JP2002346450A - Fluid ejection device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a liquid ejecting apparatus which can start and stop to eject, at a high speed and without compression and destruction of powder, various powdered materials such as adhesive, clean solder, fluorescence material, grease, paint, hot melt, chemicals and food, in the production process of the field of the electronic parts, the electric appliances and the like. SOLUTION: The liquid ejecting apparatus is provided with a driving device in the axial direction respectively driving a piston and a sleeve in almost reverse phase, and with this driving device in the axial direction, the hydrodynamics resistance is increased or decreased by changing the gap of the flow channel formed between a sleeve and a housing, and the device is constituted so that, when the dynamic characteristic compensational parameter the following formula (1) is defined as ϕ (no dimension), supposing that the sectional area and the displacement of the piston are ap , Xp, and the sectional area and the displacement of the sleeve are as , Xst , ϕ>=0.15 in the formula (1).

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention can be used in production processes in the fields of electronic parts, home appliances, etc., and can be used for various liquids such as adhesives, cream solders, phosphors, greases, paints, hot melts, chemicals, foods, etc. The present invention relates to a fluid control device and a control method for discharging a fixed amount of fluid.

[0002]

2. Description of the Related Art Liquid ejecting apparatuses (dispensers) have been used in various fields, but with the recent demand for miniaturization and high recording density of electronic components, a minute amount of fluid material can be dispensed with high precision. There is a demand for a technique for stably controlling discharge.

There is also a great demand for developing a new fluid application means for uniformly applying a phosphor on a display surface of, for example, a CRT or PDP.

In the field of surface mount (SMT), for example,
In the trend of high-speed mounting, miniaturization, high-density, high quality, and unmanned mounting, the issues of dispensers can be summarized as follows: high accuracy of application amount and miniaturization of one application amount. Shorter: Can be used for high-viscosity powder fluids that can shut off and start high-speed discharge Conventionally, a dispenser of an air pulse type, a screw groove type, a micropump type using an electromagnetic strain element, or the like has been put to practical use in order to discharge a very small amount of liquid.

[0005]

Among the prior art examples described above, a dispenser using an air pulse method as shown in FIG. 8 is widely used.
The technology is introduced in “3.25, Issue 7” and the like. The dispenser according to this method causes a constant amount of air supplied from a constant pressure source to be applied to the container 200 (cylinder) in a pulsed manner, and causes a constant amount of liquid corresponding to a rise in the pressure in the cylinder 200 to be discharged from the nozzle 201. Things.

[0006] The dispenser of the air pulse type has a drawback that response is poor.

[0007] This disadvantage is due to the compressibility of the air 202 sealed in the cylinder and the nozzle resistance when passing the air pulse through a narrow gap. That is, in the case of the air pulse method, the cylinder volume: C and the nozzle resistance: R
The time constant of the fluid circuit that can be achieved: T = RC is large, and after the input pulse is applied, a time delay of, for example, about 0.07 to 0.1 seconds from the start of ejection must be expected.

In order to eliminate the drawbacks of the air pulse method, a needle valve is provided at the inlet of the discharge nozzle.
A dispenser that opens and closes a discharge port by moving a small-diameter spool constituting the needle valve at high speed in an axial direction has been put to practical use.

However, in this case, when the fluid is shut off, the gap between the relatively moving members becomes zero, and the powder having an average particle diameter of several to several tens of microns is mechanically pressed and destroyed. Due to various inconveniences that occur as a result, it is often difficult to apply the method to application of an adhesive material, a conductive paste, or a phosphor mixed with powder.

For the same purpose, a screw groove type dispenser which is a viscous pump has already been put to practical use. In the case of the screw groove type, a pump characteristic that is hardly dependent on the nozzle resistance can be selected, so that a preferable result can be obtained in the case of continuous spraying, but the intermittent application is not good in terms of the characteristics of the viscous pump. Therefore, in the conventional screw groove type, (1) an electromagnetic clutch is interposed between the motor and the pump shaft, and this electromagnetic clutch is connected or released when the discharge is turned on or off.

(2) Use a DC servomotor to start or stop rapid rotation.

However, since the response is determined by the time constant of the mechanical system in any of the above cases, the high-speed intermittent operation is limited. The responsiveness is better than that of the air pulse method, but the shortest time is about 0.05 seconds.

In addition, since there are many uncertainties in the rotation characteristics of the pump shaft during transient response (at the start and stop of rotation), it is difficult to strictly control the flow rate, and the coating accuracy is limited.

For the purpose of discharging a small amount of fluid, a micropump using a laminated piezoelectric element has been developed. For this micro pump, a mechanical passive discharge valve or suction valve is usually used.

However, in the above-described pump which is constituted by a spring and a ball and opens and closes a discharge valve and a suction valve by a pressure difference,
It is very difficult to block and open a high viscosity rheological fluid of tens of thousands to hundreds of thousands of centipoise with low flowability at high flow rate accuracy and at high speed (0.1 seconds or less).

In the field of circuit formation, which is becoming increasingly more precise and ultra-fine in recent years, or in the fields of forming electrodes and ribs for picture tubes such as PDPs and CRTs, applying sealing materials for liquid crystal panels, and manufacturing processes for optical discs and the like. There are strong demands for fine coating technology as follows.

After continuous spraying, application can be stopped quickly and continuous application can be started rapidly after a short time. For that purpose, it is ideal that the flow rate can be controlled, for example, on the order of 0.01 second.

[0018] Capable of handling powder fluid. For example, there should be no troubles such as powder compression breakage or flow path clogging due to mechanical cutoff of the flow path.

In order to respond to various recent demands related to the above-described minute flow rate application of a high-viscosity fluid / powder fluid, the present inventors provide a relative linear motion and a rotational motion between a piston and a cylinder. At the same time, the present invention proposes a coating method in which a fluid transport means is provided by rotating motion, a relative gap between the fixed side and the rotating side is changed using linear motion, and an outflow amount is controlled. (Japanese Patent Application No. 2000-188899).

The present invention aims to further improve the above proposal based on the result of a strict theoretical analysis. For example, it is possible to cut off the flow rate at a high speed from the state of continuous coating or to steeply start continuous coating. The purpose of the present invention is to provide a configuration condition of the pump section.

[0021]

According to the present invention, there is provided a fluid supply device comprising: a sleeve for accommodating a piston; axial driving means for driving the piston and the sleeve in substantially opposite phases; and a housing for accommodating the sleeve. A fluid transport chamber formed by the sleeve and the housing;
In order to increase or decrease the fluid resistance between the fluid transport chamber and the discharge port, and the fluid drive chamber, the axial drive means is used to increase or decrease the fluid resistance between the fluid transport chamber and the discharge port. While the gap of the flow path formed between the sleeve and the housing is configured to change, the sectional area and displacement of the piston are ap, Xp, and the sectional area and displacement of the sleeve are as, Xst. , Below (1)
When the parameter of the equation: φ (dimensionless) is defined,

[0022]

(Equation 2)

It is configured such that φ ≧ 0.15.

[0024]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Hereinafter, a first embodiment in which the present invention is applied to a dispenser for surface mounting electronic components will be described.
This will be described with reference to FIG.

The structure of the dispenser of this embodiment is roughly composed of a drive section and a pump section.

The drive unit has two output shafts (double pistons) using the output displacement of both ends of the giant magnetostrictive rod, and utilizes the characteristics of a giant magnetostrictive element capable of supplying electric power in a non-contact manner, thereby using a motor to control the motor. The output shaft is further provided with a rotation function.

The pump unit includes means for transporting the fluid by rotation of the output shaft, and flow rate control for controlling the discharge amount of the fluid by changing the relative gap between the fixed side and the rotation side using linear movement of the output shaft. It has a structure having both of the means.

Hereinafter, the driving section will be described first. This drive unit is composed of members 1 to 28.

In FIG. 1, reference numeral 1 denotes a direct-acting actuator (axial driving means) using a giant magnetostrictive element. Reference numeral 2 denotes a movable sleeve driven by the actuator 1, reference numeral 3 denotes a rotary sleeve for housing the movable sleeve 2 on the front side, and reference numeral 4 denotes a housing for housing the actuator 1.

Reference numeral 5 denotes a cylindrical giant magnetostrictive rod made of a giant magnetostrictive material. The giant magnetostrictive rod 5 is fixed between the upper yoke 8 and the movable sleeve 2 also serving as a yoke material, with the bias permanent magnets (A) 6 and (B) 7 sandwiched vertically.

Reference numeral 9 denotes a magnetic field coil for applying a magnetic field in the longitudinal direction of the giant magnetostrictive rod 5, and reference numeral 10 denotes a cylindrical yoke which is housed in the housing 4.

The permanent magnets A and B apply a magnetic field to the giant magnetostrictive rod 5 in advance to increase the operating point of the magnetic field.
The closed loop magnetic circuit for controlling the expansion and contraction of the giant magnetostrictive rod 5 is formed by 6 → 8 → 10 → 2 → 7 → 5. That is, the members 5 to 10 constitute the giant magnetostrictive actuator 1 which can control the expansion and contraction of the giant magnetostrictive rod in the axial direction by the current applied to the magnetic field coil. The giant magnetostrictive material is an alloy of a rare earth element and iron. For example, bFe ₂ , DyFe ₂ , SmFe _{2 and the} like are known, and their practical use has been rapidly advanced in recent years.

Reference numeral 11 denotes a piston fixed to and integrated with the upper yoke 8 (magnetic material), and is provided through the giant magnetostrictive rod 5. The piston 11 is in close contact with the upper end of the giant magnetostrictive rod 5 with the bias permanent magnet (A) 6 sandwiched therebetween. The piston 11 also extends upward, and is supported at its upper end by a bearing 13 provided on the upper sleeve 12. The upper sleeve 12 is rotatably supported by an intermediate housing 15 with a ball bearing 14 interposed therebetween.

Like the piston 11, the movable sleeve 2 is in close contact with the lower end of the giant magnetostrictive rod 5 with the bias permanent magnet (B) 7 sandwiched therebetween.

Between the upper yoke 8 and the upper sleeve 12, and between the movable sleeve 2 and the rotary sleeve 3, a bias spring 1 for applying a mechanical axial pressure to the giant magnetostrictive rod 5 is provided.
6, 17 are provided.

With the above configuration, when a current is applied to the electromagnetic coil 9 of the giant magnetostrictive element, the giant magnetostrictive rod 5 expands and contracts in proportion to the magnitude of the applied current.

At this time, the displacement of both ends of the giant magnetostrictive rod 5 is determined by the rigidity of the two bias springs 16 and 17 for applying a preload in the axial direction to the giant magnetostrictive rod. If the stiffnesses of the bias springs 16 and 17 are equal, both ends of the giant magnetostrictive rod extend with an equal displacement with respect to the input current.

When the stroke between the movable sleeve 2 and the piston 11 is changed, the rigidity ratio between the two bias springs 16 and 17 may be changed. If the rigidity of one of the bias springs is made infinite, that is, if one of the springs is omitted and the two components (8 and 12, or 2 and 3) that interpose the spring are integrated, the axial output shaft becomes Become one.

Reference numerals 18 and 19 denote fastening bolts, and reference numeral 20 denotes a motor for imparting a rotational movement to the piston 11.
Servo motor is used, 21 is a motor rotor, 22
Is a motor stator, and 23 is an upper housing for housing the motor stator. In the upper part of the upper housing 23,
Encoder 2 for detecting rotation information of piston 11
4 are provided.

Between the movable sleeve 2 and the housing 4,
Displacement sensor 25 for detecting the end surface position of movable sleeve 2
Is arranged.

The rotary sleeve 3 that houses a part of the movable sleeve 2 on the discharge side is rotatably supported by a ball bearing 27 provided between the movable sleeve 2 and a bearing housing 26. 2
Reference numeral 8 denotes a connecting portion between the movable sleeve 2 and the piston 11.

The rotational torque given to the piston 11 from the motor 18 is given to the movable sleeve 2 by the connecting portion 28. That is, the piston 11 has a connecting portion shape (square cross section) that transmits only rotation to the movable sleeve 2 and free of relative linear motion (not shown).

Since the piston 11 is made of a non-magnetic material, it does not affect the closed loop magnetic circuit for controlling the expansion and contraction of the giant magnetostrictive rod 5. Further, since the driving torque given from the motor 18 is transmitted only to the piston 11, no torsional stress is generated in the giant magnetostrictive rod 5. Further, the gap between the inner peripheral surface of the giant magnetostrictive rod 5 and the outer peripheral surface of the piston 11 is sufficiently small, and the giant magnetostrictive rod 5 is configured to always maintain the center of rotation even during rotation. These measures can ensure the reliability and durability of the giant magnetostrictive rod, which is a brittle material that is weak against tensile stress.

With the above configuration, the drive unit of the fluid supply device of this embodiment can simultaneously and independently control the rotational movement and the linear movement of the minute displacement of the movable sleeve 2 and the piston 11.

In the embodiment, the direct-acting actuator 1
Since the giant magnetostrictive element is used, power for linearly moving the giant magnetostrictive rod 5 (and the movable sleeve 2 and the piston 11) can be applied from outside without contact. That is, in the actuator having this configuration, the two output shafts can be relatively moved in the axial direction with a high response while utilizing the characteristics of the electromagnetic strain element having a frequency characteristic of several megahertz while the motor is rotated. In the embodiment, the motor is disposed above the actuator 1 (the super-electrostrictive element), but the configuration may be reversed (not shown).

Hereinafter, the pump section 29 will be described. The pump section 29 is composed of members 30 to 42.

In this embodiment, by using the axial positioning function of the movable sleeve 2 and the piston 11, the gap between the discharge-side thrust end face of the movable sleeve 2 and the piston 11 is maintained while keeping both shafts in a constant rotation state. The size can be arbitrarily controlled. By using this function, it is possible to shut off and open the powder fluid in a state where any flow passage section from the suction port to the discharge nozzle is in a mechanically non-contact state.

When both shafts are stationary in the axial direction and the gap between the end surface of the movable sleeve 2 and the opposing surface is sufficiently small, the fluid is kept shut off by the dynamic pressure seal. When the gap is large, the effect of the dynamic pressure seal is reduced, and the fluid remains open (outflow).

In a transient state in which both shafts move in the axial direction, the movement of the fluid plays an important role in the movement of the piston 11 in addition to the movement of the movable sleeve 2.

First, the principle of sealing in a steady state in which both shafts are stationary in the axial direction will be described with reference to FIG.

(1) Principle of Sealing in Steady State In FIGS. 1 and 2, reference numeral 30 denotes a small-diameter portion of the movable sleeve 2, which is detachably attached to the movable sleeve 2 by bolts. Numeral 31 denotes a radial groove formed on the outer surface of the movable sleeve 2 for pressure-feeding a fluid to a discharge side, and numeral 32 denotes a cylinder that houses the movable sleeve 2 and is mounted on the bearing housing 24.

The movable sleeve small diameter portion 30 and the cylinder 3
Between them, a pump chamber 33 (fluid transport chamber) for obtaining a pumping action by the relative rotation of both members is formed. The cylinder 32 is provided with a suction hole 34 (inflow port) communicating with the pump chamber 33. Reference numeral 35 denotes a discharge nozzle (outflow port) attached to the lower end of the cylinder 32;
6 is a discharge flow passage formed in the discharge nozzle, 37 is a fluid seal, and 38 is a piston small diameter portion 38 housed in the small diameter portion 30 of the movable sleeve 2.

The discharge-side end surface 39 of the movable sleeve small-diameter portion 30
The thrust groove 4 for sealing is formed on the relative movement surface of the
1 is formed. Opposite surface 4 of this thrust end surface 39
The opening 42 of the discharge nozzle 35 at the center of the 0 is formed. 43 is a liquid pool part.

The radial groove 31 described with reference to FIG. 2 is a known one known as a spiral groove dynamic pressure bearing, and is also used as a screw groove pump.

The sealing thrust groove 41 is also known as a thrust dynamic pressure bearing. Now,
The steady state seal pressure at which a thrust bearing can be generated is given by the following equation. In the equation (1), ω is a rotational angular velocity, R ₀
Is the outer diameter of the thrust bearing, Ri is the inner diameter of the thrust bearing, δ is the gap between the opposing surfaces, f is a function determined by the groove depth, groove angle, groove width and ridge width, and the like.

[0056]

(Equation 3)

[0057] curve in the graph of FIG. 3 (a) under conditions of Table 1 shows the static characteristics of the sealing pressure P _S of gap δ when using the spiral groove type thrust groove.

The dynamic characteristics of the pressure when the gap δ fluctuates and when the piston also fluctuates in the opposite phase with respect to the movable sleeve will be described later in detail.

The curve (b) in the graph of FIG. 3 is an example showing the relationship between the pumping pressure of the radial groove and the gap δ at the tip of the shaft when there is no axial flow. The pumping pressure of this radial groove is the same as the thrust groove described above.
A wide range can be selected by selecting the radial gap, groove depth and groove angle. However, qualitatively, the pumping pressure Pr of the radial groove does not depend on the size of the gap at the tip of the shaft (that is, the size of the gap δ).

When the gap δ of the sealing thrust groove is sufficiently large, for example, when the gap δ is 15 μm, the generated pressure is P _S = 0.06 kg / mm ² (gauge pressure) [0.69 Mp
a (absolute pressure)].

While the shaft is being rotated, the end face of the main shaft 2 is brought closer to the opposing surface on the fixed side. Gap [delta] <becomes 10.0 [mu] m, the sealing pressure P _S is larger than the pumping pressure Pr of the radial groove, flowing out to the discharge port side of the fluid is blocked.

FIG. 2 shows a state in which the outflow of the fluid is cut off. The fluid near the opening 42 of the discharge nozzle is subjected to a centrifugal pumping action [arrows in FIG. The vicinity of the opening 42 is at a negative pressure (atmospheric pressure or lower). Due to this effect, after shutting off, the discharge nozzle 35
Part of the fluid remaining inside is sucked into the pump again. As a result, a fluid soul is not formed due to surface tension at the tip of the discharge nozzle, and stringing and dropping are eliminated.

As described above, in the embodiment of the present invention, the movable state of the fluid (ON, OFF) in the steady state can be freely selected by moving the movable sleeve 2 in the axial direction by about 10 to several tens of μm. To summarize the point, the sealing pressure by the thrust groove increases sharply as the gap δ decreases, whereas the pumping pressure of the radial groove is extremely insensitive to the change in the gap δ. .

The radial groove and the thrust groove may be formed on either the rotating side or the fixed side.

When a powder fluid such as an adhesive containing fine particles is applied, the minimum value δmin of the gap δ may be set to be larger than the fine particle diameter φd.

[0066]

(Equation 4)

In order to obtain a larger gap for the same generated pressure, increase the number of revolutions or provide a flange of a thrust seal on the movable sleeve 2 to increase the outer diameter of the collar and the groove depth, An appropriate value may be selected for the groove angle or the like (not shown).

[0068]

[Table 1]

Since the input current applied to the giant magnetostrictive element is proportional to the displacement, the open sleeve control without the displacement sensor can control the axial positioning of the movable sleeve 2. However, if feedback control is performed by providing the position detecting means as in the present embodiment, the hysteresis characteristic of the giant magnetostrictive element can be improved, and positioning with higher accuracy can be performed.

Further, in this embodiment, the piston 11 is made to penetrate the hollow giant magnetostrictive rod 5 and the bias permanent magnets 6 and 7, and both the front side and the rear side of the giant magnetostrictive rod 5 are movable ends. Are driven.

Although omitted in the embodiment, the piston 11
In order to detect the displacement of the piston shaft, a displacement sensor may be arranged at the rear end (the encoder 24 side) of the piston shaft.

(2) Sealing Principle in Transient State The sealing principle in the steady state in which the movable sleeve 2 and the piston 11 are stationary in the axial direction is as described above.

Hereinafter, the fluid behavior of the pump unit in a transient state in which the fluid is shut off and opened using the present dispenser will be described.

As already proposed by the present inventor in Japanese Patent Application No. 2000-188899 “Fluid supply device and fluid supply method”,
When the flow interruption time of the dispenser of the present invention is shortened as much as possible in order to improve the production tact, the "double piston type" having two output shafts (movable sleeve 2 and piston 11) is required to secure the flow accuracy. It is extremely effective. In addition, research after the above proposal has revealed that a problem occurs even when the discharge flow rate is released at a high speed in the single piston structure, and the double piston is an effective means for solving the problem.

Strict theoretical analysis was performed on the double piston type in comparison with the single piston type. As a result, further new knowledge was obtained. This will be described in detail.

When a viscous fluid is interposed in a narrow gap between the opposing planes and the interval between the gaps changes with time, the fluid pressure becomes squeeze action (Squeeze acti
It is obtained by solving the following Reynolds equation with the term on).

[0077]

(Equation 5)

In the equation (1), P is a pressure, μ is a viscosity coefficient of a fluid, h is a gap between opposed surfaces, r is a radial position, t is
Is time, U is the relative velocity in the x direction, and V is the relative velocity in the y direction. The right side is a term that provides a squeeze action effect that occurs when the gap changes.

Further, in the case where the liquid reservoir 43 is provided at the inlet of the discharge nozzle, the pressure of the liquid reservoir, that is, the pressure Pn on the upstream side of the nozzle is determined in consideration of the compressibility of the fluid.

[0080]

(Equation 6)

In the equation (4), Qs is the amount of inflow into the liquid reservoir, and Qn is the amount of outflow discharged from the liquid reservoir through the discharge nozzle 35 to the atmosphere. K is the bulk modulus of the fluid,
V is the volume of the liquid reservoir 43.

The upstream pressure Pn of the nozzle required for obtaining the discharge flow rate is obtained by simultaneously solving the above equations (3) and (4).

Hereinafter, the viscosity of the fluid: μ = 10,000 cps, the bulk modulus: K = 500 kg / cm ² (4.9 × 10 ⁷ N / m ² ), the boundary (the outer peripheral portion of the outer peripheral portion 45 of the thrust groove), the pressure: Ps0 = 20kg / cm ² (2.
06Mpa) (constant), an analysis was performed to determine the discharge pressure when the pump unit 29 was configured under the conditions shown in Table 2 below.

[0084]

[Table 2]

The analysis results obtained under the above conditions are summarized below.

(1) Displacement Curve of Each Output Shaft FIG. 4 shows a displacement curve of the movable sleeve 2 and the piston 11. The movable sleeve rises at t = 0.02 seconds (discharge flow path is opened)
At the same time, the piston starts to descend. Both axes stop at t = 0.03 seconds and maintain a constant position from t = 0.03 to 0.07 seconds. t = 0.07
In seconds, the piston starts to rise simultaneously with the lowering of the movable sleeve (cut off the discharge flow path), and both axes stop at t = 0.08 seconds.

(2) Comparison of pressure characteristics between double piston type and single piston type In FIG. 5, curve (a) shows a single piston type (when the piston is stationary and only the movable sleeve is driven in the axial direction). 3) shows the upstream pressure Pn of the discharge nozzle. In this case, immediately after the movable sleeve starts to rise, the pressure Pn on the upstream side of the discharge nozzle drops rapidly. The reason why the pressure drops suddenly is that there is a fluid resistance in the centripetal direction between the outer peripheral portion and the central portion of the gap portion of the thrust end surface caused by the sudden rise of the movable sleeve. Due to this fluid resistance, fluid is not easily replenished from the outer peripheral portion, and the pressure drops. In theory, Reynol
This is due to an effect that can be called an inverse squeezing effect where dh / dt> 0 in the ds equation (3 equations). Therefore, it takes t ≒ 0.035 seconds for the pressure Pn to reach the steady state pressure when the discharge is released.

The reason for the large negative pressure is that Re
This is because the ynolds equation (Equation 3) does not consider the compressibility of the fluid. Actually, the fluid pressure does not become less than zero absolute pressure (Pn <0.0MPa) due to the generation of bubbles and the like.

However, since the fluid does not flow out of the discharge nozzle during the generation of the negative pressure, the condition at which the flow starts is that Pn becomes higher than the atmospheric pressure by 0.02 seconds after the discharge start command is issued. . As a result of the experiment, the generation of this negative pressure caused air to flow in from the outlet of the discharge nozzle, and a part of the coating fluid normally filled in the flow passage of the discharge nozzle was replaced with air, so that the start of the flow of the coating fluid was further delayed. I understood.

Thereafter, for 0.035 <t <0.07 seconds, the state of continuous application is maintained.

When the movable sleeve starts lowering at T = 0.07 seconds, the pressure Pn on the upstream side of the discharge nozzle rapidly increases. The reason is due to the squeezing effect that occurs when dh / dt <0. Due to the steep pressure increase at this time (FIG. 1A), an excessive fluid is discharged immediately before the discharge is shut off.

From the above analysis results, it was found that in one cycle of opening and closing of the discharge, in the case of the single piston type, the pressure on the upstream side of the discharge nozzle is accompanied by a sharp drop and a sharp rise. These cause a delay in the discharge start time and a decrease in flow rate accuracy.

The curve (b) shows the pressure characteristics on the upstream side of the discharge nozzle in the case of the double piston type. Stroke of movable sleeve: Xst = 20 μm, stroke of piston: Xpt =
30 μm. Unlike the single piston type, there is almost no sudden drop and sudden rise in pressure. The pressure transients reach a steady state with little time lag with respect to the displacement curves of both axes (FIG. 4), but rather at a speed greater than the displacement curves.

This is because the squeeze effect and the reverse squeeze effect are simultaneously operated by driving the piston in the opposite phase with respect to the movable sleeve, thereby canceling the pressure rise and pressure drop caused by each. That is,
The piston has the effect of compensating for the failure of the dynamic characteristics when the dynamic pressure seal is used dynamically.

(3) Comparison when the piston stroke is changed in the double piston type In the above (2), by adopting the double piston type, the transient characteristics of the pressure (that is, the discharge flow rate) can be greatly improved. all right. Here, in order to evaluate the effect of the double piston type, the following dynamic characteristic compensation parameter: φ (dimensionless) is defined.

[0096]

Equation 7

In the above equation, a _s and Xst are the area and stroke of the movable sleeve, and a _p and Xpt are the area and stroke of the piston. That is, φ indicates the magnitude of the compensation amount for driving the piston in the opposite phase with respect to the movable sleeve to compensate for the pressure transient characteristic.

When φ = 0, it is the case of only a single piston having no compensating action, and when φ = 1, the size of the space in which each of the two shafts is reduced or expanded is equal, and the rate of change of the total volume is zero. Corresponds to the case of φ = ∞ is an extreme case where the movable sleeve is stationary and only the piston is movable.

FIG. 6 shows the pressure characteristics on the upstream side of the discharge nozzle when this parameter: φ is variously set.

When φ becomes 0.375 or less, the pressure decreases when the movable sleeve rises (discharge ON) and decreases when the movable sleeve lowers (discharge OF).
The pressure increase in F) begins. Conversely, when φ becomes 0.375 or more, a phenomenon opposite to that of the single piston type occurs. That is, when the movable sleeve rises (discharge ON), the pressure increases,
When dropping (discharge OFF), the pressure drops rapidly.

For example, when φ = 0.375 is seen, it can be seen from FIG. 7 that a surprisingly double piston effect can be obtained.

At the time of discharge release, Δt ₁ = 0.004 before the operation of raising the movable sleeve ends (t = 0.03 seconds).
With the rise time of second, the upstream pressure Pn of the discharge nozzle has reached the steady pressure.

At the time of discharge interruption, Δt ₂ = 0.003 seconds, which is about 3 of the time required for the movable sleeve to descend (0.1 seconds),
The pressure Pn drops under the condition that the flow rate is cut off, that is, below the atmospheric pressure.

The above results indicate that the dispenser can open and shut off the application fluid at a speed faster than the limit of the response time of the flow control valve determined by mechanical elements.

This effect is due to the fact that the action of the double piston for correcting the transient characteristic of the discharge pressure utilizes the squeeze action effect.

That is, the term (dh / dt) that gives the squeeze action effect in equation 3 is the time derivative term of the displacement h, and this is an improvement in response as if a mechanical “feed forward control” was performed. It has an effect.

Now, based on the results obtained from the above theoretical analysis, experiments were conducted by changing the compensation parameter of transient characteristics: φ.

In the case of 0 <φ <0.15 At the start of ejection, the effect of speeding up to some extent at the start up to the start of drawing is seen. However, it is not sufficient when increasing the speed.

In the case of 0.15 ≦ φ <0.25 A sufficient effect was obtained by a normal use method. It can sufficiently cope with processes requiring high-speed coating start and shut-off.

When φ ≧ 0.25 A remarkable effect was obtained at both the start and end of the ejection. At the end of application, dripping and stringing are remarkable. This is because a negative pressure is generated in the pressure upstream of the discharge nozzle immediately after the termination, and this negative pressure has an effect of instantly sucking the fluid soul at the tip of the discharge nozzle into the nozzle.

Regarding the upper limit of φ, the thrust seal 4
1 and the stroke that the giant magnetostrictive element can take practically. For example, if the outer diameter of the end face on the discharge side of the movable sleeve is Ds = 6 mm, the outer diameter of the piston is at most Dp
= 4.5mm is the limit. In addition, the stroke Xs of the movable sleeve
The limit of the piston stroke Xpt is about twice as large as Xst with respect to t. Therefore, φ ≦ (4.5 / 6) ² × (2/1) ≒ 1.2.

In this embodiment, the giant magnetostrictive element is used for the axial driving means.
The stroke of the gap δ for forming the “non-contact seal” may be on the order of at most several tens of microns, and the limit of the stroke of the magnetostrictive element such as the giant magnetostrictive element and the piezo element did not matter.

When a high-viscosity fluid is discharged, a large discharge pressure is expected due to the pumping action of the radial groove and the squeeze pressure generated on the shaft end face.
In this case, since the first actuator 1 is required to have a large thrust against high fluid pressure, it is preferable to use an electromagnetic strain type actuator capable of easily producing a force of several hundred to several thousand N. Since the magnetostrictive element has a frequency response of several MHz or more, each axis can be linearly moved with high response. Therefore, the discharge amount of the high-viscosity fluid can be controlled with high response and high accuracy.

Further, when a giant magnetostrictive element is used for the axial driving means, the conductive brush can be omitted as compared with the case where a piezoelectric element is used, so that the load on the motor (rotating means) can be reduced and the overall structure is extremely simple. Therefore, the moment of inertia of the moving part can be minimized, and the diameter of the dispenser can be reduced.

[0115] Incidentally, super-magnetostrictive element, the case of an electromagnetic strain element such as a piezoelectric element, typically, the order response of the 10- ⁴ sec is obtained.

An actuator such as an electromagnetic solenoid can also be applied to the present invention. Although the response is reduced by an order of magnitude as compared with an electromagnetic strain element, the restriction on the stroke (that is, the allowable stop time) is greatly eased.

In the embodiment, two output shafts are taken out by using one giant magnetostrictive actuator. However, two independent actuators may be arranged and controlled individually. Or only the mechanical part of the actuator
Using two, the control driver may be common. Further, the movable sleeve constituting the dynamic pressure seal may use a giant magnetostrictive actuator capable of transmitting power in a non-contact manner, and the actuator for driving the piston may use a piezoelectric element and be arranged on the fixed side.

Although the piston for compensating for the trouble caused by the movement of the movable sleeve is disposed at the center of the movable sleeve, the configuration may be reversed. That is, a dynamic pressure seal may be formed on the thrust end face on the piston side to control ON / OFF of the discharge amount by movement of the piston, and offset a sudden increase / decrease in pressure caused by the piston by the movable sleeve. In this case, the cross-sectional area and displacement of the piston are evaluated as as and Xst, and the cross-sectional area and displacement of the sleeve are evaluated as ap and Xpt.

The groove of the dynamic pressure seal may be formed on any end surface of the movable sleeve and the piston and the relative movement surface thereof.

Further, instead of the structure in which the piston is housed inside the movable sleeve, for example, a structure in which a movable member corresponding to the piston is arranged on the opposing surface of the movable sleeve may be used.
In short, the one-axis motion mechanism constituting the dynamic pressure seal may be provided with another axis motion mechanism which is driven in the opposite phase and cancels the volume change of the thrust-facing surface.

In the embodiment, the thread groove pump is used as the pressure source for supplying the application fluid to the fluid control section composed of the movable sleeve and the piston. However, the present invention can be applied to pumps other than the thread groove type. it can. For example, twin
A pump of a type such as a screw type, a trochoid type, a mono type, a gear type, and a piston type can be applied. Alternatively, an air type in which high-pressure air is applied to the application fluid may be used. A dynamic pressure seal that generates a pressure equal to or higher than the maximum pressure of these fluid pressurizing sources may be formed on the movable sleeve or the piston and their opposing surfaces.

The thrust dynamic pressure seal applied to the embodiment of the present invention is effective for completely shutting off fluid in a steady state. In a process in which a slight leak of the fluid from the discharge nozzle is allowed in the standby state of the application, the dynamic pressure seal may be omitted. In this case, there is no need to rotate the member corresponding to the movable sleeve, and no motor is required. The application may be performed using the fact that the fluid is pressurized.

[0123]

According to the fluid rotating device using the present invention, the following effects can be obtained. 1. High-speed discharge interruption and start of fluid can be performed. 2. No troubles such as clogging of the flow channel and change in fluid characteristics due to powder compression damage occur. 3. Further, the pump of the present invention can have the following features.

A high-viscosity fluid can be applied at a high speed.

An extremely small amount can be discharged with high precision.

If the present invention is applied to, for example, a surface mount dispenser, a phosphor coating for a PDP or CRT display, or a sealant coating for a liquid crystal panel, the advantages can be fully exhibited and the effect is enormous.

Further, it is believed that the ripple effect of the present invention is not limited to micro flow dispensers. Dynamic pressure seals using the principle of a fluid dynamic bearing are used in various fields, but most of them consider only a static state. In applications where the dynamic pressure seal is dynamically switched and used, as in the embodiment of the present invention, the basic principle and mechanism of the seal presented by the present invention are often used to improve the characteristics of the dynamic pressure seal in the transient state. Seems useful.

[Brief description of the drawings]

FIG. 1 is a front sectional view showing a dispenser according to a first embodiment of the present invention.

FIG. 2 is an enlarged sectional view of a discharge section of the embodiment.

FIG. 3 is a graph illustrating the principle of the present invention.

FIG. 4 is a graph showing displacement curves of a movable sleeve and a piston.

FIG. 5 is a graph showing analysis results of pressure characteristics of a single piston type and a double piston type.

FIG. 6 is a graph showing an analysis result of a pressure characteristic when a piston stroke is variously changed in a double piston type.

FIG. 7 is a graph showing pressure characteristics when φ = 0.375 in FIG. 6;

FIG. 8 is a diagram showing a conventional air pulse method.

[Explanation of symbols]

 2 shaft 7 pump chamber 8 suction port 9 discharge port 11 means for giving rotation 16 electromagnetic distortion element 17 suction hole

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Shuji Ono 1006 Kadoma Kadoma, Osaka Prefecture Matsushita Electric Industrial Co., Ltd. 72) Inventor Yamauchi Dai 1006 Kazuma Kadoma, Osaka Prefecture Matsushita Electric Industrial Co., Ltd. (72) Inventor Yoichi Nakamura 1006 Okadoma Makoto Kadoma City, Osaka Prefecture Matsushita Electric Industrial Co., Ltd. BB15 BB30 CC11 CC16 CC26 CC30 DA03 DA04 DA06 DA11 DA15 DA19 DB02 4F041 AA02 AA05 AB02 AB05 BA02 BA10 BA34 4F042 AA06 BA03 BA12 CB02 CB07 CB11 CB19

Claims (9)

[Claims] 1. A sleeve for accommodating a piston, axial driving means for driving the piston and the sleeve in substantially opposite phases, a housing for accommodating the sleeve, and a fluid formed by the sleeve and the housing. In order to increase or decrease fluid resistance between the fluid transport chamber and the discharge port, the transport chamber, a fluid inlet and an outlet formed in the housing communicating the fluid transport chamber with the outside, The directional drive means is configured to change the gap of the flow path formed between the sleeve and the housing, and the cross-sectional area and displacement of the piston are changed by a
Assuming that p, Xpt and the cross-sectional area and displacement of the sleeve are as and Xst, the following equation (1) parameter: φ (dimensionless) is defined as follows: A fluid ejection device, characterized in that φ ≧ 0.15. 2. The fluid ejection device according to claim 1, wherein 0.15 ≦ φ ≦ 1.2. 3. The fluid ejection device according to claim 1, further comprising means for relatively rotating the sleeve and the housing. 4. The fluid ejection device according to claim 3, wherein the means for increasing or decreasing the fluid resistance is a dynamic pressure seal formed between the sleeve and the housing. 5. The fluid discharge device according to claim 3, wherein a pump means for pressurizing the fluid in the fluid transport chamber by a relative rotation of the sleeve and the housing is provided. 6. The fluid discharge device according to claim 1, wherein the axial driving means is an electromagnetic strain element. 7. Using the output shafts at both ends of the electromagnetic distortion element,
7. The fluid ejection device according to claim 6, wherein the piston and the sleeve are driven. 8. An electrostrictive element having one end as a front side and the other end as a rear side, a piston penetrating through the electrostrictive element and pressed against the rear side of the electrostrictive element, A housing that houses the strain element and the piston, a rear sleeve that rotatably supports the rear side of the piston, and a sleeve that houses the front side of the piston and is pressed against the front side of the electromagnetic strain element. A front-side cylinder that rotatably supports the sleeve, a means for axially moving the piston and the sleeve, and a means for imparting rotation to one of the piston and the sleeve; In a fluid discharge device comprising an inlet and an outlet for a pressurized fluid that communicates with a fluid transport chamber formed by a side cylinder and the outside, In order to increase or decrease the fluid resistance between the fluid transport chamber and the outlet, the gap of the flow path formed between the sleeve and the housing is changed by driving the electrostrictive element. The fluid ejection device according to claim 1, wherein: 9. The fluid discharge device according to claim 8, wherein an opening of a flow passage communicating with an outlet is formed on a front end surface of the piston and a movement surface of the piston opposite to the front end surface.