JP2015203338A - compressor - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a compressor capable of efficiently restraining vibration at startup and a load to respective parts accompanied by the vibration even in a relatively large reciprocating compressor having large reactive force at startup.SOLUTION: A compressor includes a compressor body 1 compressing fluid, a motor 3 for driving the compressor body, a compressor pulley 2 rotated in accordance with driving of the compressor body, a motor pulley 4 rotated in accordance with driving of the motor, a belt 5 transmitting power of the motor pulley to the compressor pulley, a frame 7 in which the compressor body and the motor are arranged, and a vibration control member 6 for reducing vibration of the frame. A following expression is satisfied when a natural frequency in a case where the belt is used as a spring and the motor pulley is used as a weight is ω and a natural frequency in a case where the vibration control member is used as the spring and the frame and the ones arranged in the frame are used as weights is ωn.

Description

  The present invention relates to a compressor.

  As a background art of this technical field, there is Patent Document 1. The compressor of Patent Document 1 reduces vibration by supporting the compressor main body and the support member with a plurality of vibration-proof rubbers arranged in an inclined manner.

JP 2010-37961 A

  In a rotating machine represented by a compressor, vibrations caused by imbalance of rotating parts and torque fluctuations and loads generated in the machine parts due to the vibrations become a problem. For this reason, the suspension base that fixes the machine is supported by an anti-vibration spring, etc., and the relationship between the natural frequency of the system determined by the sprung mass and the spring constant and the rotational speed of the rotating machine is adjusted. A method of suppressing the amplitude of the structure and vibration transmission to the ground is taken.

  Patent Document 1 aims at reducing vibrations during steady rotation and stop deceleration by supporting a main body by a plurality of anti-vibration springs that are inclined and arranged for a reciprocating compressor as an example of a rotary machine.

  However, the vibration load actually increases not during the operation mode such as during steady rotation or during stop deceleration, but most of the vibration load is during start-up. Vibrations during steady rotation or stop deceleration tend to greatly shake the entire machine and absorb vibration displacement as a whole. On the other hand, the vibration at the time of starting is locally generated by the starting torque of the motor that is impacted on the rotating machine. Therefore, the vibration at the time of starting cannot be suppressed only by devising the support by the anti-vibration spring.

  In view of the above problems, the present invention provides a compressor capable of suppressing vibration at start-up by a simple method by setting various parameters in consideration of the natural frequency when the belt is viewed as a spring. Objective.

  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 body having an n-cylinder cylinder and compressing a 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 that transmits the power of the motor pulley to the compressor pulley; the compressor body and the motor; And a vibration isolating member for reducing vibration of the frame. The natural frequency when the belt is a spring and the compressor pulley and the motor pulley are weights is ω, and the vibration isolating member is a spring. The following equation 1 is satisfied, where ωn is the natural frequency when the frame and the one arranged on the frame are weights.

  ADVANTAGE OF THE INVENTION According to this invention, it becomes possible to provide the compressor which can suppress the vibration at the time of a start by a simple system.

  Problems, configurations, and effects other than those described above will become apparent from the following description of embodiments.

It is a block diagram of the compressor in Example 1 of this invention. It is a block diagram inside the compressor main body in Example 1 of this invention. It is a graph which shows the correlation of the crankshaft angle of the compressor in Example 1 in this invention, a cylinder internal pressure, and a shaft load torque. It is a graph which shows the correlation of the frequency ratio of a compressor and amplitude magnification in Example 1 of this invention. It is a graph which shows the correlation of the pulley rotation speed and flame | frame stress at the time of starting of the compressor in Example 1 in this invention. It is the graph which expanded a part of FIG. It is a block diagram of the compressor in the modification of Example 1 of this invention. It is a block diagram of the compressor in Example 2 of this invention. It is a graph which shows the correlation of the frequency ratio of a compressor and amplitude magnification in Example 2 of this invention.

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

  First, the basic compressor configuration in this embodiment, the idea of conventional vibration countermeasures, and the concept of this embodiment will be described.

  FIG. 1 shows a schematic diagram of a basic compressor in the present embodiment. FIG. 2 shows a schematic diagram of the internal structure of the compressor body in FIG.

  The compressor shown in FIG. 1 includes a support frame 7 that is supported on a floor surface or the like via an anti-vibration spring 6, and a compressor body 1 that is installed on the frame 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 pistons and cylinders for simplification of explanation, but there are cases where a plurality of pistons and cylinders are arranged in series or radially with respect to the crankshaft. is there.

  The compressor body 1 is fixed to the upper surface of the support frame 7 with the crankshaft 24 disposed in parallel with the rotation shaft of the motor 3, and the compressor pulley 2 is attached to the crankshaft 24 and the motor pulley 4 is attached to the motor shaft. Is fixed. A compressor pulley 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, thereby promoting heat dissipation of the compressor main body 1.

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

  Here, the excitation frequency ω due to the reciprocating inertia force in the steady state will be described.

  The reciprocating compressor as shown in FIG. 1 and FIG. 2 performs the operation of compressing fluid discontinuously on its mechanism, so that the cylinder internal pressure applied to the piston 25 and the shaft generated on the crankshaft 24 thereby. The load torque varies greatly while the crankshaft 24 makes one revolution. FIG. 3 shows an example of changes in the rotation angle of the crankshaft 24 and the accompanying cylinder internal pressure and shaft load torque.

  Such a large change in the shaft load torque increases the effective value by changing the current value of the motor 3 and becomes a load on the electric system. Therefore, the reciprocating compressor generally has a flywheel for smoothing the torque fluctuation, and in particular, this function is often achieved by giving the compressor pulley 2 a moment of inertia. This is because the moment of inertia inevitably increases when trying to give the compressor pulley 2 a large blade for cooling the compressor main body 1 and rigidity to withstand the tension of the transmission belt 5. It is because it is convenient from.

By the way, when the above-described reciprocating compressor is operated, the compressor main body 1 vibrates due to the inertial force caused by the reciprocating motion of components such as the piston 25. In order to suppress the load caused by this vibration, a balance weight 24a is attached to the crankshaft 24, and so-called half balancing is performed in which a part of the reciprocating inertia force is canceled by a centrifugal force accompanying the rotation. Furthermore, for the unbalanced component that cannot be suppressed even by this half balancing, the frame 7 that fixes the compressor body 1 is supported by a vibration isolating spring (vibration isolating member) 6, A countermeasure is taken by adjusting the relationship of the total mass of the sprung structure composed of the compressor body 1, the compressor pulley 2, and the frame 7. That is, when the total mass of the sprung structure is m, the spring constant of the vibration isolating spring is k, the maximum value of the excitation force is F0, and the angular frequency of the excitation is ω, (Vertical direction) equation of motion is

It is written. From the above equation, the following equation is obtained as the natural angular frequency ωn when the vibration-proof spring 6 is a spring and the frame 7 and the frame 7 are weights.

Further, the following equation is obtained as the amplitude magnification X of this system.

  Note that xst is the deflection of the spring when the excitation force F0 acts statically, and xst = F0 / k.

  FIG. 4 shows the correlation between ω / ωn and amplitude magnification in this one-degree-of-freedom vibration system. In the figure, it can be seen that when ω / ωn = 1, that is, when the rotation speed of the compressor body 1 becomes equal to the natural frequency, resonance occurs and the amplitude becomes infinite. In addition, the amplitude magnification is 1 at ω / ωn = √2, and in the region beyond that, the amplitude of the sprung structure can be suppressed to be equal to or less than the amplitude of the excitation by the reciprocating inertia force. In other words, the anti-vibration effect by the anti-vibration spring is exhibited. On the other hand, in the region where ω / ωn <√2, the amplitude magnification is 1 or more. Therefore, the closer to the resonance point, the worse the vibration of the sprung structure.

Therefore, as a countermeasure against vibrations of the system, the relationship of the following equation between the natural angular frequency ωn of this system and the rotation speed of the compressor body 1, that is, the ratio ω / ωn of the angular frequency ωn of vibration due to the reciprocating inertia force is Set to meet.

  In addition, the excitation frequency ω due to the reciprocating inertial force at the time of steady rotation is a number of cylinders that is the number of rotations of the main body. In the present embodiment, for the sake of simplicity, the compressor body 1 is assumed to be a single-cylinder machine having one cylinder, but there is actually a reciprocating compressor having two or more cylinders. In this case, the excitation frequency ω at the time of steady rotation is calculated by multiplying the number of cylinders by the number of main body rotations.

  However, it is rare that excessive vibration occurs during steady rotation. In most cases, excessive vibration occurs at the time of starting, and before the rotation of the compressor body 1 reaches the steady rotational speed. Therefore, in this embodiment, the excitation frequency ω at the start is calculated, and various parameters are set so that the vibration at the start can be suppressed.

  Note that even a reciprocating compressor having two or more cylinders is calculated using the excitation frequency ω.

  A method for calculating the excitation frequency ω at the start will be described with reference to FIGS.

  FIG. 5 shows an example in which, at the time of starting a certain reciprocating compressor, the number of rotations thereof and the stress generated in the welded portion of the fixed frame 7 of the compressor body 1 are simultaneously measured. From the figure, it can be seen that the stress amplitude reaches a peak at a time of several tens of msec from the start before the rotation speed of the compressor reaches the steady rotation. Moreover, it can be confirmed that the amplitude of the strain after reaching the steady rotation is smaller than that at the start.

  This is because vibration during steady rotation greatly shakes the whole sprung structure shown in FIG. 1 and tends to absorb vibration displacement as a whole. This is generated locally by the starting torque of the applied motor 3, and as a result, the load applied to the frame 7 or the like is much more severe at the time of starting.

  Hereinafter, a mechanism for generating the load at the start will be described.

  FIG. 6 is an enlarged view of the rising portion of the rotational speed of the motor pulley 4 in FIG. From the figure, it can be seen that the number of revolutions of the motor pulley 4 rises while slightly vibrating at the same period as the stress amplitude of the frame 7. On the other hand, the rotation speed of the compressor pulley 2 does not show the vibration like the motor pulley 4 and starts to rotate smoothly.

  This suggests that the compressor pulley 2 having a large moment of inertia cannot immediately follow the rotation at the moment when the motor pulley 4 starts to rotate at the start, and a slight rotational delay occurs with respect to the motor pulley 4. . Due to this delay, the transmission belt 5 is stretched, and the force due to the spring property resists the torque of the motor 3, so that the rotational speed of the motor pulley 4 slightly decreases. Next, when the compressor pulley 2 starts to rotate with a delay, the elongation of the transmission belt 5 decreases and the spring force decreases, so that the rotational speed of the motor pulley 4 is recovered. By the series of operations as described above, the rotational speed of the motor pulley 4 increases while vibrating.

  In the above vibration phenomenon, the motor torque that rises as a step input when the motor 3 is started excites the natural vibration of the two-degree-of-freedom rotational vibration system that uses the motor pulley 4 and the compressor pulley 2 as weights and the transmission belt 5 as a spring. In other words.

  Therefore, in this embodiment, the natural frequency is obtained by paying attention to the above-described vibration phenomenon, and various parameters are set based on the conditions that can suppress the vibration at the start.

The equation of motion of the vibration system is as follows.

  Where I is the moment of inertia of the pulley, θ is the rotation angle of the pulley, x is the extension of the transmission belt 5, c is the damping coefficient of the transmission belt 5, kB is the spring constant of the transmission belt, and the subscripts M and V are the motor pulleys, respectively. The compressor pulley is shown.

Since the extension x of the transmission belt is considered as the difference in arc length with respect to the rotation angles of the motor pulley and the compressor pulley,

  Where D is the pitch circle diameter of the pulley.

For simplification, the natural angular frequency is obtained after excluding the attenuation term, and the following equation is obtained.

  Since ω1 = 0 has no meaning as a solution in the rotational vibration system, ω2, which is the natural angular frequency of the second-order mode, is the eigenvalue of the vibration described above.

  By the way, the rotational vibration phenomenon of the motor pulley, when viewed from the compressor body 1, means that the entire compressor body 1 is vibrated at a cycle of ω2 with the crankshaft as the excitation point. Therefore, when ω2 and the natural frequency ωn of the sprung structure represented by (1) are close, a resonance phenomenon occurs as described above. As a result, a stress waveform having a sharp peak as shown in FIGS. 5 and 6 is generated.

  Therefore, in this embodiment, in order to reduce the vibration at the time of starting, each parameter was adjusted so that the above-described equation (4) when ω = ω2 was satisfied.

  In the above description, as shown in FIG. 1, the motor 3 serving as the excitation source of the two-degree-of-freedom rotational vibration system including a pulley and a belt is insulated from the frame on which the compressor body 1 is fixed (vibration It was described as being arranged (so as not to communicate).

  Here, if the compressor body 1 and the motor 3 exist on the same spring as shown in FIG. 7, if the frame 7 is regarded as a rigid body, the force caused by the vibration of the pulley / belt system will cause the reaction inside the rigid body to react. Because it balances with the force, it does not affect the vibration of the sprung mass system.

  However, since the frame that supports the actual compressor is made of welded sheet metal or the like, it is considered to be a continuous body with some elasticity, not a rigid body. A mode in which the upper mass system vibrates occurs. Therefore, since the same frequency ratio relationship can be applied to the compressor having the configuration as shown in FIG. 7, it is possible to reduce the vibration at the start by setting the equation (4).

In this figure, for simplicity of explanation, the sprung structure is treated as a one-degree-of-freedom vibration system that vibrates only in the vertical direction. However, the relationship of equation (4) is replaced with a vibration system of any degree of freedom. It is also possible to think. For example, when the compressor main body 1 and the motor 3 are juxtaposed on the same base as shown in FIG. 8, the natural angular frequency ωn satisfies the following equation when this is replaced with a two-degree-of-freedom vibration system. .

Here, k1 and k2 are the spring constants of the left and right anti-vibration springs supporting the frame 7 in FIG. 8, I is the moment of inertia of the sprung structure, m is the mass of the sprung structure, and s is the sprung This is the horizontal distance between the center of gravity G and point E of the structure. Further, the position of point E can be obtained by the following equation.

  It is possible to reduce the vibration at the time of starting by adjusting each parameter so as to satisfy the natural frequency obtained from the equation (10) or satisfying the equation (4).

  Note that the excitation frequency ω at the time of start is the same as in equation (9).

  As described above, according to this embodiment, various parameters are set in consideration of the natural frequency of the two-degree-of-freedom rotational vibration system in which the motor pulley 4 and the compressor pulley 2 are weights and the transmission belt 5 is a spring. Thus, vibration at the start of the compressor can be effectively suppressed.

  In this embodiment, even when the relationship between the natural frequency ωn of the sprung mass system and the natural frequency ω2 of the pulley / belt system does not satisfy the formula (4), the configuration can reduce the vibration at the start. explain.

  Since the parameters in equation (4) are greatly related to the specifications of the compressor body 1, there are many cases where there is no adjustment allowance. For example, the diameter of the compressor pulley 2 is determined from the diameter of the fan for cooling the cylinder head, and as described above, in order to cancel the torque fluctuation, the inertia moment as the rotating body is required to be a certain level or more. On the other hand, the motor pulley 4 has a diameter ratio with the compressor pulley 2 determined by the design value of the number of revolutions of the compressor body 1, and the transmission belt 5 is a commercially available product because the type is limited based on the transmission power. In some cases, the spring constant is naturally limited. Taking these matters into consideration, there is a case where the balance of the entire product is deteriorated, for example, an extremely large pulley must be used if Equation (2) is to be satisfied.

  Therefore, in this embodiment, a method for taking measures against the vibration phenomenon of the motor pulley from the electrical side, not the mechanical measures as described above, will be described.

  The natural vibration of the pulley-belt system in question is excited by the starting torque of the motor 3, and the vibration amplitude is proportional to the magnitude of the starting torque. On the other hand, since the starting torque of the motor 3 is proportional to the supply voltage, it is possible to reduce the starting torque and the vibration amplitude by performing electrical control so as to reduce the supply voltage only for a predetermined time from the start.

  Therefore, in this embodiment, the vibration at the time of starting is suppressed by using a load reducing circuit that reduces the supply voltage to the motor 3 at the time of starting.

  The electrical control by the load reducing circuit at this time has an effect of protecting the wiring system by preventing an excessive current, and various systems such as a star delta system and a condorfa system have been put into practical use. By installing these starting methods, vibrations may be able to be suppressed even in the configuration in which equation (4) is not satisfied.

Hereinafter, a vibration reducing method when the star delta method is used will be described as an example. Taking the compressor configuration shown in FIG. 1 as an example, equation (1) is rewritten as follows.

  Here, a is the rate of change of the motor starting torque due to the change of the starting method. The torque fluctuation rate at the start is the square of the supply voltage fluctuation rate at the start. For example, in the star delta start system, the supply voltage is suppressed to 1 / √3 by changing the winding connection of the motor 3 for a certain period of time from the start, so the change rate of the start torque is a = 1/3. .

From the above formula,

From the relationship of equation (3),

FIG. 9 shows the correlation between the amplitude magnification X and ω / ωn when a = 1 and 1/3. From the figure, the amplitude magnification X is lowered by changing the starting method of the motor 3, and as a result, the condition of the equation (2) is relaxed as the following equation.

  Therefore, if the equation (2) cannot be satisfied, the parameters in the equation may be determined so as to satisfy the equation (15) by changing the starting method of the motor 3. In this embodiment, the star delta start method is taken as an example, and a = 1/3. However, in actuality, the start method may be selected so that the formula (15) is satisfied, and a is from 0 to 1. Can take any value.

  As described above, according to the present embodiment, it is possible to effectively suppress vibration at the time of starting the compressor without greatly changing the specifications of the compressor body 1 such as the compressor pulley 2 and the motor pulley 4.

DESCRIPTION OF SYMBOLS 1 ... Compressor main body 2 ... Compressor pulley 3 ... Motor 4 ... Motor pulley 5 ... Transmission belt 6 ... Anti-vibration spring 7 ... Support frame 21 ... Crank chamber 22 ... Cylinder 23 ... Cylinder head 24 ... Crank shaft 24a ... Balance weight 25 ... Piston

Claims (7)

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;
A frame on which the compressor body and the motor are disposed;
A vibration isolating member that reduces vibration of the frame,
The natural frequency when the belt is a spring and the motor pulley is a weight is ω, the natural vibration frequency is when the vibration isolator is a spring, and the frame and the one disposed on the frame are weights, and the natural frequency is ωn. When the compressor is satisfied, the following formula 1 is satisfied.

The compressor according to claim 1, wherein ωn is given by the following equation (2).

However, k is the spring constant of the vibration isolating member, and M is the total mass of the frame and what is disposed on the frame. The compressor according to claim 1, wherein ω is given by Equation 3 below.

Where kb is the spring constant of the belt, IM is the moment of inertia of the motor pulley, IV is the moment of inertia of the compressor pulley, DM is the diameter of the motor pulley, and DV is the diameter of the compressor pulley.
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;
A frame on which the compressor body and the motor are disposed;
A vibration isolating member for reducing vibration of the frame;
A load reduction circuit that reduces the supply voltage to the motor at the start of the motor,
The natural frequency when the belt is a spring and the motor pulley is a weight is ω, the vibration isolating member is a spring, and the natural frequency when the frame and the one disposed on the frame are a weight is ωn, The compressor satisfying the following expression 1 when a reduction rate of the motor starting torque of the load reducing circuit is a.

  The compressor according to claim 4, wherein a = 1/3. The compressor according to claim 4 or 5, wherein ωn is given by the following equation (4).

However, k is the spring constant of the vibration isolating member, and M is the total mass of the frame and what is disposed on the frame. 6. The compressor according to claim 4, wherein ω is given by the following expression (5).

Where kb is the spring constant of the belt, IM is the moment of inertia of the motor pulley, IV is the moment of inertia of the compressor pulley, DM is the diameter of the motor pulley, and DV is the diameter of the compressor pulley.

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