JPH0689959B2 - Free-piston Stirling refrigerator - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a free piston Stirling refrigerator, and in particular, cools electronic equipment and the like at a cryogenic temperature of liquefied nitrogen temperature (−196 ° C.) or lower. The present invention relates to a free piston type Stirling refrigerator having a regenerator that is suitable for use.

[Conventional technology]

For the conventional free piston type Stirling refrigerator, for example, Proceedings of International Conference on Advanced Inflated Detectors and Systems (19
1981) (Proceedings of International Conference on
Small Spiral Stirling Coolers by A, K, de Jonge, as described in Advanced Infrared Detectors &Systems'81) Small Spli
t Stirling Coolers for IRDetectors), or similarly, Proceedings of International Conference on Advanced Inflated Detectors and Systems (1982).
(Proceedings of International Conference on Advan
The Oxford University mini by G. Davey, described in ced Infrared Detectors &Systems'82).
ature cryogenic refrigerator) is known.

This refrigerator has a piston for compressing a working fluid,
A linear motor for reciprocating the piston, a displacer for expanding the working fluid, and a reciprocating motion of the piston and the displacer to form a compression space area and an expansion space area of the working fluid. Of the cylinder, a regenerator for communicating the compression space region and the expansion space region to recover cold heat energy, and a working fluid flow path connecting these elements to form a reverse Stirling cycle. As in the cited document, the regenerator may be placed inside the displacer or may be separately placed in the middle of the flow path connecting the compression space and the expansion space. Usually, a direct acting electric motor (linear motor) is attached to the piston side, and the working gas is compressed by reciprocating the piston. The compressed gas flows into the expansion space area above the displacer through the regenerator, but due to the pressure loss when passing through the regenerator, there is a pressure between the expansion space area above the displacer and the compression space area below. There is a difference. This pressure difference serves as a driving force for the displacement of the displacer, but the size and direction of the pressure difference change depending on the pressure change in the compression space region. Therefore, the mass of the displacer and the spring constant of the spring supporting the displacer cause the displacer to vibrate with a certain phase difference from the piston. That is, if the structure of the regenerator or the flow path is determined, the change in pressure difference is uniquely determined, and the vibration phase difference is determined accordingly, so the refrigerating capacity is determined. However, if the refrigeration load changes or the removal of compression heat generated in the compression space region is incomplete or unstable, the pressure difference change fluctuates, and the displacer does not vibrate at the initially designed vibration phase difference position. The refrigerating capacity and thus the efficiency cannot be fixed at the optimum position, but this point has not been considered in the prior art.

[Problems to be solved by the invention]

As described above, in the conventional technology, when the load fluctuation or the cooling of the compression heat is incomplete, the phase difference between the piston and the displacer or the amplitude of the displacer changes, and the optimum efficiency point and the optimum performance point are deviated. It will be driven in a different position.

In order to avoid this problem, the vibration of the displacer is
There is also a method of fixing the phase difference and amplitude by controlling the input power to the motor by using a direct drive motor separate from the piston instead of the pressure difference, but the mechanism becomes complicated while the control circuit It is also difficult to ensure stability.

As another problem, if the structure of the regenerator is taken up, even if the structure is thermally sufficient, if the required pressure difference on both sides of the regenerator cannot be secured, the displacer amplitude can be secured sufficiently. Instead, it will lead to a lack of freezing capacity.

The present invention has been made in order to solve the above-mentioned problems of the prior art, and even if the working fluid cooling state in the load or the compression space region changes under the thermally sufficient regenerator structure. , The purpose is to provide a free piston type Stirling refrigerator that can maintain stable refrigeration capacity and refrigeration efficiency by keeping the vibration phase difference between the piston and the displacer and the vibration amplitude of the displacer at optimum values. There is.

[Means for solving problems]

In order to achieve the above object, the structure of a free piston type Stirling refrigerator according to the present invention has a piston for compressing a working fluid, a linear motor for reciprocating the piston, and an expansion of the working fluid. And a cylinder for reciprocating the piston and the displacer to form a compression space region and an expansion space region of the working fluid, and a cold energy for communicating with the compression space region and the expansion space region. In a free piston type Stirling refrigerator consisting of a regenerator for recovering and a working fluid passage connecting these elements to form an inverse Stirling cycle, a working fluid passage portion at one end or both ends of the regenerator Is provided with a fluid throttle means.

In addition, in order to solve the above-mentioned problems, it is sufficient to prevent the driving force of the displacer from changing even if the load condition changes.

This can be achieved by controlling the pressure difference between both ends of the displacer, that is, the pressure difference between both ends of the regenerator to be constant according to the operating conditions.

If the pressure difference before and after the displacer is insufficient,
Since the amplitude of the displacer is reduced and the degree of expansion is reduced, sufficient cooling capacity cannot be expected.Therefore, if the working fluid flow path on both sides of the regenerator is forced to have a pressure difference above a certain value, The amplitude of the displacer can be secured.

Therefore, if a throttle mechanism is provided in front of and behind the regenerator and a pressure difference above and below the displacer is set to a certain level or more, the amplitude of the displacer can be secured. In addition, if this throttle mechanism that can electrically change the opening is adopted, the throttle part can be controlled by controlling the phase difference between the piston and the displacer, the amplitude of the displacer, and the pressure difference between the compression chamber and the expansion chamber. The optimum performance can be maintained by controlling.

[Action]

The operation of the present invention will be verified analytically. The equation of motion for the displacer is given by equation (1).

Md + (Pe-Pc) Ad + kdx = 0 (1) where Md: mass of displacer x: displacement of displacer Pe-Pc: pressure difference before and after displacer Ad: cross section of displacer kd: support spring Spring constant The pressure difference Pe-Pc is the speed between the displacer and piston,
And it is determined by the flow path loss in the regenerator, etc. Pe-Pc = Cdd + Cpp (2) where Cdd: Coefficient indicating the pressure loss in the regenerator Cpp: Coefficient indicating the pressure loss in the regenerator y: Piston It becomes a displacement. Note that the loss due to damping is not taken into consideration in equations (1) and (2). Substituting equation (2) into equation (1) and assuming that the piston and displacer sine-oscillate with a phase difference α, The sine vibration between the piston and the displacer is assumed to be the equations (4) and (5).

Where ω is the angular frequency, Is the vibration amplitude and t is the time.

If the equation (3) holds for an arbitrary time t, the equations (6) and (7) are obtained.

α = tan ⁻¹ {M _d (ω ² −ω _d ² ) / (C _dd ω Ad)} ......
(7) Here, ωd is expressed by equation (8).

ω _d ² = K _d / M _d (8) On the other hand, the refrigerating capacity Qe is given by the equation (9) as can be seen from the above-mentioned known example.

Here, Z: coefficient due to pressure change, etc.As shown above, the vibration phase difference α between the piston and the displacer and the amplitude of the displacer can be determined by the coefficients Cdd and Cpp determined by the pressure loss in the regenerator. Can be decided.

Further, since the vibration phase difference α and the amplitude determine the refrigerating capacity Qe as shown in the equation (9), the pressure difference between the working fluid before and after the regenerator becomes a large factor that affects the refrigerating capacity and thus the refrigerator efficiency. . Therefore, if the pressure difference is controlled according to the operating conditions, the refrigerator can be operated at the optimum operating point.

For that purpose, it is necessary to observe a physical quantity indicating the operating state of the refrigerator as a reference for determining whether the control of the pressure difference is appropriate.

One of them is a method of directly detecting the phase difference. That is, the pressure difference before and after the displacer,
If the relationship with the vibration phase difference between the piston and displacer is known in advance, performance can be maintained by adjusting the diaphragm mechanism to match the vibration phase difference (target vibration phase difference) in the optimum operating state. It will be possible.

Furthermore, if the relationship between the pressure difference and the change of the displacer, that is, the refrigerating capacity is grasped in advance, the pressure difference may be adjusted by detecting the change of the displacer.

As described above, if the pressure difference in the regenerator, that is, the pressure difference before and after the displacer is controlled, it is possible to operate the refrigerator at the optimum performance point.

〔Example〕

Embodiments of the present invention will be described below with reference to FIGS. 1 to 7.

First, FIG. 1 is a vertical cross-sectional view of a free piston type Stirling refrigerator according to an embodiment of the present invention.

In FIG. 1, reference numeral 1 is a piston for compressing a working fluid, for example, helium gas, and 2 is a linear motor for reciprocating the piston 1, for example, a voice coil motor. 3 is a displacer for expanding the working fluid, 4 is the piston 1 and the displacer 3
Is a cylinder for reciprocating to form a compression space region 5 and an expansion space region 6 of the working fluid. In this cylinder 4, a piston 4 reciprocates to form a compression space region 4a portion, a display. The sag 3 reciprocates to form an expansion space region 6b, and a compression space region 5 and an expansion space region 6c.

Reference numeral 7 denotes a regenerator for communicating with the compression space region 5 and the expansion space region 6 via a working fluid flow passage 8 to collect cold energy, and 9 denotes a fluid throttling means provided in the working fluid flow passage 8. This is the diaphragm mechanism.

Reference numeral 10 is a support spring that elastically supports the reciprocating motion of the displacer 3 in the cylinder 4c portion, 11 is a spring that elastically supports the reciprocating motion of the piston 1, and 12 is a cooler.

The operation of such a refrigerator will be described.

The linear motion electric motor 2 moves the piston 1 upward in the figure to compress the working fluid in the compression space region 5. The compressed working fluid flows into the expansion space region 6 from the cylinder 4c through the expansion mechanism 9 and the regenerator 7. However, due to the pressure loss in the regenerator 7 and the expansion mechanism 9, the pressure in the expansion space region 6 is reduced. Since the pressure becomes lower than the pressure in the compression space region 5, the displacer 3 moves upward.

At this time, the heat of compression applied to the compressed working fluid is removed by the cooler 12.

Next, as the piston 1 moves downward, the pressure in the compression space region 5 drops, so the working fluid in the expansion space region 6 passes through the throttle mechanism 9 and the regenerator 7 and moves into the compression space region 5. .

At this time, contrary to the above case, the pressure in the expansion space region 6 becomes higher than the pressure in the compression space region 5 due to the pressure loss, so that the displacer 3 moves downward. For this reason,
The isothermal expansion of the working fluid absorbs heat from the outside and causes a refrigerating action in the expansion space region 7.

Thus, the inverse Stirling cycle of isothermal compression, constant volume heat dissipation, isothermal expansion, and constant volume endotherm is established.

As described above, the movement of the displacer 3 is caused by the pressure loss, but the phase difference of the vibration with the piston 1 is also influenced by the mass of the displacer 3 and the spring constant of the support spring 10. receive. Now, the displacer 3 and the support spring 10
If the structure is determined, the vibration phase difference is determined by the regenerator 7 and the throttle mechanism 9.

Next, FIG. 2 is a partial sectional view of a free piston type Stirling refrigerator according to another embodiment of the present invention, in which the throttle mechanism 9 in FIG. 1 is replaced with a variable throttle mechanism 13. In the figure, the same reference numerals as those in FIG. 1 and parts not shown are the same as those in FIG. 1, and therefore their explanations are omitted.

FIG. 3 is a diagram showing the relationship between the pressure loss coefficient, which is the controlled object of the present invention, and each characteristic.

According to this embodiment, the variable throttle mechanism 13 is operated when the operating condition, for example, the cooling condition in the cooler 12 changes, or the heat load attached to the expansion space region 6, that is, the cooled object (not shown) or the like changes. The refrigeration cycle can be set to the required conditions by adjusting the throttle (opening).

This operation is shown in the graph of FIG.

FIG. 3 shows the inside of the regenerator 7 which is one of the pressure loss coefficients of the flow path from the compression space region 5 to the expansion space region 6 by adjusting the throttle of the variable throttle mechanism 13 shown in FIG. 7 shows changes in the displacement of the displacer, the refrigerating capacity Qe, and the vibration phase difference α when the pressure loss coefficient C _dd indicating the pressure loss is changed.

Since the characteristics are almost the same even if the pressure loss coefficient Cpp is changed, only Cdd is shown here. Cdd and Cpp are
It is the coefficient shown in equation (2) above. Figure 3 shows The characteristics are shown when the driving frequency and the working fluid filling pressure are constant, and when these characteristics change, the characteristics shown in FIG. 3 change. As is clear from the figure, the refrigerating capacity Qe becomes maximum at a certain pressure loss coefficient, and the phase difference α and the displacer displacement are uniquely determined. That is, if the refrigerator characteristics shown in FIG. 3 are understood with respect to the displacement of the operating conditions such as the piston displacement, the drive frequency, and the filling pressure described above, the pressure loss coefficient can be adjusted by adjusting the variable throttle mechanism 13. By setting, it becomes possible to operate the refrigerator in the optimum state.

4 to 6 show a control means for more accurately implementing the operation control described above.

4 to 6 are all control system diagrams of a free piston type Stirling refrigerator according to still another embodiment of the present invention, and a refrigerator main body not shown in FIGS. 4 to 6 is shown. Is equivalent to FIG. Further, since the parts having the same reference numerals as those in FIG. 1 are the same as those in the embodiment of FIG. 1, the description thereof will be omitted.

Note that each of these examples describes the case where the conditions such as the enclosed pressure and the drive frequency described above are constant, but even when these conditions fluctuate, the information shown in FIG. 3 is used as data. If so, there is no particular problem in control. Moreover, the above-mentioned conditions are usually operated under constant conditions.

The embodiment shown in FIG. 4 is an example in which the displacement x of the displacer 3 is detected and the variable aperture mechanism 13 is controlled so as to attain the target displacement.

In FIG. 4, reference numeral 14 is a displacement detecting means for detecting the displacement x of the displacer 3.
Various types of devices can be considered, such as those applying a differential transformer or a gearup sensor, but the individual detecting means is not limited here.

15 is a central processing unit such as a micro computer (hereinafter
CPU), 16 is a diaphragm mechanism drive circuit for controlling the diaphragm of the variable diaphragm mechanism 13 in response to a command signal from the CPU 15,
5. The calculation control means is constituted by the diaphragm mechanism drive circuit 16.

Displacement x of the displacer 3 detected by the displacement detecting means 14
However, if it is larger than the target displacement which is the vibration amplitude set in advance, the variable throttle mechanism 13 is opened to reduce the pressure loss coefficient. That is, the characteristics are shifted to the left in the diagram of FIG. When the displacement x is smaller than the target displacement, the variable throttle mechanism 13 is narrowed down to increase the pressure loss coefficient and shift the characteristic to the right in FIG.

According to the present embodiment, since the control means capable of determining the pressure loss of the regenerator to a desired state is provided, it is possible to operate in an optimal state of the displacer displacement corresponding to the operating condition. Even if external conditions such as cooling of the compressed fluid fluctuate, the performance can be properly maintained.

Next, in the embodiment shown in FIG. 5, by adding displacement detecting means for the piston 1, the position phase difference from the displacer 3 is detected, and the vibration phase difference is used as a control target to operate in an optimum performance state. That is. In addition, in FIG.
The parts having the same reference numerals as those in the figure are the same as those in the embodiment of FIG.

In FIG. 5, 17 is a piston displacement detecting means for detecting the displacement of the piston 1, and 18 is a displacement detecting means 14 and a piston displacement detecting means 17 for the displacer, which are provided between the piston 1 and the displacer 3. 2 is a phase difference detection circuit for detecting the vibration phase difference of.

The phase difference detection circuit 18, the CPU 15, and the diaphragm mechanism drive circuit 16 constitute the arithmetic control means.

From the displacement of the displacer 3 detected by the displacement detecting means 14 and the displacement of the piston 1 detected by the piston displacement detecting means 17, the vibration phase difference α between the piston 1 and the displacer 3 is detected by the phase difference detecting circuit 8. To be done. If this vibration phase difference α is larger than the target vibration phase difference α _{0, the} variable throttle mechanism 13 is narrowed down by the action of the CPU 15 and the throttle mechanism drive circuit 16 to increase the pressure loss coefficient, and the characteristic is shifted to the right in FIG. . If the vibration phase difference α is smaller than the target vibration phase difference α ₀ , the variable throttle mechanism 13 is opened to reduce the pressure loss coefficient, and the characteristic is shifted to the left in FIG.

According to the present embodiment, the operation can be performed with the vibration phase difference between the piston 1 and the displacer 3 which is suitable for the operating condition.
The same effect as that of the embodiment shown in the figure is expected.

In the embodiment of FIG. 5, various methods can be considered for the phase difference detection circuit 18, but any method can be used as long as it satisfies the control accuracy.

Next, in the embodiment shown in FIG. 6, the variable throttle mechanism 13 is controlled by detecting the pressure difference between the compression space region 5 and the expansion space region 6 with the regenerator 7, variable throttle mechanism 13 and the like interposed therebetween. It is a thing. In FIG. 6, the same reference numerals as those in FIG. 4 are equivalent to those in the embodiment of FIG.

In FIG. 6, 19 is a pressure difference detecting means for detecting the pressure difference between the compression space area 5 and the expansion space area 6, that is, the pressure difference between the working fluids on both sides of the regenerator 7, and 20 is the cold storage from the detected pressure difference. It is a pressure loss coefficient calculation circuit for calculating the pressure loss coefficient of the device 7.

Pressure loss coefficient calculation circuit 20, CPU15, throttle mechanism drive circuit 16
In addition, an arithmetic control unit that outputs a control signal in accordance with the pressure difference detected by the differential pressure detection unit 19 and the relationship between the pressure difference and the pressure loss coefficient stored in advance is configured.

The pressure difference changes due to the vibration (reciprocating motion) of the piston 1 and the displacer 3, that is, the amount of change in the drive frequency. By this pressure difference displacement, the pressure difference (Pe-Pc) and the pressure loss coefficient Cdd, according to the equation (2) shown above,
If the relationship with Cpp is grasped in advance, the pressure loss coefficient can be determined by directly controlling the pressure loss by the variable throttle mechanism 13, and the operation in the optimum performance state becomes possible.

FIG. 7 is a diagram showing the relationship between pressure loss and pressure loss coefficient.

In FIG. 7, the horizontal axis represents the pressure loss and the vertical axis represents the regenerator flow rate Q, the broken line Q ₀ is the reference flow rate, and the solid line with an arrow is the pressure loss coefficient. It is clear from the figure that the pressure loss increases as the pressure loss coefficient increases for a constant flow rate.

In each of the above-described embodiments, the displacement of the displacer, the vibration phase difference between the piston and the displacer, the pressure difference between the working fluids on both sides of the regenerator, etc. are detected independently and the variable throttle mechanism is used to store the cold energy. Although the example of adjusting the flow rate through the device and controlling so as to obtain the optimum pressure loss state has been described, the present invention is not limited to the above-mentioned three independent control means, and two or more of these control means may be used. The variable diaphragm mechanism may be controlled by combining the above.

〔The invention's effect〕

As described above, according to the present invention, under the thermally sufficient regenerator structure, even if the working fluid cooling state in the load or the compression space region changes, the vibration between the piston and the displacer Keeping the phase difference and the vibration amplitude of the displacer at optimum values,
It is possible to provide a free piston Stirling refrigerator that can maintain stable refrigeration capacity and refrigeration efficiency.

[Brief description of drawings]

1 is a vertical sectional view of a free piston Stirling refrigerator according to an embodiment of the present invention, and FIG. 2 is a partial sectional view of a free piston Stirling refrigerator according to another embodiment of the present invention. FIG. 3 is a diagram showing the relationship between the pressure loss coefficient which is the control symmetry of the present invention and each characteristic, and FIGS. 4 to 6 are all free piston type Stirling according to still another embodiment of the present invention. Fig. 7 is a control system diagram of the refrigerator.
It is a diagram which shows the relationship between a pressure loss and a pressure loss coefficient. 1 ... Piston, 2 ... Linear motor, 3 ... Displacer, 4 ... Cylinder, 5 ... Compression space area, 6 ... Expansion space area, 7 ... Regenerator, 8 ... Working fluid flow Road, 9 ...
… Throttle mechanism, 13 …… Variable diaphragm mechanism, 14 …… Displacement detection means, 15 …… CPU, 16 …… Diffusion mechanism drive circuit, 17 …… Piston displacement detection means, 18 …… Phase difference detection circuit, 19 …… Differential pressure detection means, 20 ... Pressure loss coefficient calculation circuit.

Claims (5)

[Claims] 1. A piston for compressing a working fluid, a linear motor for reciprocating the piston, a displacer for expanding the working fluid, and a reciprocating piston and displacer. A cylinder for forming a compression space region and an expansion space region of the working fluid, a regenerator for communicating cold and heat energy by communicating with the compression space region and the expansion space region, and connecting and connecting these elements to each other. A free-piston type Stirling refrigerator comprising a working fluid flow path constituting a Stirling cycle, characterized in that a working fluid flow path portion at one end or both ends of the regenerator is provided with a fluid throttling means. refrigerator. 2. A variable throttle mechanism according to claim 1, wherein the fluid throttle means provided in the working fluid passage is a variable throttle mechanism, and a control means for controlling the throttle of the variable throttle mechanism is provided. A free piston Stirling refrigerator. 3. The control means of the variable diaphragm mechanism according to claim 2, wherein at least displacement detecting means of the displacer is provided, and the detected displacement is compared with a preset target displacement. A free-piston Stirling refrigerator equipped with arithmetic control means for outputting a control signal. 4. The variable aperture mechanism control means according to claim 2, wherein at least a piston displacement detection means and a displacer displacement detection means are provided.
The free-piston type Stirling refrigerating machine is provided with arithmetic control means for comparing the vibration phase difference between the piston and the displacer detected by these displacement detection means with a preset target vibration phase difference and outputting a control signal. Machine. 5. The variable throttle mechanism control means according to claim 2, wherein at least differential pressure detection means for detecting a pressure difference between the working fluids on both sides of the regenerator is provided.
A free piston type Stirling refrigerator having arithmetic control means for outputting a control signal in accordance with the detected differential pressure and the relationship between the pressure difference and the pressure loss coefficient stored in advance.