Development of Electro-pneumatic Fast Switching Valve And

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Development of electro-pneumatic fast switching valve andinvestigation of its characteristics

Elif Erzan Topcu, _Ibrahim Yuksel *, Zeliha Kams

University of Uludag, Faculty of Engineering and Architecture, Department of Mechanical Engineering, 16059 Bursa, Turkey

Received 15 March 2005; accepted 18 January 2006

Abstract

The paper describes the development of an electrically operated pneumatic fast-switching valve for pneumatic position control sys-tems. The valve was designed and fabricated as an onoff valve with 2/2-way function and its characteristics were investigated. Thedetailed theoretical analysis and preliminary tests were carried out on the manufactured valve and showed that the switching time ofthe single stage valve is about 4.5 ms and provides a 460 l/min of flow rate at standard air conditions. Production and assembling ofthe valve is rather simple and does not need a manufacturing operation with tight tolerance and therefore, it is envisaged as a low costvalve. The four of 2/2-way function valves are applicable for the realization of pneumatic actuators driving with a well-known establishedtechnique of the pulse width modulation (PWM). 2006 Elsevier Ltd. All rights reserved.

Keywords: Electro-pneumatic; Fast switching solenoid valve; PWM; Mechatronic

1. Introduction

Electro-pneumatic control valves are used as interfacesto electronic controls to allow the infinitely variable electri-cal remote adjustment of fluid flow of the valve outputdriving a pneumatic actuator. There are two types ofelectro-pneumatic valves used in the realizing control offluid flow of the pneumatic actuator. These are servovalvesand onoff valves.

The servovalves are used to achieve high linear control

accuracy in pneumatic actuators, but they have complexstructures and they are very costly. On the other hand,owing to low cost, high flow rate gain, small size and simplestructure, the fast-switching onoff valves that are used tocontrol the pneumatic actuators have received considerableattention[1,2]. They are also known[3,4]as digital valves.The onoff switching valves are inherently nonlinear dis-crete electro-pneumatic converters. To obtain similar linear

characteristics of the servovalves with simple onoff switch-ing valves, pulse width modulation technique is used. Toreach a large modulated linear area, the valve switchingtimes must be very rapid during the short cycle time. Thelimits of the modulation depend on the valve switchingtimes and the modulation frequency.

The conventional single stage solenoid operated onoffvalves are very bulky and their dynamic performances arelow. With these valves, fine motion control is difficult toachieve because of the limitation of the valve response

time. Therefore, it needs a fast switching valve to get agood performance. As a result, a considerable amount ofresearch has been devoted to develop various position con-trol systems using fast response onoff solenoid valves[57].

A very early application of pulse width modulationschemes in fluid power systems started with the adaptationof conventional electro-hydraulic servo valves for using asswitching valves[8]. However, the first application of pulse-modulated control in pneumatic systems emerged towardsthe end of 1960. Goldstein and Ricardson[9] investigated

0957-4158/$ - see front matter 2006 Elsevier Ltd. All rights reserved.

doi:10.1016/j.mechatronics.2006.01.005

* Corresponding author.E-mail address: [email protected]( _I. Yuksel).

Mechatronics 16 (2006) 365378
mailto:[email protected]:[email protected]


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the principles of pulse-modulated control using free float-ing-flapper disc switching valves. A study was carried outby Taft and Harned [10] on the development of a four-way electro-pneumatic valve utilized permanent and ferro-magnetic valve switching elements. Yuksel [11] developedtwo types of novel electro-hydraulic floating disc type offast switching valves where a disc type of magnetic circuitwas used to operate the valves electrically. Later the inves-tigation on the development of the disc valve has been car-ried out by several researchers[1215].

PWM technique in the application of fast switching sole-noid valves that are used to drive linear hydraulic actuatorswere investigated by several authors. Muto et al. [16] andalso Suematsu et al. [17] have used differential PWMmethod to control a hydraulic actuator with three-wayhigh-speed on/off solenoid valves. Keles and Ercan [18]has shown that no stick-slip phenomenon is observed inhydraulic systems with slowly moving loads if they aredriven by pulse-modulated inputs.

In pneumatic systems, Morito et al. [19]and Noritsugu[20]implemented the PWM technique in pneumatic manip-ulators to control pressure and contact force. Ye et al. [21]

investigated a model for determining the maximum operat-

ing modulation ratio for pneumatic PWM solenoid valves.A novel PWM technique that was used to drive fast switch-ing solenoid valves was introduced by Varseveld and Bone[1] for accurate control of a pneumatic actuator. Theydescribed the minimum and the maximum duty cycle forPWM signals depending on the dead times of the valveswitching. Shih and Ma[5,6]used a combination of the slid-ing mode and modified differential PWM method to controlthe position of a pneumatic rodless cylinder with high speedsolenoid valves. The proposed controller was designed andimplemented in microcomputer experimentally.

The object of this study is to develop an inexpensivesingle stage solenoid valve to provide a fast switching timewith a considerably large amount of flow rates.

2. Description and design of the valve

The valve was designed as a solenoid operated onoff,2/2, spring return with a disc element. A disc type of mag-netic circuit is used for fast switching of the valve.

Fig. 1shows a cross-sectional view of one of the fabri-cated four prototype valves. The valve consists of two major

plates. The upper plate houses an electromagnetic core with

Nomenclature

Bg useful air gap flux density, f(x, i) (Wb/m2)

e applied input voltage (V)fpwm modulation frequency (Hz)

Fd external force (N)Fm magnetic force (N)Fprs pressure force affected on the disc (N)H magnetic intensity (A-turns/m)I current in the coil (A)M air mass into the chambers (kg)NI magnetomotive force (A-turns)Pdown absolute downstream stagnation pressure of the

valve (N/m2)Pup absolute upstream stagnation pressure of the

valve (N/m2)Qn nominal flow rate (m

3/s)Td, Tt disc delay and traveling time (s)

Tso, Tsc opening and closing switching time of the valve(s)

tp the valve switching pulse time (s)V volume of cylinder (m3)x motion of the disc (m)y piston position (m)dm/dt mass flow rate (kg/s)DP net pressure (N/m2)/ effective flux of the magnetic circuit (Wb)Q mean value of this flow oscillation (m3/s)MR modulation ratioAe effective area of the magnetic circuit

(256

10

6

m

2

)

Ai piston effective area (A1: 1256 106, A2:

1056 106 m2)Av effective cross-sectional area of fluid flow path

(12.5

10

6

m

2

)b damping coefficient (varies N/(m/s))Cd discharge coefficient (0.65)d diameter of nozzle (8.6 103 m)k spring coefficient of the valve (20000 N/m)L piston stroke (0.4 m)md total mass of moving elements (28 10

3 kg)N number of turns (120)Pcr critical pressure ratio (0.528)Patm atmosphere pressure (1 10

5 N/m2)Ps supply pressure (5 10

5 N/m2)Rc coil resistance (2.7 X)R ideal gas constant (287 J/kg K)

r radius of nozzle (4.3 103 m)T upstream stagnation temperature (293 K)Vtank volume of tank (1.8 10

3 m3)w land width of nozzle (1 103 m)xh disc travel distance (0.46 10

3 m)xt total air gap that includes holding gap and

traveling gap (0.6 103 m)c ratio of specific heat (=cp/cv) (1.4)l permeability of the air gap (4p107 H/m)

Index

i cylinder chamber index (1, 2)

366 E.E. Topcu et al. / Mechatronics 16 (2006) 365378


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a coil and provides a drain connection. The lower plateforms a disc chamber and provides a pressure connection.

The disc forms a seating type of valve and also consti-tutes a moving part of magnetic circuit. It moves betweena nozzle and a preloaded spring under resultant force of

magnetic, spring and flow forces. The diameter of the discchamber is slightly less than the diameter of the magneticcore to prevent the magnet being pulled into the chamberby magnetic force. A nozzle made of bronze is pressed intothe centre of the lower plate. The upper and lower platesare made of hard aluminum, and spring is made of alloyedsteel to reduce magnetic flux and remanence. The valveassembly is bolted together with four cap-headed screws.The valve plates and magnetic core are sealed with O-ringsas it is shown in Fig. 1.

The core around the coil and disc are made of SAE1008magnetic material. The diameter of core and disc are35.5 mm and 34 mm, respectively. The disc is 4 mm wideand has four stops core side and four 3 mm of drill holes.A 12 mm diameter and 1.5 mm thick nonmagnetic elasto-mer material is pressed in to the nozzle side of the disc toprovide a good seal. The coil consists of 120 turns of0.44 mm of isolated diameter copper wire on a polyimidespool. It is sealed in the magnetic iron core with epoxy.The epoxy seals two external wires to the coil and also pro-vides a smooth face on the disc chamber end of the coil.

Operation of the valve is shown in Fig. 2. The valve isarranged to be normally closed by the preloaded spring act-ing on the disc keeping the nozzle closed against the pres-surized flow path. When the coil is energized, magnetic

attraction force pulls the disc against spring force and the

valve opens and the fluid flows through the flow path ofthe valve as it is illustrated in Fig. 2. In fact, as the valvebegins to open, the flow force helps the magnetic force tokeep the valve open.

A position control of a pneumatic actuator with the four

2/2 valves configuration shown in Fig. 1can be realized asschematically shown inFig. 3. When four of the valves areclosed, the piston is kept at the required position. Whenvalve 1 and valve 3 are open, the piston moves forward.Similarly, when valve 2 and valve 4 are open, the pistonmoves backward. The electromagnet can be actuated usingpulse width modulation (PWM) technique with controlimplemented via analog circuits, digital logic circuits or amicroprocessor-based system.

The dimensions of the valve were chosen in order to sup-ply the required amount of air (460 l/min at standard pres-sure and temperature) for a cylinder diameter of 40 mmand for a pneumatic actuator of 400 mm of long. As it isillustrated schematically in Fig. 2, the disc-nozzle systemwas used in flow control of the valve. The necessary flowcontrol area is estimated as 12.5 mm2 for 460 l/min of nom-inal flow rate at 7 105 N/m2 of absolute supply pressureand 1 105 N/m2 of pressure drop. This is the circumferen-tial area between the disc and the nozzle, and it is equiva-lent to a nozzle diameter of 8.6 mm and a disc traveldistance of 0.46 mm. Here, the travel distance of the discis kept as small as possible in order to benefit from highmagnetic attraction force.

The close position of the spring kept valve is determinedby the maximum flow force acting on the disc type of seat-

ing element. The maximum fluid force equals to the force

Fig. 1. Cutaway view of prototype valve.

E.E. Topcu et al. / Mechatronics 16 (2006) 365378 367


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exerted by the static pressure, which, in turn, can be esti-mated from the pressure on the valve in close positionand the area acted on by this pressure.

The proposed valves are intended for use as PWM con-trolled solenoid valves, which are generally composed of aPWM circuit, a power amplifier, and solenoid valves. The

PWM circuit can be realized with an analog circuit drivenby microprocessor based software or a hybrid analog-digital circuit. A simplest analog PWM generator canbe realized using a comparator to modulate a sawtoothwave generator. Generally, PWM signal and modulatedwave are periodic with the same period. The ratio of the

Fig. 2. Operation principle of the valve.

Displacement Sensor

aM bM

..

Ps PsPatm atmP1 2 3 4

P1, V 1, T 1 P2 , V2 , T 2

Fout

y(t)

mA B

PowerAmplifier

DAC PC

ADC

PWMCircuit

Input

command

Fig. 3. Electro-pneumatic position control system with four prototype valves.

R2R1

R3 R4

+12 V+24 V

R5

C1

R7

R6

(47)

Sensing

Resistance

(0.1)

IGBT(IXGH24NCD1)

C2

(100 nf 250 VAC)

R8

(1003W)Valve

BJT2

(BC237B)BJT1

(BC237B)Buffer

PWM

Signal

Optocoupler

R7

DSEI12

Fig. 4. Power output stage with gate driver.



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time the on state to PWM period is defined as the dutyratio or pulse width ratio. Regardless type of PWM signalgenerator, a change of the signal applied to the input of thePWM generator produce a change of pulse width ratio ofthe output of PWM generator output.

In this project, digitally generated PWM signals

obtained using MATLAB/Real Time Workshop environ-ment with an Adventech PCL-818H data acquisition boardare used to drive the valves through power amplifiers. TheIGBT type of power amplifiers shown in Fig. 4were man-ufactured and used to drive the valves. IGBT types of tran-sistors are preferred due to their isolation against highvoltage and current and it is appropriateness for drivingwith PWM signals. In fact, IGBTs are very high frequencyswitching elements in comparison with the switching fre-quency of the valve. For the experimental test works, inaddition to use of digital PWM signals, an analog PWMsignal generator is designed and constructed with mainlywell known 555 Timer elements.

The main component of power output stage is BJT,MOSFET or IGBT type of transistors. MOSFET or IGBTtypes of transistors work at high current and they can bealso driven by pulsed width modulated signals. In this studyIGBT type of transistors are preferred due to their isolationagainst high voltage and current. Fig. 4 shows an IGBTtype of the power output stage with the gate driver.

3. Mathematical model of the valve

The valve was designed as a command element for pro-viding fast air flows through chambers of pneumatic actu-

ator. The characteristic of the valve is governed by a coupleof electromechanical and flow equations. The relationshipsof equations are strongly coupled with each other. Thevalve can also be considered as an electro-fluid converter.The block diagram given in Fig. 5 shows the signal flowof the electro-fluid conversation.

The main characteristic equations of the electro-pneumatic valve system can be derived as follows.

3.1. Electromagnetic sub-system

This sub-system consists of an electrical and magnetic

circuit. The electrical circuit is the actual coil which isrepresented with an ideal inductance L in series with aresistance Rc of the actual coil. Thus the inputoutput

equation for the electromechanical sub-system can bewritten by applying Kirchhoffs law:

et Rcit dN/tdt

1

The magneto-motive force is given by the number of

ampere turns NI, which must provide the flux requiredfor given force and displacement. Therefore, the followingequation may be written[22]:

NI 2Bgxtl

X

Hili 2

where 2Bgxt

l represents the magneto-motive force necessary

to establish the flux (in the air gap at a density ofBgacrossan air gap of length xt) and

PHilirepresents the magneto-

motive force necessary to establish the flux in the iron partsof circuit and Bg= f(x, i) is the useful air gap flux density(Wb/m2), xt is the total air gap that includes holding gapand traveling gap (m).

The magnetic attraction force generated as a result ofmagnetic flux, /(t) can be expressed as

Fm /2

Ael B

2gAe

l 3

Furthermore, the magnetic flux /(t) is a function of cur-rent i, disc movement x (/=/(i(t), x(t))). Then, it can beshown that the magnetic force equation can be written as:

Fm lAeNI2

4xtx2 4

where Ae is effective cross-sectional area of the magneticcore.

3.2. Mechanic subsystem

Referring toFig. 6and applying Newtons second law ofdynamic to the moving elements, the equation of motionfor the valve disc can be written as

Fmt m d2xtdt2

b dxtdt

kxt Fdt Fprs 5

where m is the mass of the disc (kg), b is the damping

coefficient (N/(m/s)), kis the spring coefficient of the valve(N/m), Fd is the external force (N) and Fprs(N) is the pres-sure force affected on the disc (N).

ElectricalCircuit

MagneticCircuit

MechanicalCircuit

FluidCircuit

Disc

m,b,k

FluidCoil

L,R

MagneticCircuit(t)Electrical

Signal

Electromechanical Subsystem Mechanic-Fluid Subsystem

Fluid

Signal

e(t) i(t) Fm(t) x(t)(t)M

dx/dt

Fig. 5. Block diagram of mathematical model of the valve.



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3.3. Fluid subsystem

The electro-mechanical part of the valve is used to con-trol the fluid flow through the valve orifice. As the fluid inthe valve and in the pneumatic system is compressible,when the upstream to downstream pressure is larger thanthe critical value Pcr, the mass flow depends nonlinearly

on both pressures. Whereas, when the pressure ratio issmaller than Pcr, the flow attains sonic velocity (chokedflow) and depends linearly on the upstream pressure.

The standard equation for mass flow rate through anorifice of area Av is [23,25]

dm

dt

0:0405CdAvPupffiffiffi

Tp for sonic flow

where PdownPup

6 0:528

CdAvPuffiffiffiT

p 2cRc1

PdownPup

2c Pdown

Pup

c1c

1=2for subsonic flow

where PdownPup

> 0:528

8>>>>>>>>>>>>>>>:

6where Av is the effective cross-sectional area of fluid flowpath (m2), Cdis the discharge coefficient, Pupis the absoluteupstream stagnation pressure of the valve (Pa), Pdownis theabsolute downstream stagnation pressure of the valve (Pa)and finally Tis the upstream stagnation temperature (K).

The variable control area of the valve orifice is given bythe disc position relative to the nozzle and circumference ofthe nozzle as it is shown in Fig. 6. The circumferential areaof the nozzle-disc system can be expressed as

Av pdxh 7where d is the diameter of nozzle (m), xh is the disc traveldistance (m).

The static fluid force acting on the disc is critical fordetermining of the maximum system pressure and springconstant. This force is equal to maximum fluid force andit can be estimated as

pr2DP6 F 6 pr w2DP 8For a very small disc displacement (motion), the flow

force can be obtained from the Navier Stokes equations as

F pr w2 r2

2 ln1 w=r DP 9

where w is the land width of nozzle (m) and r is the radius

of nozzle (m).

4. Mathematical model of the cylinder chamber

The governing equation of the pressure dynamics of thecylinder chambers can be derived by analyzing the thermo-dynamics of the system through energy conservation andcontinuity equations.

Referring toFig. 3and considering the control volume

V, with density q, mass m, pressure P, and temperatureT, the ideal gas law can be written as

PiVi miRTi 10where i= 1,2 is the cylinder chamber index and Pi, Vi, miand Ti are respectively the pressure, the control volume,the mass and the temperature on the either side of the piston.

Differentiating and then rearranging the above equationwith energy equation and the heat transfer terms yields [26]

dPidt

cRTVi

dmidt

cPiVi

dVidt

11

Choosing the origin of the piston displacement at the

middle of the stroke, the volume of each chamber can beexpressed as shown in Eq.(12)

Vi V0i Ai 12Ly

12

whereV0iis the dead volume at the end of stroke and con-necting ports (m3), Aiis the piston effective area (m

2), L isthe piston stroke (m), and yis the piston position (m).

By substituting Eq.(12)and its time derivative into Eq.(11), the time derivative for the pressure in the pneumaticcylinders becomes

dPi

dtc

RT

V0i Ai L2 y

dmi

dtc

PiAi

V0i Ai L2 y

dy

dt

13where dmi/dtis the mass flow rate into the chamber takenfrom the valve ports. The leakage between the chamberswas neglected in Eq.(13)by assuming rubber seals are usedfor regular pneumatic cylinder. But the leakage can besignificant for low friction cylinders that have graphite orTeflon seals. In the case of a low friction cylinder, the termofminmoutmust be replaced with miin Eq.(13). Heremoutis the leakage between the chambers.

The general form of the pressure dynamic (i.e. Eq. (13))

can be adopted for the charging and discharging processes

Fig. 6. Free body diagram of the moving element of the valve.



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of the cylinder chambers or any other control volumes.When Vi is the constant at the case of a constant volumetank, then the pressure rate equation can be shown as

dPidt

cRTVi

dmidt

14

5. Switching characteristics of the valve

In order to determine static and dynamic characteristicsof the valve, the equations of the mathematical modelwere solved numerically with MATLAB and Simulinkenvironment.

Fig. 7shows the static characteristics of the magnetic cir-cuit of the valve. The resultswere obtained by the solution ofEqs.(2)(4)with a series flux and permeability relationships[22]. The valve spring characteristics were added to theresults in order to determine the current required to switchthe valve and find an optimum working distance of the disc

element. A 0.6 mm of total gap with 0.14 mm of holdinggap and 0.46 mm of travel distance has been chosen as anoptimum working gap. With this working gap, an appliedcurrent of 6 A or higher provides a sufficient magnetic forceto overcome the spring force at the beginning of switching ofthe valve. At the end of switching, about 1.75 A of appliedcurrent is just enough to hold the disc against to the springforce with 0.14 mm of holding gap. Henceforth the currentcan be reduced to a holding level after the end of switchingso that the consumed energy is reduced. This equallyimproves the switching time of the valve closing.

Fig. 8 shows the flow characteristics of the valveobtained from solution of Eq.(6).

The switching time of the valve is governed by the elec-tromechanical part of the valve. The dynamic characteris-tics of the valve based on the nonlinear equations (1)(5)are obtained from a simulation program developed on

the Simulink environment. The program is based on per-manence formulae and leakage coefficient of dimensionedmagnetic circuit and simulation starts to estimate the fluxdistributions with the data taken from BHcurve of thematerial used in the manufacturing of the valve magnet.

Typical switching time characteristics of the valve areshown in Fig. 9. The current in the coil approximatelyexponentially increases until the magnetic attraction forceexceeds the spring force. The disc will start to move andequally the valve starts to switch at this moment which isdenoted as the disc time delay, Td. The time delay is gov-erned by the force balance of the spring force, pressureforce and magnetic force. In fact, the pressure force helps

the magnetic force to open the valve. As the disc starts tomove, the inductance of the coil starts to increase due tothe decreasing gap between the disc and the magnet. Asthe gap decreases, the effective time constant increases,thereby causing the current to decrease to its local mini-mum when the disc reaches its final position. The time per-iod between the delay time and the time when the discreaches its final position is denoted as the disc travel time,Tt. The total opening time of the valve, Tso, therefore con-sists of the disc delay time, Td, and the disc travel time,Tt.The simulation results show that the total opening time of

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 10

25

50

75

100

125

150

175

200

225

250

275

300

Displacement (mm)

Magne

ticForce(N)

1. i=0.5 A2. i=1.25 A3. i=1.75 A4. i=3.0 A5. i=4.0 A6. i=5.0 A7. i=6.0 A8. i=7.0 A

1

2

3

4

5

6

7

Spring characteristics

Holding Gap Travelling Gap

Total Gap

8

Fig. 7. Static characteristics of the magnetic circuit of the valve.



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the valve is approximately 4.65 ms, and the disc delay timeis 3.2 ms, which is 69% of the total switching time.

The closing time of the valve is as critical as the openingtime of the valve in selection of the cycle time of PWM sig-nal. The valve starts to close when the coil is de-energized.The total closing time of the valve is similar to the opening

time of the valve and is defined as the sum of the closingdelay time,Tdcand the disc travel time of the valve closing,Ttc. Figs. 9 and 10 show the definition of the switchingtimes for a total switching pulse or a switching cycle. Theclosing delay time is governed by the force balance betweenthe spring force, the fluid force and the magnetic holding

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 10

1

2

3

4

5

6

7

8

9x 10-3

Pdown/Pup

Mass

flow

rate(kg

/s)

Pup:7*105 Pa

Pup:6*105 Pa

Pup:5*105 Pa

Pup:4*105 Pa

Pup:3*105 Pa

Pup:2*105 Pa

Pup:1*105 Pa

Fig. 8. Flow characteristics of the valve.

0 0.005 0.01 0.015 0.02 0.025 0.030

1

2

3

4

5

6

7

8

9

10

Time (s)

Current(A),Input(V)

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1x 10

-3

Displace

ment(m)

Current

Input

Tso

Td

Tsc

Tdc

Tt

tp

Displacement

Ttc

Fig. 9. Switching time characteristics of the valve.



9/14

force. The current and holding force decay in the de-energized magnet are not instantaneous but they follow atransient state. This retards the beginning of the motionof the disc depending on the level of holding current. Thesimulation and experimental results show that the closingtime delay can be reduced as much as about a half of the

maximum time delay. Thereby the holding current can bereduced to a level of 25% of the maximum applied current,which is just enough to hold the disc against the springforce. The disc travel time is almost the same with thetravel time of the valve opening.

6. PWM characteristics of the valve

PWM control method is used in fluid power systemsto obtain a linear relationship between the fluid flow andan input signal with on/off switching solenoid valves. Toachieve a full linear relationship with an on/off valve, the

switching action of the ideal valve must be instantaneous.Under ideal switching conditions, the volume of fluidpassing through an ideal valve is a function of nominalflow rate, Qn and the valve switching pulse time tp

VpZ tp

0

Qndt Qntp 15

As the valve is opened periodically for a certain timeinterval tp, a series of volume increment, Vp is obtained.It can be shown[24]that the mean value of this flow oscil-lation Q is a relation between rated flow, time of valveopening and the switching frequency, fpwm

Q Vpfpwm Qntpfpwm 16

The rated flow of the ideal valve is valid for the actualvalve under following conditions: The change of flow hasto be identical for each on and off switching operationi.e. Tso=Tsc. In this case, the area integral of the flowcurve has the same amount as that of the ideal valve. Allvariables of the valve must reach their steady state between

consecutive valve opening and closing. The valve switchingtime, Tscan be defined as the time when all variables havereached their steady state. From this condition, the rela-tionship between the valve switching time, Ts switchingpulse, tp and modulation frequency, fpwm or modulationperiod, Tpwmcan be expressed as

Ts 6 tp 6 Tpwm Ts or Tpwm P 2Ts 17Furthermore, the opening switching time Tso and clos-

ing switching time Tsc of the solenoid-operated springreturn valve are not identical which reduces linearity moreseverely. In this case, the following switching condition

must be satisfied in order to ensure the on/off switchingoperation of the valve can be executed completely

Tso 6 tp 6 Tpwm Tsc or Tpwm P Tso Tsc 18As the modulation ratio, MR relating to the switching

pulse, tpto the modulation period, Tpwmis MR =tp/Tpwmthen from Eqs.(17) and (18), the minimum and maximumoperating modulation ratio of a PWM modulated valvecan be shown[1] respectively as

MRmin TsoTpwm

19

MRmax

1

Tsc

Tpwm 20

0 0.005 0.01 0.015 0.02 0.025 0.03 0.035 0.040

1

2

3

4

5

6

7

8

Time (s)

Current(A),InputSignal(V

)

0

1

2

3

4

5

6

7

8x 10

-4

Displacement(m)

Current

Displacement

Holding input

Switching

input

Fig. 10. Effects of holding current on switching time.



10/14

The minimum and the maximum modulation ratiosensure where the valve still response at the switching pulsetime for respectively opening and closing of the valve. Theminimum modulation ratio indicates the dead zone on therelationship between the modulation ratio and the flowcharacteristics of the valve output. On the other hand,

the maximum modulation ratio indicates the end of theproportional range between the modulation ratio of inputand flow output of the valve.

The proportional band of the modulation ratios wherethe valve operates as quasi-linear can be expressed as

Mpb 1 2TsTpwm

or Mpb 1 Tso TscTpwm

21

Eqs.(19)(21), indicate that the smaller the valve open-ing and closing times the greater is the PWM frequency,fpwm(=1/Tpwm) with the greater proportional band, Mpb.This can be achieved with a fast switching valve.

The discrepancy between the ideal and the real condi-tions of the valve introduces a dead zone on the relationshipof the duty ratio of the valve input and flow characteristicsof the valve output as it is shown inFigs. 9 and 10. To over-come the dead zone and improve the linearity several differ-ent PWM methods such as D-PWM; M-D-PWM have beenoffered several authors[5,9,17,20].

In this study, an overdriving pulse technique is intendedto use with newly developed PWM methods[1,2]which aredifferent to existing methods in order to improve the linear-ity and to reduce the dead zone. The simulation and theexperimental results have shown that the electrical time

delay, Te(=L(x)/R) constitutes the main portion of the

delay in the switching time of the valve. When a highamplitude pulse driving high current is applied to the coilof the valve, the electrical delay and hence the total openingtime of the valve is reduced. Furthermore, after the open-ing of the valve is completed the current can be reducedto a holding level that is just enough to hold the disc which

opens the orifice. When the valve is switched off, the hold-ing current drops more rapidly and hence the magneticretarding force. Therefore, the delay in the valve closingtime reduces. With this adjustment, even the switching onand switching off the valve can be made identical.

7. Experimental tests

Experimental tests were carried out to investigate staticand dynamic characteristics of the valves. The switchingconditions of the valves were investigated through seriesof the tests. The data of dynamic tests were controlledand recorded with an Adventech PCL-818H data acquisi-

tion board. An opto-couple type of position sensor was fit-ted the valve to detect the position of the disc movementwith no load conditions. The movement of the disc is trans-ferred to the opto-couple with a rod connected to the disc.The changes in the current of the coils in the valves weremeasured through the sensing resistance on the power unit(seeFig. 4).

The experimental results of the current history for fourdifferent valves are shown in Fig. 11. For comparison,the simulation results are added to Fig. 11. As it can beseen in these results, the switching times of the valves arevery close to each other. In addition, the simulation results

show a good agreement with the experimental results.

0.6 0.602 0.604 0.606 0.608 0.61 0.612 0.6140

1

2

3

4

5

6

7

8

9

Current(A)

Time (s)

1

234

5

1,2,3,4: experimental5 : simulation

Fig. 11. Experimental current histories of valves.



11/14

Fig. 12 shows typical characteristics of current historyand the history of the disc motion of one of the valves whena high voltage pulse is applied to the coil of the valve underno load conditions. From the characteristics curves, theswitching time of the valve is determined using the localminimum in the current curve together with the curve ofthe disc motion. The minimum indicates the end of the discmotion when it is evaluated with the curve of the discmotion. The dead time of the valve switching is measuredas the time between the beginning of the valve switchingand starting of disc motion. Various switching conditionsof the valves were tested under different electrical controlof the valve.

The dynamic switching characteristics of the valvesunder pressure were determined with a test setup shown

inFig. 13. The outlet pressure history of the valves is mea-sured by the transducer 1 for determining the switchingtime of the valves under pressure. The pressure transducer1 is fitted to downstream side of the valve as close as pos-sible in order to avoid a flow transition lag. The totalswitching time of the valve is determined from the currenthistory curves and pressure history curves and the deadtime is determined from the pressure curves.Fig. 14showstypical switching characteristics of the valve under pres-sure. The opening time of the valve under pressure is about3 ms which is much less than comparing with 4.5 ms of theopening time of the valve with no load. This is expected asthe pressure force helps to open the valve. On the otherhand the closing time of the valve under pressure increasescomparing with the closing time of the valve with no load.

0 0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.10

1

2

3

4

5

6

7

8

Current(A),InputSignal(V)

0 0.050

1

2

3

4

5

6

7

8x 10

-4

Time (s)

Displacement(m)

Current

InputDisplacement

Fig. 12. Switching characteristics of valve under no load condition.

Prototype

valve

Pressure

Sensor 2

Patm

Psupply

PowerAmplifier

DAC PCPWMCircuit

ADC

Pressure

Sensor 1

Inputcommand

Fig. 13. Experimental setup for switching of valves under pressure.



12/14

Nevertheless the total switching period (opening time plusclosing time) of the valve remains almost same which isabout 9.5 ms.

One extreme switching condition of the valves is toapply a very high voltage pulse which drives a maximumcurrent through the coil of the valve and keeps the currentas long as the valve is open. This condition provides a min-imum switching time of the valve opening. But, on theother hand, it increases the switching time of the valve clos-ing due to high holding current. As the current decreasesexponentially it cannot drops instantaneously. Hence, itretards the closing of the valve.

One other extreme switching condition is to apply a lowvoltage input pulse just enough to switching and hold toopen position the valve. This increase the switching timeof the valve opening but it reduces the time delay in thevalve closing.Table 1shows the switching times for variousswitching conditions.

0 0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.10

1

2

3

4

5

6

7

Time (s)

Curren

t(A),Pressure

(Pa

),Inpu

ts

ign

al(V)

Input pressure of valve (105Pa)

Output pressure of valve (105Pa)

Current (A)

Input signal (V)

Fig. 14. Switching characteristics of valves under pressure.

Table 1Switching times for various working conditions

Workingcondition

Openingtime tso(ms)

Closingtime tsc(ms)

Various applied current values

9.6 A (24 V) 4.5 8.5

8.0 A (20 V) 5.5 5.76 A (16 V) 8.0 4.5 A (12 V) 16.5

Without the reducedholding current (no load)

4.5(Theoretical:4.65)


With the reducedholding current (no load)



With the reduced holding current(under 7 105 N/m2 pressure)

3.0 6.0

Prototype

valve

TANK

PsupplyPatm

Pressure

Sensor

Power

Amplifier DAC PC

ADC

PWM

Circuit

PressureSensor

Input

command

Fig. 15. Experimental setup for determination steady PWM characteristics of valves.



13/14

PWM controlled flow characteristics of the valves aretested on a closed volume tank as it is shown in Fig. 15.In the test setup, while one of two valves is accommodatedto pressurize the tank and the other one is accommodatedto ventilate the tank to atmosphere. The smallest pulseduration of the PWM signal is chosen at least equals to

the switching time of the valve opening. While a PWM sig-nal with certain pulse duration is applied to the valve in thepressurized side of the tank, the inverse of the PWM signalis applied to the ventilation side of the tank. The PWM fre-quencies driving the valves are various between 10 Hz and

50 Hz. The steady state values of the pressure in the tankwere measured for various pulse durations and the resultsare shown inFig. 16for dimensionless pressure change ver-sus pulse duration ratios.

During the tests, the opening time of the valves weremeasured as 4 ms and the closing time of the valves were

measured 6 ms. Then the maximum and minimum operat-ing modulation ratios (Eqs. (19) and (20)) were chosenaccording the above results. As it can be seen fromFig. 16, quite good linearity between the valve input andthe valve output are obtained up to 33 Hz of PWM

0 10 20 30 40 50 60 70 80 90 100

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

Md=td/Tc

P/P

up

--x--:fc=10 Hz--+--:fc=20 Hz-- --:fc=33 Hz--o--:fc=50 Hz

Fig. 16. PWM characteristics of the valves.

0 0.1 0.2 0.3 0.4 0.5 0.60

0.05

0.1

0.15

0.2

0.25

0.3

Time (s)

Displacement

(m)

Fig. 17. Step response of position control system.



14/14

frequency. When the PWM frequency is increased up to50 Hz the linearity deteriorates due to the valve dynamicscannot follow the input changes.

To evaluate the feasibility of the valves driven withPWM on a pneumatic position control system, stepresponse tests were carried out on the system given in

Fig. 3. The PWM frequency applied to the valves was35 Hz and a pneumatic cylinder of 400 mm of long wasmoved up to mid-position. Fig. 17 shows the resultsobtained with a simple PI control tuning of the system.As it can be seen from the figure very good position ac-curacy is obtained.

8. Conclusions

A simple fast switching valve controlled with PWM forapplications of pneumatic position control has been inves-tigated. Four prototype valves have been built and thebasic mode of operation confirmed. In addition, also elec-

tronic circuits for driving the valves and for PWM controlof the valves have been designed and constructed as a com-plete mechatronic system.

The switching and PWM characteristics of the valvesboth theoretically and experimentally have been investi-gated. The theoretical and experimental results have shownthat the opening time of the valves were as fast as 3 ms andthe closing time of the valves was 6 ms under 7 105 N/m2

of supply pressure. PWM controlled valves provide a goodlinearity up to 33 Hz.

It has been demonstrated that the switching speed of thevalves can be enhanced by the application of overdriving

current to the coils during the switching action and thenby reducing the applied current to a holding current levelafter the switching is completed. With this way, the electri-cal power required to hold the close after switching is com-pleted can be reduced about 6% of the peak power and theclosing time of the valves can be reduced about 50% of thehighest closing time.

Simulated results of the valves dynamics were in agree-ment with the experimental results, and thus the validityof the proposed mathematical model was confirmed. Asthe valves developed in this study have a simple construc-tion and provide higher flow rates (460 l/min) as single stagevalves than conventional servovalves and fast switchingvalves, they should be economical to use in pneumatic posi-tion control systems. The work in this area is continuing.

References

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[2] Gentile N, Giannoccaro NI, Reina G. Experimental tests on positioncontrol of a pneumatic actuator using on/off solenoid valves.Industrial Technology, IEEE ICIT 02, vol. 1, Bangkok, Thailand,December 2002. p. 5559.

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[16] Muto T, Yamada H, Suematsu Y. Digital control of hydraulic servosystem operated by differential pulse width modulation. JSME Int J1993;36(4):618.

[17] Suematsu Y, Yamada H, Tsukamoto T, Muto T. Digital control ofhydraulic actuator system operated by differential pulse widthmodulation. JSME Int J 1990;33(4):6418.

[18] Keles O, Ercan Y. Theoretical and experimental investigation of apulse-width modulated digital hydraulic position control system.Control Eng Pract 2002;10:64554.

[19] Morito YS, Shimizu M, Kagawa T. An analysis on pneumatic PWMand its application to a manipulator. In: Proceedings of theinternational symposium on fluid power control and measurement,Tokyo, 1987. p. 38.

[20] Noritsugu T. Pulse-width modulated feedback force control of apneumatically powered robot hand. In: Proceedings of the interna-tional symposium on fluid power control and measurement, Tokyo,1987. p. 4752.

[21] Ye N, Scavarda S, Betemps M, Jutard A. Models of a pneumaticPWM solenoid valve for engineering applications. Trans ASME1992;114:6808.

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Development of Electro-pneumatic Fast Switching Valve And