JP6609492B2 - Compressor - Google Patents

Description

  The present invention relates to a compressor.

  As a background art of this technical field, there is Patent Document 1. The engine disclosed in Patent Document 1 has an anti-resonance frequency generated by superimposing the vibration mode of the rotational vibration system from the driving force transmission mechanism to the sub flywheel and the vibration mode of the roll vibration system accompanying the rotation of the crankshaft. Roll vibration is reduced by making it substantially coincide with one of the frequencies obtained by multiplying the rotation frequency at a predetermined rotation speed of the shaft by (natural number / 2).

JP 11-325186 A

  Although the engine of Patent Document 1 is considered with respect to reduction of roll vibration, reduction of tension applied to a belt as a driving force transmission mechanism is not considered and is insufficient.

  In view of the above problems, the present invention considers the belt as a spring and sets the various parameters in consideration of the natural frequency of the two-degree-of-freedom rotational vibration system composed of two pulleys. The purpose is to prevent and ensure the reliability of the belt for a long time.

  In order to solve the above problems, for example, the configuration described in the claims is adopted.

  The present invention includes a number of means for solving the above-mentioned problems. For example, a compressor main body that compresses fluid, a motor that drives the compressor main body, and driving of the compressor main body are included. A compressor pulley that rotates in response to the drive of the motor, and a belt that transmits the power of the motor pulley to the compressor pulley, the compressor body having M cylinders, When the belt is a spring and the motor pulley is a weight, the angular frequency at the anti-resonance point at which the fluctuation of the rotation speed of the motor pulley stops is ω1 * with respect to the fluctuation of the rotation speed of the compressor pulley, and the compressor Provided is a compressor characterized by satisfying Mω <ω1 * when the rotational angular frequency of the main body is ω.

  According to the present invention, it is possible to prevent excessive tension generated in the belt and ensure the reliability of the belt for a long period of time.

It is a figure which shows the whole structure of the compressor in Example 1, 2 of this invention. It is a figure which shows the structure of the compressor main body 1 in Example 1 of this invention. 4 is a graph showing the relationship between the phase of the crankshaft 24, the pressure in the cylinder 22, and the axial load torque of the crankshaft 24. It is a graph which shows the time change of the rotational speed of the compressor pulley 2 and the motor pulley 4, and the tension concerning a belt. It is a graph which shows the time change of the rotational speed of the compressor pulley 2 and the motor pulley 4, and the tension concerning a belt. 6 is a graph showing the relationship between the fluctuation phase difference between the compressor pulley 2 and the motor pulley 4 and the tension fluctuation piece amplitude of the belt 5 with respect to the angular frequency of the compressor pulley 2. It is a figure which shows the structure of the compressor main body 1 in Example 2 of this invention.

  Hereinafter, Example 1 will be described with reference to the drawings.

  First, the basic configuration and concept of the compressor according to the first embodiment of the present invention will be described.

  FIG. 1 shows a schematic diagram of a basic compressor in Embodiment 1 of the present invention. 2 shows a schematic diagram of the internal structure as seen from the side of the compressor body in FIG.

  The compressor shown in FIG. 1 includes a tank 7 installed on a floor surface or the like, and a compressor main body 1 installed on the tank 7.

  As shown in FIG. 2, the compressor main body 1 compresses fluid. As shown in FIG. 2, the compressor body 1 closes the crank chamber 21, one cylinder 22 projecting vertically from the crank chamber 21, and the upper portion of the cylinder 22. And a crankshaft 24 that is rotatably supported at the center of the crank chamber 21. In the compressor main body 1, the crankshaft 24 in the crank chamber 21 rotates, whereby the piston 25 installed in the cylinder 22 reciprocates in the vertical direction. As a result, the fluid is sucked, compressed, and discharged. . In FIGS. 1 and 2, the compressor shape is a single cylinder having only one pair of piston / cylinder for simplification of explanation, but it has a plurality of pistons / cylinders in series or radially with respect to the crankshaft 24. It may be.

  The compressor body 1 is fixed to the upper surface of the tank 7 with the crankshaft 24 disposed in parallel with the rotation shaft of the motor 3. The compressor pulley 2 is attached to the crankshaft 24, and the rotation shaft of the motor 3 is attached to the rotation shaft of the motor 3. The motor pulley 4 is fixed. The compressor pulley 2 attached to the compressor main body 1 has blades, and generates cooling air toward the compressor main body 1 along with the rotation of the compressor pulley 2 to promote heat dissipation of the compressor main body 1.

  A belt 5 for power transmission is wound around the compressor pulley 2 and the motor pulley 4. Accordingly, the crankshaft 24 as the drive shaft of the compressor body 1 is rotationally driven through the motor pulley 4, the belt 5 and the compressor pulley 2 according to the rotation of the motor 3, and the compressor body 1 compresses the fluid.

  In the present embodiment, the crankshaft 24 cranked as the drive shaft will be described as an example. For example, a non-cranked drive shaft is connected to an eccentric position of the base end portion of the piston 25 and the base of the piston 25 is connected. The end may be eccentric, and the drive shaft does not necessarily have to be cranked.

  Here, the shaft load torque generated in the crankshaft 24 during steady operation will be described.

  Since the reciprocating compressor as shown in FIGS. 1 and 2 performs the operation of compressing the fluid discontinuously due to its mechanism, the axial load torque generated in the crankshaft 24 is reduced during one rotation of the crankshaft 24. It fluctuates greatly. Here, the axial load torque means that a component transmitted to the crankshaft 24 through the connecting rod out of the force applied to the piston 25 by the pressure of the compressed air in the cylinder 22 causes the crankshaft 24 to rotate. It means the power to try. FIG. 3 shows an example of changes in the rotation angle of the crankshaft 24 and the accompanying cylinder 22 internal pressure and shaft load torque. Here, the crank angle indicates an angle at which the crankshaft 24 is rotated from the start when the rotation angle (phase) at the start of the crankshaft 24 is 0 degree.

  Such a large fluctuation in the shaft load torque increases the effective value by changing the current value of the motor 3 to increase the load on the electric system, increase the reciprocating inertia force, and reduce the vibration of the compressor body. Or cause it. Therefore, a reciprocating compressor generally has a flywheel for smoothing the rotational speed fluctuation resulting from this torque fluctuation. Specifically, by providing an annular weight portion on the compressor pulley 2 or the motor pulley 4, the inertia moment of the pulley is increased. By doing so, since the rotational kinetic energy of the pulley in the rotating state increases, even if some variation occurs in the axial load torque, it is possible to suppress a change in rotational speed due to the influence. Usually, the compressor pulley 2 and the motor pulley 4 are designed as cast parts, and therefore, in many cases, annular pulleys are provided on the outer peripheral portions of these pulleys to also function as a flywheel.

  In general, in a small reciprocating air compressor as shown in FIG. 1, the rated rotational speed of the compressor body 1 is designed to be smaller than the rated rotational speed of the motor 3. Therefore, the compressor pulley 2 has a larger diameter than the motor pulley 4, and it is easy to increase the moment of inertia if a weight is provided on the outer peripheral portion. Therefore, the flywheel effect is often provided by the compressor pulley 2 side. However, if it is difficult to obtain rotational kinetic energy even if the inertia moment of the compressor pulley 2 is increased, such as when the rated rotational speed of the compressor body 1 is extremely low, a flywheel may be provided on the motor pulley 4. Done.

  Here, in the compressor in the present embodiment, the size and number of the belts 5 are appropriately selected depending on the magnitude of the transmission power. However, in a reciprocating compressor, since compression occurs intermittently, the torque applied to the crankshaft 24 varies greatly while the crankshaft 24 is rotating once. Accordingly, the transmission power that the belt 5 undertakes during this period, that is, the tension generated in the belt 5 also fluctuates. For this reason, even if the average transmission power is appropriate, the allowable range may be exceeded momentarily. The excessive tension at this time causes wear or breakage of the belt 5.

  Further, when the pulley diameter is extremely small, for example, when the designed rotational speed of the compressor main body 1 is extremely low with respect to the rotational speed of the motor 3 that drives the compressor body 1, the curvature in a state where the belt 5 is wound around the pulley. The radius becomes smaller. If the radius of curvature of the belt 5 is small, the cross section of the belt 5 is distorted and causes local contact with the belt groove on the pulley side, which also causes problems such as the aforementioned wear.

  In addition to this, as the rotational speed of the compressor body 1 decreases, the rotational kinetic energy of rotating parts such as pulleys decreases, so that the above torque fluctuations directly act on the belt 5 and wear due to excessive tension. To promote. Such a problem becomes particularly prominent when the compressor body 1 having a relatively large capacity is used at a low speed for the purpose of low noise.

  Next, fluctuations in tension applied to the belt 5 by driving the compressor body 1 will be described.

  Since the power transmission between the compressor pulley 2 and the motor pulley 4 is performed by the belt 5, the surplus that was not completely suppressed by the rotational kinetic energy of the compressor pulley 2 among the fluctuations in the axial load torque generated in the compressor body 1. The minute is transmitted to the motor pulley 4 via the belt 5. Therefore, even if the motor pulley 4 has a sufficient moment of inertia, and the entire rotation system including the compressor body 1 and the motor 3 can suppress fluctuations in the rotational speed, the tension of the belt 5 does not fluctuate. It is happening. Furthermore, as the moment of inertia of the motor pulley 4 increases, the response of the motor pulley 4 is delayed with respect to the rotational speed fluctuation of the compressor body 1 and does not follow. The tension of the belt 5 may fluctuate significantly depending on the difference in rotational fluctuation between the compressor pulley 2 and the motor pulley 4 at this time. This example is shown below.

  FIGS. 4a and 4b show the results of simulating the rotational speeds of the compressor pulley 2 and the motor pulley 4 and the time variation of the tension generated in the belt 5 of the compressor rotating under a certain condition. 4a and 4b differ only in the moment of inertia of the motor pulley.

  4a and 4b, the vertical axis indicates the rotational speeds of the compressor pulley 2 and the motor pulley 4 and the tension generated in the belt 5, and the horizontal axis indicates the crank angle. As in FIG. 3, the crank angle indicates an angle at which the crankshaft 24 is rotated from the start when the rotation angle (phase) at the start of the crankshaft 24 is 0 degree.

  In FIG. 4a, it can be seen that the rotational speeds of the compressor pulley 2 and the motor pulley 4 are substantially the same in crank angle at which the change reaches a valley and fluctuates in the same phase. Here, the belt 5 can be regarded as a spring that expands and contracts due to the difference in rotational fluctuation of each pulley. In the example shown in FIG. 4a, since the rotational fluctuation is in phase and the difference is small as described above, the tension fluctuation width of the belt 5 corresponding thereto is also kept low.

  On the other hand, in FIG. 4b, the rotational speeds of the compressor pulley 2 and the motor pulley 4 fluctuate in opposite phases in the mountain valley. Therefore, the tension of the belt 5 corresponding to this shows a fluctuation having a large peak every 360 degrees where the rotational speed of the motor pulley 4 is a trough and the rotational speed of the compressor pulley is a crest.

  Therefore, in this embodiment, a condition for increasing the tension fluctuation is obtained by paying attention to the phase difference of the rotational speed fluctuation of each pulley, and various parameters are set based on this condition.

  When a motion equation of a two-degree-of-freedom rotational vibration system including the compressor pulley 2, the motor pulley 4, and the belt 5 is established, the following equation is obtained.

  However, the moment of inertia of the pulley, θ is the rotation angle of the pulley, K is the spring constant of the belt 5, r is the radius of the pulley, T is the axial load torque on the compressor body 1, the subscripts M and C are the motor pulley 4, The compressor pulleys 2 and ω indicate the rotational angular frequency of the compressor body 1 (compressor pulley 2). Here, it is assumed that the axial load torque on the compressor side is sinusoidal.

  When the natural angular frequency of this system is obtained from the equations (1) and (2), the following equation is obtained.

  Further, the tension amplitude FB of the tension generated in the belt 5 is given by the following equation.

  From the above equation, it can be seen that as the rotational angular frequency of the compressor approaches the natural angular frequency ω2, the half amplitude of the tension of the belt 5 becomes very large and a resonance state is obtained. At this time, the rotational speed fluctuations of the compressor pulley 2 and the motor pulley 4 are also maximized.

  Further, from the equations (1) and (2), even if the rotational speed of the compressor pulley 2 varies, the rotational speed of the motor pulley 4 does not vary, that is, against the periodic variation of the rotational speed of the compressor pulley 2. The angular frequency ω1 * at which the anti-resonance state where the fluctuation of the rotation speed of the motor pulley 4 stops can be obtained, and the following equation is obtained.

  In this two-degree-of-freedom rotational vibration system, the relationship between the fluctuation phase difference of each pulley and the belt tension piece amplitude with respect to the angular frequency of the axial load torque is as shown in FIG. From FIG. 5, it can be seen that the tension of the belt 5 increases in the range of ω> ω1 * where the fluctuation of each pulley is in the opposite phase (180 °). Therefore, when the rated rotational speed of the compressor body 1 is Nc [min-1], each parameter may be set so that the following equation is established.

  There are the following three methods for establishing the above relationship.

1) Increasing the diameter rM of the motor pulley 4 2) Increasing the spring constant K of the belt 5 3) Decreasing the moment of inertia IM of the motor pulley 4 Considering 1), the rated rotational speed of the compressor body 1 is maintained. In order to do so, the diameter of the compressor pulley 2 also needs to be increased.

  In general, the inertia moment of the pulley increases in proportion to the pulley diameter. That is, as the diameter rM of the motor pulley 4 is increased, the inertia moment IM of the motor pulley 4 is also increased. Further, in order to reduce the inertia moment IM of the motor pulley 4, it is necessary to reduce the diameter rM of the motor pulley 4. Therefore, the above 1) and 3) are contradictory.

  Therefore, considering 2), in addition to changing the type of the belt 5, it is possible to take measures such as making multiple belts. At this time, if the spring constant of each belt is k and the number of belts is NB, K = NBk. Therefore, the belt number NB may be set without changing the diameter rM of the motor pulley 4 and the inertia moment IM of the motor pulley 4 so as to satisfy the relationship of Expression (6).

  As described above, according to the present embodiment, the angular frequency of the compressor body 1 is set as described above in consideration of the natural angular frequency of the two-degree-of-freedom rotational vibration system composed of two pulleys by taking the belt 5 as a spring. By doing so, it is possible to suppress excessive tension on the belt and to ensure the reliability of the belt for a long time.

  In the present embodiment, a description will be given of a configuration capable of suppressing the fluctuation in the tension of the belt 5 even when the expression (6) cannot be satisfied due to the restriction of the rated rotational speed of the compressor body 1 or the like. The same components as those of the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

  FIG. 5 shows that in the range of ω> √2 * ω2, the tension fluctuation can be suppressed even though the fluctuation of each pulley is in reverse phase. Therefore, each parameter may be set so that the following equation is established.

  There are the following three methods for establishing the above relationship.

1) Decreasing the diameter rc of the compressor pulley 2 2) Decreasing the spring constant K of the belt 5 3) Increasing the inertia moment Ic of the compressor pulley 2 Considering 1) above, the compressor body 1 In order to maintain the rated rotational speed, the diameter of the compressor pulley 2 needs to be reduced. However, the inertia moment of the pulley usually increases in proportion to the pulley diameter. That is, as the diameter rc of the compressor pulley 2 is reduced, the inertia moment rc of the compressor pulley 2 is also reduced. Further, in order to increase the inertia moment Ic of the compressor pulley 2, it is necessary to increase the diameter rc of the compressor pulley 2. Therefore, the above 1) and 3) are contradictory.

  Therefore, in the present embodiment, as shown in FIG. 6, a flywheel 2a having a diameter ΦD1 larger than the effective diameter ΦD0 is provided for the compressor pulley 2. By doing so, the inertia moment of the compressor pulley 2 can be increased without increasing the diameter of the compressor pulley 2. Note that the arrangement of the flywheel is preferably provided on the side closer to the compressor body 1 than the center of the belt 5 in consideration of the alignment adjustment workability of the compressor pulley 2 and the motor pulley 4.

  As described above, according to the present embodiment, even when the expression (6) cannot be satisfied due to the limitation of the rated rotational speed, the belt can be obtained by setting the angular frequency of the compressor body 1 as described above. It is possible to suppress excessive tension on the belt and to ensure the reliability of the belt for a long time.

  In the first and second embodiments, the equations (6) and (7) indicate that the target compressor is a single cylinder as shown in FIG. 1, and the angular frequency of the axial load torque is It is assumed that it matches the rotational angular frequency of the compressor body 1. However, for example, when the target compressor has two cylinders and the respective phases are shifted by 180 degrees in the crank angle, the angular frequency of the axial load torque is equal to the rotation frequency of the compressor body 1. Doubled. Similarly, in an M cylinder machine, it becomes M times. Therefore, in this case, the equations (6) and (7) are rewritten as follows.

  Up to this point, the reciprocating compressor in which the piston 25 reciprocates in the cylinder 22 has been described as an example of the compressor body 1 as an embodiment of the present invention. However, the present invention is not limited to the reciprocating compressor, and the scroll The present invention can also be applied to other fluid machines such as a hydraulic fluid machine and a screw compressor.

  The embodiments described so far are merely examples of implementation in carrying out the present invention, and the technical scope of the present invention is not limitedly interpreted by these. That is, the present invention can be implemented in various forms without departing from the technical idea or the main features thereof.

DESCRIPTION OF SYMBOLS 1 ... Compressor body 2 ... Compressor pulley 3 ... Motor 4 ... Motor pulley 5 ... Belt 7 ... Tank 21 ... Crank chamber 22 ... Cylinder 23 ... Cylinder Head 24 ... Crankshaft 24a ... Balance weight 25 ... Piston

Claims (5)

A compressor body for compressing fluid;
A motor for driving the compressor body;
A compressor pulley that rotates as the compressor body is driven;
A motor pulley that rotates as the motor is driven;
And a belt for transmitting power of the motor pulley to the compressor pulley,
The compressor body has M cylinders,
When the belt is used as a spring and the motor pulley is used as a weight, the angular frequency at the anti-resonance point at which the fluctuation of the rotation speed of the motor pulley stops with respect to the periodic fluctuation of the rotation speed of the compressor pulley is ω1 *. The compressor satisfies Mω <ω1 *, where ω is the rotational angular frequency of the compressor body. 2. The angular frequency ω <b> 1 * at the anti-resonance point satisfies the following expression when the diameter of the motor pulley is rM, the moment of inertia of the motor pulley is IM, and the spring constant of the belt is K: 2. Compressor.

A compressor body for compressing fluid;
A motor for driving the compressor body;
A compressor pulley that rotates as the compressor body is driven;
A motor pulley that rotates as the motor is driven;
A belt for transmitting the power of the motor pulley to the compressor pulley,
The compressor body has M cylinders,
When the belt is a spring and the motor pulley is a weight, the rotational frequency of the compressor body is ω, and one of the natural angular frequencies of the rotational vibration system including the compressor pulley, the motor pulley, and the belt is ω2. when, to satisfy the following equation,

Ω2, which is one of the natural angular frequencies, is the compressor pulley diameter rC, the motor pulley diameter rM, the compressor pulley inertia moment IC, the motor pulley inertia moment IM, and the belt spring. compressor according to claim Succoth satisfy the following formula when the constant set to K.

  The compressor according to claim 3, wherein a flywheel having a diameter larger than that of the compressor pulley is provided on the compressor pulley.   The compressor according to claim 3, wherein the flywheel is provided at a position closer to the compressor body than the belt.

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