force depending on pressure at rod advancing

Fig.3.3. Variation of force depending on pressure at rod withdrawal

4. Conclusions

The stand has a large area of applicability at enterprises that manufacture pneumatic cylinders or recondition them, in education within technical high schools, in training people under POSDRU/HRD projects, and in case that it is computerized with transducers and computing system it may be a component of the pneumatics, tribology, mechatronics laboratories in higher education for experiments and research conducted by students and PhD students.

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’’, Magazine: Hidraulica no. 1 - 2 / 2011, ISSN 1453 – 7303 [2] Assofluid, ’’Hydraulics in industrial and mobile application’’, Milano, September, 2007 [3] C. Cristescu, P.Drumea, “Mathematical modeling and numerical simulation of the tribologic behaviour of mobile translation sealings subjected at high pressures”, Magazine no.2 (22), September 2008, ISSN 1453-7303 http://www.festo.com/net/SupportPortal/Files/142581/8000912d6.pdf

http://www.festo.com/cat/ro_ro/products_DNC

56

http://www.festo.com/net/SupportPortal/Files/142581/8000912d6.pdf

http://www.festo.com/cat/ro_ro/products_DNC


ANALYSIS OF PRESSURE IN BUCHAREST BETWEEN 2009-2012

Rusănescu Carmen Otilia1, Paraschiv Gigel1, Voicu Gheorghe1, Dorel STOICA1

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

Abstract: In this paper, we monitored the weather station atmospheric pressure based on pressure values recorded by the station, we calculated the statistical analysis of the minimum, maximum and average atmospheric pressure in the years 2009-2012, and step values in the same range baric analyzed .

Keywords: pressure, step barrel

1. Introduction

The pressure of the gaseous envelope surrounding the globe is called atmospheric pressure or barometric pressure.

Studying the atmospheric pressure regime and the distribution shows great theoretical importance because it allows explaining local circulation of the atmosphere, as well as meteorological processes such as gas exchange between the atmosphere and soil, evaporation or evapotranspiration.

Gravity is what keeps the atmosphere around the Earth and all her weight down its print them. Value in any point on the Earth's surface is equal to the weight of a column of air between that point and the upper limit of the atmosphere click on a unit area.[1]

Atmospheric pressure varies from one area to another depending on geographical latitude and temperature.

Pressure determinations at any point on the Earth's surface and in all geographic conditions, show that the pressure does not remain constant over time, presenting variations.

Pressure decreases with altitude. The decrease is not linear, but exponential - increase height in arithmetic progression, the pressure drops in geometric progression.

For Romania maxima occur at 500 and 1400 hours (in January) and 400, 1700 (in July), and minima occur at times in 1000, 2200 (January) and 900, 2400 (in July).

Figure 1 Daily variation of pressure [4]

Non-periodic variations and disturbances often called the most significant category designates the pressure changes. Their main cause is the temperature variations that generates dynamic processes in the atmosphere. Accidental pressure oscillations generally vary between 970mb and 1040 mb. In

57



exceptional cases, where depressions pressure can drop up to 925 mb and if anticyclones can grow up to 1070mb.

Atmospheric pressure varies with weather. In general, an increase in pressure means an improvement in the weather, and a decrease in pressure means a worsening of weather. Atmospheric pressure varies with altitude. As altitude is higher, the pressure will be lower. A good approximation is: for every 100 m climb in altitude, the pressure varies by 10 mb. Is approximately valid up to approximately 3000 m above sea level.

2. Materials and methods

Pressure was monitored with Weather Station: AWS / EV of ISB faculty, program version 2.3.2 Geco MICROS SIAP program automatically records the following parameters: pressure, air temperature, wind direction and speed, atmospheric humidity, solar radiation, rainfall [8]. Atmospheric pressure sensor (TBAR-V) The sensor is in accordance with international specifications WMO (World Meteorological Organization). To optimize the measurement, the sensor is equipped with electronic control inside that automatically compensates for temperature variations, ensuring good accuracy throughout operational. Structure protection range is made of plastic. Internal circuits are protected from atmospheric discharges and polarity. This is an analog sensor with linear output signal ranging from 0 V to +2 VDC. - Measuring range: 700 - 1100 mb (hPa) - Sensitivity: 0.1 mb (hPa) - Accuracy: + / - 1.5 mb (hPa) - Resolution: 0.1 mb (hPa) - Linearity: + / - 0.15% (full scale) - Temperature: - 30-60 0C - Power supply: + 10 to + 16 VDC - Output signal: 0 V (700 MB / hPa) at V + 2 (1100 MB / hPa) - The sensor connector: 4 - pin female - Mounting: with support provided (Ф mast 48-50 mm)

3. Results and discussion Established law Laplace pressure variation with altitude. This complex is a logarithmic function.

For ease of calculation was introduced step barrel. This is the vertical distance, in meters, for which there is a decrease in air pressure of 1 millibar.

Baric is calculated every step that can approximate a linear decrease of the pressure value as follows:

- At sea level falls to 1 mb to 8.4 m and 1 mm Hg for every 11.2 m; - From 5000 m to 1 mb pressure drop every 16 m; - From 11000 m to 1 mb pressure drop every 32 m [6] - In this paper, we calculate the barometric stage based on the following mathematical

relationships:

)1(8000 tp

h α+=

where: p = pressure; α = coefficient of expansion of the gas (0,04); t = temperature at that time;

58


Evolution of atmospheric pressure values from one month to another is closely correlated with

both air temperature, and especially with the dynamic atmosphere that carry latitudes investigated area baric air masses with different characteristics.

By statistical analysis based on the values of atmospheric pressure to stop we calculated the minimum, maximum, average atmospheric pressure and the results are shown in Figures 1 a - 1 d

Figure 1 a Variation of the minimum, maximum, average pressure in 2009 in the city of Bucharest

It is apparent from the graph that the atmospheric pressure in the year 2009 in the month of January 1027 mbar maximum and lowest value in September 700 mbar.

Figure 1 b Variation of the minimum, maximum, average pressure in 2010 in the city of Bucharest

Note that the maximum value was recorded in January 1029 mbar and 933 mbar minimum in December when the temperature was lower and.

59


Figure 1 c Variation of the minimum, maximum, average pressure in 2011 in the city of Bucharest

The maximum value is 1026 mbar in March and the minimum pressure of 988 mbar was recorded in July.

Figure 1 d Variation of the minimum, maximum, average pressure in 2012 in the city of Bucharest

This graph shows that the maximum value is 1027 mbar in January, and the minimum is 981

mbar station recorded all in January. Note that the air pressure imposed temporary knows a great variability of the general

circulation of the atmosphere. Given the predominance of a regime anticyclone in winter, when heat activates continental anticyclones strongly developed (Siberian and East European) and whose dorsal extend beyond South East, it records the highest values of atmospheric pressure. Weather influences the atmospheric pressure so a place can change over time. If atmospheric pressure grows we hope to

60


have nice time, if the atmospheric pressure decreases we expect the weather to become ugly. So atmospheric pressure varies from one place to another, varies with weather, altitude. When atmospheric pressure decreases (that is the lower temperature in the mountains) air is thinning and high atmospheric pressure air is more dense, more (when temperatures are high), sea.

We calculated the mean, maximum and minimum step baric in 2009-2012, based on the

values of air temperature and atmospheric pressure and the results are shown in Figure 2 a - 2 d. Other authors in their work studied the statistical analyzes various materials [3,4,5,6].

Figure 2 a The minimum, maximum, average baric stage in 2009

In 2009 the value of the step baric was 19 m in July, the lowest value in January of 1 m

Figure 2 b The minimum, maximum, average baric stage in 2010

In 2010 the highest value of baric stage was about 19 m in August, the lowest value in January

of about 1 m

61


Figure 2 c The minimum, maximum, average baric stage in 2011

In 2010 the highest value of baric stage was about 19 m in July, the lowest value in January of

about 1 m.

Figure 2 d The minimum, maximum, average baric stage in 2012

Conclusions

From the foregoing it follows that Bucharest is still located in the temperate climate and the Biotechnical Faculty of Engineering where the weather station is situated, is a protected area of influence of winds but has the advantage in terms of the values of solar radiation, which may be considered useful in the event the installation of solar panels.

Bucharest City fall transition moderate continental climate characteristic of the SE part of the Pannonian Plain, with some Mediterranean influences (version Adriatic). Its general features are marked by diversity and irregularity of atmospheric processes.

Movement of air masses from the west so in the cold persists and in the warm seasons and is characterized by mild winters, often with liquid precipitation. Frequently, even during winter, arriving from the Atlantic humid air masses, bringing significant rain and snow, less cold waves. Polar circulation is determined by cyclones in the North Atlantic Ocean and is characterized by decreases in temperature, cloudiness and precipitation in the form of sharp showers, and winter snowfall is accompanied by intensification of the wind. From September to February is manifested frequent intrusion of continental polar air masses coming from the East. However the influence is strongly felt cyclones and warm air masses from the Mediterranean, which generates thaw winter and summer

62


periods require stifling heat. Tropical circulation causes mild winters with large amounts of rainfall and the summer a while unstable with rain showers and thunderstorms.

Cold and dry air masses are accompanied by land surface pressures higher than warm and humid air masses. In turn, hot air, dry determined pressure greater than a mass of hot air as but not wet. Replacement of air masses with different can mean changes in atmospheric pressure and time, but the surface pressure fluctuations may present even without changing air masses because pressure may decrease or increase as local air is heated or cooled. In addition to changes in air pressure resulting from variations in the temperature and water vapor content, the pressure may also be influenced by the type of air circulation. Winds diverge from a central point on the earth's surface causes the center air descending from above, the diverging air taking place, if the surface diverges more air than the top down, the air density and pressure drop. In the case of wind converging to a point on the earth's surface, if more air than converge rises to higher altitudes, the air density and pressure increase.

REFERENCES

[1] APOSTOL L. – „Meteorologie si climatologie”, Editura Universitatii Suceava 2000 [2] N. Topor, V. Mosoiu, N. Vancea - Meteorologie Aeronautica 1967 [3] D. Stoica, G Stanciu: “Influence the degree of sorting the separation process a conical sieve” Digest Journal of Nanomaterials and Biostructures Vol. 8, No. 2, April - June 2013, p. 513 - 518 [4]Irina Istrate, Diana Cocârță, Silvia Neamțu, Talida Cirlioru, The assessment of an alternative treatment for PCB polluted soils in the romanian context - bench scale tests, 2012, Water air and soil pollution, vol 224 (4),DOI: 10.1007/s11270-013-1516-2, ISSN: 0049-6979 (print version), ISSN: 1573-2932 (electronic version); [5] Dumitru Popovici, George Strejea, Ion Mihaila – Performante si limite umane in aviatie , Bucuresti 2009. [6] C. O. Rusănescu, G. Paraschiv, G. Voicu, M. Rusănescu - Comparative Analysis of Atmospheric Temperature Values, Relative Humidity In 2009 And 2010 In West Side Of Bucharest City, Bulletin USAMV Agriculture, 68(2)/2011, Print ISSN 1843-5246; Electronic ISSN 1843-5386, pag. 130-138 [7] Rusănescu Carmen Otilia – Îndrumar de meteorologie si climatologie, Editura Matrix Rom, Bucuresti 2013

63


EXPERIMENTAL TESTING OF A LOW SPEED HYDRAULIC MOTOR WITH AXIAL PISTONS

drd. ing. Laura GRAMA1, dr.ing. Daniel BANYAI2, dr. ing. Liviu VAIDA3 1 Technical University of Cluj-Napoca,[email protected] 2Technical University of Cluj-Napoca, [email protected] 3Technical University of Cluj-Napoca, [email protected]

Abstract:In the paper is presented a new design of low speed hydraulic motor with axial pistons. It work stable at low speeds providing high torque. Another important advantagethat brings this new model of stepper motor is the precision of the angular position that can be achieved. Also it is cheaper then electric motors with the same power.The paper presents the results from tests with an experimental model. Analyzing these results can be said that the motor fulfills the requirements on its steady and dynamic behavior.

Keywords:Hydraulic motor, axial pistons, low speed.

1.Design of the new low speed hydraulic motor

The hydrostatic motors have apppeared and been used from the 17th century. Inclined shaft motors, working at high pressure have been developed by Thoma in 1930. In 1950 axial piston motors with tilted disc successfully emerged. Over time there have been researches on hydraulic piston motor, in increasing their performances and accuracy. [2], [3], [4], [7].

Rotary hydraulic motors have to ensure great stability of movement in a wide range of variation of the output values (1 rev/min ... 3000 rev/min) and aratio between active torqueand moment of inertia, grater then electrical drives for the same power. [2], [3], [4].

In case of heavy loads at low speed, we need “slow” running motors that work stable at low speeds and provide high torque. Starting with this premise, it is developed a version of a hydraulic motor with axial pistons that is shown in figure 1. Through this design it was intended to adjust the motor speed without dissipating energy by adjusting the feed rate, but which is able to transmit to the shaft of the motor a greater torque for the entire speed range. [1], [5], [6].

The piston block is attached to the motor housing (8) and secured against rotation. The pistons (4) are arranged in a circle in the block, having the axes parallel to the axis of the piston port block. In line movement is determined by supplying the piston chambers with pressure oil via rotary distributor (6). Turning it is done with a rotary electric motor (stepper motor or DC electric motor (7)).

Pistons of the hydraulic motor (4) under the action of the pressurized fluid perform a reciprocating motion in contact with a tilting drive. The pistons are acting on the disc by means of connecting rods having both ends spheric. Removing the piston rod and the disk is prevented by their crimping pistons, or by means of a retaining plate attached to the disc.

Due to the alternative movement of the pistons, to a full turn of the rotary distributor, the tilted disc performs a swinging movement around its axis. Rotating the distributor allows the pistons to connect to the supply and tank connections through the holes in the block and windows made inside the block.

64


The rotary distributor must ensure hermetic sealing of cylinder pistons near dead points,

areas that allow connecting the cylinders alternatively to the pressure or tank. Through the rotary distributor it separates the high of low pressure zone of the motor.

Fig.1. Low speed hydraulic motor: 1 – shaft, 2 – driven gear, 3 – drive gear (tilted disc), 4 – piston,5 – piston block port, 6 – distribuitor, 7 – electric motor, 8 – housing.

2. Description of the experimental stand

Schematic diagram of the testingstand is shown in Figure 2 and the panoramic description of the rig is shown in Figure 3.

Fig. 2. Scheme of the experimental stand.

65


The elements numbered in Figure 3 are the following:

1. Control panel for electrical motor;

2. Electric stepper motor;

3. Low speed hydraulic motor;

4. Hydraulic brake;

5. Pressure sensor;

6. Hand pump;

7. Device for measuring the pressure;

8. dSpace panel;

9. Computer.

Fig. 3. Experimental rig (overview image).

Figure 4represents the application made in MathLab Simulink software, performed for command and data acquisition of the experimental stand.[9]

66


Fig. 4. Control and data acquisition application

Graphic blocks marked "From Real World" and "To Real World" correspond to functions for accessing hardware resources dSpace system. Amplification blocks serve to scale physical quantities in the interval [-1,1]. Block "Signal Generator" transmits the frequency that control the electric stepper motor which drives the distributor.

With the Control Desk software was made a GUI for controlling and viewing the parameters. According to figure 5, this interface contains a block for entering numerical frequency for electrical stepper motor, two display windows for pressures. Speed of the hydraulic motor and distributor are plotted in two separate windows. [8]

67


Fig. 5. GUI for data acquisition and control

3. Conducting experiments

There has been two series of experiments. In a first series, no load was considered, for different supply pressures. For each value of the supply pressure were ordered seven different speeds, introducing at the keyboard interface feed rate values corresponding to the stepper motor.

In this case the power supply of the motor was determined by the relationship:

[ ]600sourceQ pP kW⋅

= (1)

Where:

Q- flow [l/min];

p- pressure [bar].

The measured values from the experimetal research, are shown in Table 1.

Tab. 1. Testing without load

Nr.

crt.

Supply

pressure [bar]

Speed at

stepper

motor

[rot/min]

Speed at

hydraulic

motor

[rot/min]

Pressure

hydraulic

brake

[bar]

Braking

torque

[Nm]

Power

source

[kW]

Power

hydraulic

motor

[kW]

1

30

100 2

0 - 0,675 -

2 150 3

3 200 4

4 250 5

5 300 6

6 400 8

7 500 10

8

45

100 2

0 - 1,01 -

9 150 3

10 200 4

11 250 5

12 300 6

13 400 8

14 500 10

15

60

100 2

0 - 1,351 - 16 150 3

17 200 4

68


18 250 5

19 300 6

20 400 8

21 500 10

A second series of experiments was performed by placing a load torque on the motor, again for different supply pressures and the drive speed of the distributor.

For the test conditions, the values of the power supply has been calculated with the equation (1), the braking torque determined with (2) and the motor power with (3):

[ ]2med

fr p frdM F Nmµ= ⋅ ⋅ (2)

[ ]motorP M kWω= ⋅ (3)

Where:

dmed- the mean diameter of the braking disc [mm];

μ- friction coefficient;

Fp fr- braking force [N];

ω- angular velocity [rot/min];

M- torque [Nm].

The parameter values for this series of measurements are presented in Tables 2, 3, 4, for maximun pressure of 30, 45, 60 bar.

The values for the measurements at the supply pressure of 30 bar is given Table 2.

Tab. 2. Testing at 30 bar

Nr.

crt.

Speed at

stepper motor

[rot/min]

Speed at

hydraulic

motor

[rot/min]

Pressure

hydraulic

brake

[bar]

Braking

torque

[Nm]

Power

source

[kW]

Power

hydraulic

motor

[kW]

1 100 2 3,6 19,94

0,675

0,268

2 150 3 3,2 17,72 0,278

3 200 4 2,7 14,95 0,313

4 250 5 1,8 9,97 0,261

5 300 6 1,2 6,64 0,208

69


6 400 8 0,8 4,43 0,185

7 500 10 0,4 2,21 0,116

Data from measurements made at a pressure of 45 bar can be seen in Table 3.


Nr.

crt.

Speed at

stepper motor

[rot/min]

Speed at

hydraulic

motor

[rot/min]

Pressure

hydraulic

brake

[bar]

Braking

torque

[Nm]

Power

source

[kW]

Power

hydraulic

motor

[kW]

1 100 2 5,9 32,68

1,01

0,342

2 150 3 5,5 30,47 0,478

3 200 4 4,8 26,59 0,556

4 250 5 4,1 22,71 0,594

5 300 6 1,8 9,97 0,313

6 400 8 1,2 6,64 0,278

7 500 10 0,6 3,32 0,173

Table 4 displays the experimental data from testing at 60 bar.


Nr.

crt.

Speed at

stepper motor

[rot/min]

Speed at

hydraulic

motor

[rot/min]

Pressure

hydraulic

brake

[bar]

Braking

torque

[Nm]

Power

source

[kW]

Power

hydraulic

motor

[kW]

1 100 2 6,9 38,22

1,35

0,400

2 150 3 6,6 36,56 0,573

3 200 4 6,3 34,90 0,730

70


4 250 5 6 33,24 0,869

5 300 6 5,3 29,36 0,922

6 400 8 3,8 21,05 0,881

7 500 10 2,7 14,95 0,782

4. Experimental results

For experiments without load at the motor shaft, there has been done the following actions:

• It was fix the supply pressure of the low speed hydraulic motor at 60 bar;

• The values for the speeds of the rotary distributor (100, 150, 200, 250, 300, 400, 500 rot /

min);

• For each value of speed were drawn response diagrams depending of the slow speed

hydraulic motor (Figure 6).

Fig. 6. Diagram of speed at 100 rot/min value.

For the other speeds distributor diagram is still linear, with values from Table 1.

For sets of measurements with load:

• Three values for supply pressure were successively set (30, 45, 60 bar);

• For each value for maximum pressure, the values for the speeds of the rotary

distributor (100, 150, 200, 250, 300, 400, 500 rot / min);

• For each value for maximum pressure,was increased the load until the motor speed

was zero;

• For each set of measurements diagrams were drawn (Fig. 7, Fig. 8,Fig. 9).

71


Fig. 7. Supply pressure 30 bar, speed distribuitor 100 rot/min.

For other speeds of the distributor, the diagrams look the same except that the values given

in Table 2.


The other diagrams are the same, only the values, shown in Table 3.


72


For other values of speed, diagrams have a similar shape, the values are presented in

Table 4.

5. Conclusions Were clarified the parameters which strongly influences the behavior of the low speed

motor, leading thus to find ways to improve their performance.

The trend of increasing the pressures is limited for these motors because the pressure in thepistons chambers made forces acting on the gear teeth so the materials used require special measures.

Finally we can say that the new hydraulic motor behaves as a dynamic system very stable at real operating conditions. Compared to equivalent electrical systems, they are robust, reliable and have a small size, also can achieve very small angular displacements and high torques.

6. REFERENCES [1]Banyai D.V., „Metode noi în sinteza maşinilor hidraulice, cu volum unitar variabil şi reglare electro-

hidraulică”, teză de doctorat, 2011. [2]Banyai D., Vaida L., Năşcuţiu L., Opruţa D., Giurgea C, Marcu I.L., Synoptic view of the latest trends in

hydraulic actuation, ICMS 2009, Buletinul Institutului Politehnic din Iași, Tom.LVI (LX), Fasc.1, Secția Construcţii de Maşini, p.148-155, ISSN 1011-2855, Iași, 2010.

[3]Deacu L., “Hidraulica maşinilor unelte”, Ed. Dacia, Cluj-Napoca, 1989. [4]Drumea P., Echipamente mecatronice cu structură hidraulică, Simpozion HERVEX, Olănești, 1999. [5]Vaida, L.,Năşcuţiu, L., Deacu, L., Stand experimental pentru determinarea caracteristicilor motoarelor

hidraulice, Buletin informativ – HIDRAULICA, Institutul de cercetări pentru hidraulică şi pneumatică, Bucureşti, 1999. ISSN/453-7303.

[6]Vaida L., Opruţa D., Năşcuţiu L., Implementarea pompelor si motoarelor reglabile electro-hidraulice în sisteme informatizate de conducere a maşinilor, Lucrările celui de-al III-lea Simpozion, Metode moderne în ingineria echipamentelor hidro, Editura Politehnica, Hidrotim S.A., p. 134, ISBN 973-8247-45-4, Timişoara, 27-28 septembrie 2001.

[7]*** www.assofluid.it. [8]*** www.dspace.de. [9]*** www.mathworks.com.

73

http://www.assofluid.it/

http://www.dspace.de/

http://www.mathworks.com/


SONIC EFFECTS OF A UNCONVENTIONAL HEAT INSTALLATIONS Prof.dr. eng. Carmen BAL1,prof. dr. eng.Nicolaie BAL2, lecturer eng. Lucian Marcu3,eng.

Carmen Ioana IUHOS4

1 Technical University of Cluj Napoca, [email protected]; 2,3Technical University of Cluj Napoca, [email protected]. 3S.C. BROKER SA; [email protected].

Abstract:The present work aims, through the theme addressed, to make some contributions to a better knowledge of the problems related to the study of theoretical and experimental sonicity in heat transfer. Based on these considerations, it has built a stand allowing implementation of the theory of sonics actuator systems with harmonic flow rates to send sonic energy and remote to turn into heat by friction between the phenomenon of the column of fluid friction and resistance.The stand consists of hydraulic pump, capacitors, tubes and resistance of friction.In the present work the experimental version of the installation is the serial installation: small condenser monted after friction resistance Keywords: sonic pressure, sonic flow, sonic capacity, friction resistance, condenser..

1.Introduction

The present work aims, through the theme addressed, to make some contributions to a better knowledge of the problems related to the study of theoretical and experimental sonicity in heat transfer. Sonic enables the actuators make optimum facilities offered by ease of processing electrical signals (low energy) actuators make.sonic with high power and efficiency, which give the possibility to eliminate the biggest parts of a classic waterworks (reservoirs, hydraulic valves, differential pressure control, flow path, etc.), resulting in an actuator that combines the best opportunities offered by processing technique of low-power signals and sonic compact actuators make, high-efficiency, low-volume, so very economical. The great inventor Gogu Constantinescu has spent an enormous energy to convince the world that liquids are compression connectors with a lot more than they accept, and this property is essential for the propagation of vibrations in liquids. As has been said right from her appearance, the sonics is in correspondence with electricity, and sonic transmissions are similar to AC electrical transmissions.Considering the above analogy, compressing liquids is equivalent to proving the charge build-up in a capacitor. Sonic transmission is achieved by vibration, and at the beginning of the 20th century, believed that the energy of vibration is a degrading form of energy that can not only turn into heat. It was unthinkable that in a system of vibration can get work with high efficiency. Based on these considerations, it has built a stand allowing implementation of the theory of sonics actuator systems with harmonic flow rates to send sonic energy and remote to turn into heat by friction between the phenomenon of the column of fluid friction and resistance. The stand consists of hydraulic pump, capacitors, tubes and resistance of friction. To find the version with maximum efficiency experiments were made in several versions of the stand: 1. stand consisting of: pump, resistance and a capacitor connected in series; 2. stand consisting of: pump, stamina and two capacitors fitted in parallel; 3. stand consisting of: pump friction resistance and two capacitors fitted in the series; 4. stand consisting of: pump friction resistance and capacitive with two cylinders mounted in the

extension.

74



2. Experimental research

Experimental research focused on getting the calorie effect of heat transmission through the vibrations (sonic waves in fluids). Such research has been conducted on the stand, starting from the different frequencies of the propulsion engine piston sonic generator. For each frequency were conducted three static pressure measurements in the plant having the values(0; 0,25;0,5; 0,75; 1; 1,25; 1,5) bar. In the present work the experimental version of the Installation in the series:Small condenser monted after friction resistance. The stand on which the trials were made is the one shown in Figure 1. Sonic generator from leaving a pipe which connects withthe resistance of friction "Rf" which is connected to the small condenser whose volume is V = 1462,411 103 mm3.

After processing files with experimental data obtained from the three sensors fitted in the system is the result of primary form of histogram shown in figure 2. This illustrates the pressure generator and developments at the two capacitors. Also you can see the speed of the generator (via view-position curve generator). Pressure curve evolution reveals a phase shift between the generator and pressure the pressure in the capacitors.

Fig. 1 Scheme of experimental stand 1. electric motor; 2- proximity sensor; 3- elastic coupling;

4- Sonic pump; 5, 8, 10- pressure sensors; 6- temperature sensor; 7- resistance of friction; 9- small capacitor; 11- large condenser; 12, 14- tap, 13- pump; 15- oil tank.

1 2 3 4 6 7 8 9 10 11 12 13

14

15

A B CA B C

A B C

e

M

5

Fig.2. The evolution in time to mount pressure withsmall capacitor in series

75


Research has been conducted on the arranged as in Figure 1. Figure 3 and 4 have represented diagrams of pressure and temperature variation in function of time and speed for static pressure of 0 Pa. pressures on those two cylinders not capacitive changed instead to pressure generator remains constant around smaller 40E+05 Pa.

In the graphs of resultsnote: - ΔG- sonic's pressure variation at sensor pump 5; - ΔS1-pressure variation obtained from the pressure sensor 8; - ΔS2-pressure variation obtained from the pressure sensor 10; - T-temperature.

Although the speed was very high temperature of 2200 rpm failed to exceed 30° C, there may be air in the plant whereas after a period of time it was stopped by the electric motor. In Figure 5 and Figure 6 he represented diagrams of pressure and temperature variation in function of time and speed for static pressure of 0,5E+05 Pa. In this situation it becomes apparent that pressures on pressure sensors are on cylinders are approximately equal and constant having a pressure of about 50E + 05 Pa. Pressure generator is labile, and after about 50 got a jump up to 250E+05 Pa, at which point the engine shutdown occurs. The temperature does not exceed 48°C until turning off the electric motor.

n = 1400 rot/min ps = 0 Pa

Fig.3 Chart of the variation of pressure and temperature depending on the

time of static pressure of 0 Pa

05101520253035

05

101520253035404550

0 20 40 60 80

tem

pera

tura

[ºC

]

pres

iuni

[E+0

5 Pa

]

timp [ms]

Δ G Δ S1 Δ S2 T [ºC]

Fig.4 Diagram speed variation depending on time for static pressure 0 Pa

76


n = 800 rot/min ps = 0,5E+05 Pa

Fig. 5.Diagram of variation of pressure and temperature depending on the time of static pressure of 0,5E+05 Pa

ps = 0,5E+05 Pa

Fig. 6.Diagram speed depending on variationfor static pressure 0,5E+05 Pa

n = 800 rot/min ps = 2 E+05 Pa

Fig.7.Diagram of variation of pressure and temperature depending on the time of static pressure of 2E+05 Pa

01020304050607080

020406080

100120140160180200

0 10 20 30 40 50

tem

pera

tura

pres

iune

[E+0

5 Pa

]

timp [s]

Δ G Δ S1 Δ S2 T [ºC]

77


ps = 2E+05 Pa

Fig. 8 Diagram speed depending on variationfor static pressure 2E+05 Pa

Figure 7 and 8 he represented diagram of pressure and temperature variation in function of time, as well as the stages for static pressure of 2E + 05 PA. The speed was 1400 RPM. the pressures the two cylinders are approximately constant and equal to the smaller 40E+05 Pawhile the pressure generator increases continuously up to 150E+05Pa, after which there is a drop in pressure taking place and stop the engine electric. The maximum temperature reached is 70 °C.

3. Conclusions

From the analysis of the above charts can be drawn the following conclusions: -the largest influence on the increased pressure and temperature in the plant a revolution, which has the higherpower the higher the faster the pressure and temperature; -static pressure in the plant to a lesser extent influences the evolution of temperature in friction resistance; -the pressure in the cylinder increases much faster than in the case of linking in large cylinder series; -because the electric motor turns off after a short period of operation (about 1 min) is not recommended for the standconsists of a resistance of friction and a cylinder.

ps = 2E+05 Pa

Fig. 8 Diagram speed depending on variationfor static pressure 2E+05 Pa

78


Experimental results achieved with other pressures and revs were not significant strengthening the conviction not to recommend us connections in this form.

REFERENCES

1. Abrudean, M., 1998, Teoria sistemelor şi reglare automată, Editura Mediamira, Cluj-Napoca,. 2. Constantinesco, G., 1918, Theory of Sonics: A Treatise on Transmission of Power by

Vibrations, The Admiralty, London, 3. Constantinescu, G., 1985, Teoria sonicităţii, Editura Academiei R.S.R., Bucureşti , 4. Carmen Bal, (2007), The caloric effect in the circuits by harmonic flow, Cluj Napoca, Ed. ALMA

MATER, ISBN 978-973-7898-75-3. 5. Carmen Bal, (2006), Research and contributions about the drive systems with the harmonic

flow, the doctoral thesis Technical University of Cluj Napoca. 6. Pop I. Ioan, Marcu Lucian,Carmen Bal ş.a., (2007), Sonicity applications. Experimenthal

results. Iaşi, Ed. Performantica, ISBN 978-973-730-391-2. 7. Pop, I., Khader, M., Marcu, I. L, s.a, 1999, Modern hydraulics. Pneumatics, Editura U.T.Pres,

Cluj-Napoca, 8. Pop, I., Marcu, I. L., s.a,2004, Acţionări hidraulice moderne. Pneumatică, Editura U.T.Pres,

Cluj-Napoca,. 9. Pop, I., Marcu, I. L., 2004., Personalităţi în ştiinţă şi tehnică - Gogu Constantinescu, Editura

AGIR, Bucureşti, 10. Pop, I., 2006, Tratat de teoria sonicităţii,Editura Perfomantica, Iasi.

79


ANALYTICAL MODEL OF THE CONNECTION PIPES OF THE ALTERNATING FLOW DRIVEN HYDRAULIC SYSTEMS

Ioan-Lucian MARCU1, Daniel-Vasile BANYAI1 1 Technical University of Cluj-Napoca, [email protected] 1 Technical University of Cluj-Napoca, [email protected]

Abstract: The paper presents the specific equations that are defining the transfer functions of the connection pipes of the alternating flow driven hydraulic systems, determined based on each component specific characteristics like fluid and pipe elasticity, inertia, friction and leaks. Also there are presented the relations which are defining the amplitude and the phase alteration.

Keywords: analytical model, alternating flow, hydraulic system

1. General aspects

Alternating flow driven systems involves a new approach of the driving systems using pressurized liquids, because we have here, in the entire system, along the pipes, an energy transmissions without volumetric flow transportation between the energy converters, hydraulic generator and hydraulic motor. [5], [7], [8]

Generally, an alternating flow driven hydraulic transmission consists in a alternating flows and pressures generator (G) and a motor (M), the connection between them being realized with a number of pipes equal with the number of phases (Phase 1, Phase 2 and Phase 3), the pipes being filled with fluid at a certain pressure (pre-established with an hydraulic accumulator Ac), figure 1. During the functioning of the system the pressure and the flow within each pipe varies in a sinusoidal way, around an average value.

Rz1

Rz2

G

ME ~

C

Rz3

Ac

R

Phase 3

Phase 2

Phase 1M

~

Fig. 1. Principle schema of an alternating flows and pressures drive hydraulic system.

In order to have a proper functioning it is compulsory that this average pressure from each pipe to have the same value and to have a constant value in time. Therefore, to obtain the correct functionality we create from the beginning either a pressure in each phase, higher than the amplitude maximum value, or this pressure is modifying itself during the functioning.

80



2. Analytical model of the alternating flow driven hydraulic systems

The precision of the analyzed aspects, the design and the physical realization of an automatically system depends on the complete modeling possibilities, using characteristic equations for each component. It is also very important the correct determination of each constant value from the equations structure which involve the establishment of the frequency functions. The transfer functions of the elements (components) represent the rate between the Laplace transformations of the output respectively input signal, for null initial conditions. The general effect of using the Laplace transformation is the reducing of difficulty order of the problems. The transfer function algebra contains some rules which allow combining the transfer functions of many components and finally to obtain the transfer function of the entire assembly of individual elements. [1], [4] The transfer function of the entire system was determined considering that the input signal is a harmonic one (sinusoidal), signal provided by alternative movement of the hydraulic generator piston, and the output signal is also harmonic, but having an amplitude and phase angle alteration. [8] The dynamic system schema is represented in figure 2.

Fig. 2. Representation of the hydraulic pressure losses in the connection pipe.

We assume that the governing equation for the instantaneous flow are: [5], [7]

( )0max sin ϕω +⋅= tQQ ai (1) In which:

2maxgg

a

ShQ

⋅⋅=ω (2)

The instantaneous flow Qi, equation (1), provided by the generator, is producing a variation due to the combined effect of the frictions, capacity, inertia and leaks, figure 2. The dependences between these instantaneous pressures and instantaneous flow are: [7]

ifif QCp ⋅= (3)

dtdQ

Lp iiL ⋅= (4)

∫ ⋅⋅= dtQC

p ih

iC1 (5)

P

iiP C

Qp = (6)

The friction coeficient is:

cc

c

f dSg

lC ef

V

⋅⋅

⋅⋅⋅=

2

γε

⋅

5msN (7)

in which:

81


ε - no dimensional coefficient; γ - the specific weight of the oil [kg/m3]; lc - the length of the pipe [m]; Vef - the oil speed [m/s]; g - the gravitational acceleration [m/s2]; Sc - the pipe area [m2]; dc - the inside diameter of the pipe [m]. Using experimental researches the coefficient ε is estimated by the equation: [7]

cef dV ⋅+=

18,002,0ε (8)

The inertia or the hydraulic inductance is defined by the equation:

c

c

Sgl

L⋅⋅

=γ

⋅5

2

msN (9)

The hydraulic capacity, considering the oil and pipe compressibility is:

−

+

⋅

⋅+⋅=

1

5,15,211

2

2

0

int

int

c

c

c

c

conductauleih

dd

dd

EEVC

ext

ext

N

m5

(10)

in which: V0 - the initial oil volume from the pipe [m3]; Eulei - elasticity modulus of the oil [N/m2]; Econducta - elasticity modulus of the pipe [N/m2];

extcd - the inner diameter of the pipe [m];

intcd - the outer diameter of the pipe [m].

The coefficient which is defining the hydraulic leaks can be evaluated using the formula: [2]

ηπ

12

3hdC p⋅⋅

= (11)

The combined effect of the friction, inertia, capacity and leaks, figure 2, is expressed by summing the specific pressures, defined by the relations (3), (4), (5) and (6), obtaining in this way:

iPiCiLifi ppppp +++= (12) or:

P

ii

h

iifi C

tQdttQ

CdttdQ

LtQCtp)(

)(1)()()( +⋅⋅+⋅+⋅= ∫ (13)

Using the Laplace transformation we obtain:

)(1)(1)()()( sQC

sQsC

sQsLsQCsp iP

ih

iifi ⋅+⋅⋅

+⋅⋅+⋅= (14)

or:

)(11)( sQsC

sLC

Csp ihP

fi ⋅

⋅

+⋅++= (15)

In this way: )()()( * sQsFsp ici ⋅= (16)

82


in which:

sCsL

CR

sQsp

sFhP

fi

ic ⋅

+⋅++==11

)()(

)(* (17)

is representing the transfer function of the pipe. The equation (17) can be reconsidered like:

sC

sCC

CsCLsF

h

hP

fh

c ⋅

+⋅⋅

++⋅⋅

=11

)(

2

* (18)

The Laplace operator is ω⋅= js [6], and in this way the equation (18) is:

)(

1)(1)()(

2

*

ω

ωωω

jC

jCC

CjCLjF

h

hP

fh

c ⋅

+⋅⋅

++⋅⋅

=

or:

h

Pfhh

c CjC

CCjCLjF

⋅⋅

+⋅⋅⋅+⋅⋅−

=ω

ωωω

11)(

2

* (19)

considering: 12 −=j . The transfer function of the pipe can by expressed like:

⋅

−⋅⋅+

+=

hPfc C

LjC

CjFω

ωω 11)(* (20)

The real and the imaginary components are:

Pf C

C 1Re += (21)

and:

hCL

⋅−⋅=ω

ω 1Im (22)

In this way, the amplitude and the phase alteration of the transfer function of the pipe is defined by the equations:

22* 11)(

⋅

−⋅+

+=

hPfc C

LC

CjFω

ωω (23)

and:

Pf

hc

CC

CL

arctgjF1

1

)(*

+

⋅−⋅

=<ω

ωω (24)

Knowing that the input signal, the instantaneous flow is harmonic equation (1), then the output signal will be also harmonic, but having a different amplitude and phase alteration. Then, the output signal which is the instantaneous pressure is defined by:

83


+

⋅−⋅

++⋅

⋅

−⋅+

+⋅=

Pf

h

hPfai

CC

CL

arctgtC

LC

CQp1

1

sin110

22

maxω

ωϕω

ωω (25)

From equation (25), using the equations (23) and (24), we obtain the amplitude of the output pressure, like:

22

maxmax11

⋅

−⋅+

+⋅=

hPfaa C

LC

CQpω

ω (26)

and the phase alteration:

Pf

hc

CC

CL

arctg1

1

+

⋅−⋅

=ω

ωψ (27)

If we not taking into account the hydraulic leaks coefficient Cp , then the transfer function defined by the equation (18) is:

+⋅+⋅⋅=

hfc C

sCsLs

sF 11)( 2* (28)

Another form of the transfer function of the pipe is:

+⋅+⋅⋅

⋅= 1

211)( 22

* ssCs

sFnc

c

nchc ω

δω

(29)

in which the natural frequency is defined by:

hnc CL ⋅=

1ω (30)

and the damping ratio of the pipe is:

LCC hf

c ⋅=2

δ (31)

Considering the Laplace operator ω⋅= js we obtain a new expression of the transfer function:

h

ncnc

c

c C

jjF

⋅

−⋅+

=ω

ωω

ωδ

ω1

2

)(2

2

* (32)

The natural frequency becomes:

−

+

⋅

⋅+⋅⋅

=

1

5,15,211

1

2

2

2

int

int

c

c

c

c

conductaulei

c

nc

dd

dd

EEgl

ext

ext

γ

ω (33)

Also, the damping ratio of the pipe will be:

84


−

+

⋅

⋅+⋅⋅

⋅⋅⋅⋅

⋅⋅⋅=

1

5,15,211

4 2

2

2

int

int

c

c

c

c

conductaulei

c

cc

c

c

dd

dd

EESg

dSg

l

ext

ext

efV

γ

γεδ (34)

In this way, the new equation defining the amplitude and phase alteration are: 2

2

2

2

2* 1

41)(

−+⋅

⋅=

ncnc

c

hc C

jFωω

ωδ

ωω (35)

−⋅=< 1

2)( 2

2*

ncc

ncc arctgjF

ωω

δω

ω (36)

3. Conclusions The objective of this study was a new approach of the hydraulic drives, in which the pressure and flow is not continuously transmitted between the energy converters (pumps and motors).

The analytical model concern the connection pipes, taking into account some specific characteristics like fluid and pipe elasticity, the fluid inertia, the fluid friction and also the leaks. Simulating this model we can observe the possibility to adjust, during the functioning, the input parameters, in order to obtain the anticipated output values of some parameters. REFERENCES

[1] Coloşi, T., Ignat, I., "Elemente de teoria sistemelor şi reglaj automat", Lito. IPCN, Cluj-Napoca, 1981. [2] Deacu, L., Pop, I., "Hidraulica maşinilor-unelte", Lito. I.P.C.N., Cluj-Napoca, 1983. [3] Marcu, I.L., Pop, I. “Interconnection possibilities for the working volumes of the alternating hydraulic motors”

Proc. of the 6th International Conference on Hydraulic Machinery and Hydrodynamics - HMH2004 in Trans. of Mechanics, Tom 49 (63), Timisoara, October 2004, ISSN 1224-6077, pp. 365-370.

[4] Ogata, K., System dynamics, Pretince-Hall, Englewood Clifs, New Jersey 07632. [5] Pop, I. et all, "Conventional Hydraulics", Ed. U.T.PRES, Cluj-Napoca, 1999. [6] Pop, I., et all, "Machines-tools acoustics and dynamics", Editura U.T.Pres, Cluj-Napoca, 2000. [7] Pop, I., "Sonic Theory Treatise", Ed. Performantica, Iasi, (2006). [8] Pop, I. et all, "Sonics Applications. Experimental Results", Ed. Performantica, Iasi, 2007.

85

euInternational Salon of Hydraulics & Pneumatics - HERVEX

Hidraulica Magazine 3 2013

Documents

Transcript of Hidraulica Magazine 3 2013