JP2009008028A - Variable capacity turbine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a variable capacity turbine, suppressing breakage of a turbine blade caused by nozzle wake resonance, without changing a device constitution, at low cost. <P>SOLUTION: The variable capacity turbine comprises a turbine wheel disposed to a gas inlet passage and rotatable when a plurality of turbine blades receive exhaust gas; and a plurality of variable nozzle vane changing a flow rate of exhaust gas flowing into the turbine wheel. In each of the turbine blades, allowable stress σc allowing a load by its rigidity is set. When σb is centrifugal stress generated by rotation of the turbine and σa is vibration stress caused by nozzle wake resonance generated by the rotation of the turbine wheel, characteristic frequency for each turbine blade is set so that a total sum (σa+σb) is below the allowable stress σc, and the nozzle wake resonance generated by rotation of the turbine wheel is generated at a 70% or less of the maximum revolution number Nmax of the turbine wheel. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a variable capacity turbine capable of changing the flow velocity of exhaust gas flowing into a turbine wheel by a plurality of variable nozzle vanes.

  Conventionally, as this type of variable capacity turbine, the exhaust gas is received by a plurality of rotor blades (turbine blades), and the turbine wheel that rotates and rotates around the rotating shaft to change the flow rate of the exhaust gas that flows in. A plurality of nozzle vanes that can be used (variable nozzle vanes) are known (see Patent Document 1). In this variable capacity turbine, the rotation axis of each nozzle vane is positioned within 1/3 L from the gas outlet trailing edge of each nozzle vane with respect to the length L of the nozzle vane in order to prevent damage to the moving blade due to nozzle wake resonance. It is arranged.

  Here, the blade is damaged by the nozzle wake resonance because the sum of the centrifugal stress caused by the rotational drive of the turbine wheel and the vibration stress caused by the nozzle wake resonance exceeds the allowable stress that the blade can accept. is there. Therefore, according to the above configuration, the vibration stress applied to the moving blade can be reduced, so that the sum of the centrifugal stress and the vibration stress can be contained within the allowable stress of the moving blade. The blade is prevented from being damaged by the wake resonance.

Japanese Patent Laid-Open No. 10-205340

  However, in the above-described configuration, since the rotation axis of the nozzle vane is disposed on the trailing edge side of the nozzle vane, a large torque is required when rotating the nozzle vane. For this reason, it is necessary to mount a large actuator to rotationally drive the nozzle vanes. This increases the cost and increases the size of the variable capacity turbine.

  Accordingly, an object of the present invention is to provide a variable capacity turbine capable of suppressing damage to turbine blades due to nozzle wake resonance at a low cost without changing the device configuration.

  The variable capacity turbine of the present invention is disposed in a gas inflow passage, and is rotatable by receiving exhaust gas by a plurality of turbine blades, and is disposed on the upstream side of the turbine wheel in the gas inflow passage and flows into the turbine wheel. A plurality of variable nozzle vanes capable of changing the flow rate of exhaust gas to be generated. Each turbine blade is set with an allowable stress that allows a load depending on its rigidity, and is generated by centrifugal stress generated by rotation of the turbine wheel and rotation of the turbine wheel. The natural frequency of each turbine blade is such that the sum of the vibration stress due to the nozzle wake resonance is less than the allowable stress and the nozzle wake resonance caused by the rotation of the turbine wheel occurs at 70% or less of the maximum rotation speed of the turbine wheel. Is set.

  In this case, it is preferable that the natural frequency of each turbine blade is set to a natural frequency corresponding to a vibration mode in which the inflow end of each turbine blade into which exhaust gas flows flows becomes a vibration antinode. .

  Also, in this case, the natural frequency of each turbine blade is calculated by multiplying the resonance frequency calculated by multiplying the number of variable nozzle vanes by the maximum number of rotations per second of the turbine wheel by a percentage of 70% or less. It is preferred that

  In this case, it is preferable that the natural frequency of each turbine blade is set so that the nozzle wake resonance occurs at 50% or more of the maximum rotational speed of the turbine wheel.

  The variable capacity turbine according to the present invention has an effect that the damage of each turbine blade can be suppressed at low cost without changing the device configuration.

  Hereinafter, a variable capacity turbine according to the present invention will be described with reference to the accompanying drawings. Note that the present invention is not limited to the embodiments.

  Here, FIG. 1 is a cross-sectional view of the turbocharger 1, and FIG. 2 is a schematic view around the turbine wheel as seen from the axial direction. FIG. 3 is a graph relating to the rotational speed of the turbine wheel and the total stress received by the turbine blade in the conventional configuration, and FIG. 4 is a fatigue limit diagram of the turbine blade. Further, FIG. 5 is a graph regarding the rotational speed of the turbine wheel and the total stress received by the turbine blade in the present embodiment, and FIG. 6 is a Campbell diagram of the turbine blade.

  First, prior to the description of the variable capacity turbine, a turbocharger to which the variable capacity turbine is applied will be described with reference to FIG. The turbocharger 1 is a so-called VG (Variable Geometry: variable capacity) type turbocharger. The turbocharger 1 includes a variable capacity turbine 2 disposed on the right side in the figure, a compressor 3 disposed on the left side in the figure, and a rotor shaft 4 interposed between the variable capacity turbine 2 and the compressor 3. ing. The variable capacity turbine 2 rotates when exhaust gas discharged from the internal combustion engine flows in. The rotor shaft 4 is accommodated in the center housing 5 and transmits the rotational force of the variable capacity turbine 2 to the compressor 3. The compressor 3 takes in air and compresses the air by the rotational force transmitted from the rotor shaft 4 and sends it to the internal combustion engine. That is, the turbocharger 1 is configured such that when the variable capacity turbine 2 is operated by the exhaust gas from the internal combustion engine, the rotational force is transmitted to the compressor 3 via the rotor shaft 4 so that the compressor 3 is operated. ing.

  The compressor 3 includes a compressor wheel 10 fixed to one end of the rotor shaft 4 and a compressor housing 11 that houses the compressor wheel 10. Further, the compressor housing 11 is formed with an air inflow path 15 for sending the taken-in air into the combustion chamber.

  The compressor housing 11 is integrally formed by a compressor cylindrical portion 11a disposed coaxially with the shaft axis direction of the rotor shaft 4 and a compressor ring portion 11b provided in an annular shape along the outer peripheral surface of the compressor cylindrical portion 11a. ing.

  The compressor cylindrical portion 11 a is formed in a cylindrical shape, and an intake inlet 13 is circularly opened in the shaft axial direction of the rotor shaft 4 on the outer end surface on the left side in the drawing. In addition, a compressor wheel accommodating portion 14 for accommodating the compressor wheel 10 is formed inside the compressor cylindrical portion 11 a and communicates with the intake inlet 13.

  The compressor ring portion 11b is formed in an annular shape, and the compressor side scroll 12 is formed in a spiral shape along the circumferential direction inside the compressor ring portion 11b. Although not shown, an intake outlet is formed in the compressor ring portion 11b. The compressor side scroll 12 communicates with the compressor wheel accommodating portion 14 on the upstream side of the air inflow path 15, and the intake outlet communicates with the downstream side of the air inflow path 15.

  And the air inflow path 15 formed in the compressor housing 11 is comprised by the order of the intake inlet 13, the compressor wheel accommodating part 14, the compressor side scroll 12, and the intake outlet from the upstream.

  When the compressor wheel 10 rotates with the rotation of the rotor shaft 4, air is taken in from the intake inlet 13, and the taken-in air flows into the compressor wheel housing portion 14 and is compressed by the compressor wheel 10, and then the compressor side scroll 12. Is exhausted from the intake outlet.

  The center housing 5 is formed in a cylindrical shape, and a shaft hole 16 into which the rotor shaft 4 is inserted is formed through the shaft center. A pair of bearings 17 composed of bearings or the like are disposed on the inner periphery of the shaft hole 16, and the rotor shaft 4 is rotatably supported by the pair of bearings 17. The compressor housing 11 is fixed to the left end surface of the center housing 5 in the figure, and a turbine housing 21 to be described later is fixed to the right end face in the figure.

  The rotor shaft 4 has the above-described compressor wheel 10 fixed to one end thereof, and a turbine wheel 20 described later fixed to the other end thereof. For this reason, the rotor shaft 4, the compressor wheel 10, and the turbine wheel 20 rotate integrally.

  Here, the variable capacity turbine 2 in the present embodiment will be described in detail. The variable capacity turbine 2 is disposed along the periphery of the turbine wheel 20, the turbine wheel 20 fixed to the other end of the rotor shaft 4, the turbine housing 21 that houses the turbine wheel 20, and flows into the turbine wheel 20. And a plurality of variable nozzle vanes 22 capable of changing the flow rate of the exhaust gas to be discharged. Further, the turbine housing 21 is formed with a gas inflow path 30 for discharging exhaust gas sent from the combustion chamber.

  The turbine housing 21 is configured in substantially the same manner as the compressor housing 11, and is provided in a ring shape along the outer peripheral surface of the turbine cylindrical portion 21a and the turbine cylindrical portion 21a disposed coaxially with the shaft axis direction of the rotor shaft 4. The turbine ring portion 21b is integrally formed.

  The turbine cylindrical portion 21a is formed in a cylindrical shape, and an exhaust outlet 24 is circularly opened in the shaft axial direction of the rotor shaft 4 on the outer end surface on the left side in the drawing. Further, a turbine wheel accommodating portion 25 that accommodates the turbine wheel 20 is formed inside the turbine cylindrical portion 21 a and communicates with the exhaust outlet 24.

  The turbine ring portion 21b is formed in an annular shape, and the turbine-side scroll 23 is formed in a spiral shape along the circumferential direction inside the turbine ring portion 21b. Moreover, although illustration is abbreviate | omitted, the exhaust_gas | exhaustion inlet_port | entrance is formed in the turbine ring part 21b. In the turbine side scroll 23, the above-described turbine wheel accommodating portion 25 communicates with the downstream side of the gas inflow path 30, and the exhaust inlet communicates with the upstream side of the gas inflow path 30.

  The gas inflow path 30 formed inside the turbine housing 21 is configured from an upstream side to an exhaust inlet, a turbine side scroll 23, a turbine wheel accommodating portion 25, and an exhaust outlet 24 in this order.

  When exhaust gas flows from the exhaust inlet, the exhaust gas passes through the turbine-side scroll 23 and flows into the turbine wheel housing portion 25, and the turbine wheel 20 is rotationally driven by the exhaust gas that has flowed in. Thereafter, the exhaust gas is discharged from the exhaust outlet 24.

  As shown in FIG. 2, the plurality of variable nozzle vanes 22 are disposed in the gas inflow path 30 between the turbine side scroll 23 and the turbine wheel accommodating portion 25 and are disposed along the periphery of the turbine wheel 20. ing. Each variable nozzle vane 22 is formed in a wing shape and rotates between an open position and a closed position about a rotation shaft 35, thereby changing the flow rate of the exhaust gas flowing into the turbine wheel housing portion 25. That is, when each variable nozzle vane 22 rotates to the closed position, the gas area in the gas inflow path 30 is reduced, and the gas pressure in the gas inflow path 30 upstream of the variable nozzle vane 22 is increased. The flow rate of the exhaust gas flowing into the section 25 is increased. On the other hand, when each variable nozzle vane 22 is rotated to the open position, the gas area in the gas inflow path 30 is increased, so that the gas pressure in the gas inflow path 30 on the upstream side of the variable nozzle vane 22 is reduced. The flow rate of the exhaust gas flowing into the portion 25 becomes slow. The rotating shaft 35 is provided coaxially with the shaft axis direction and is pivotally supported by the turbine housing 21.

  The turbine wheel 20 includes a wheel body 40 and a plurality of turbine blades 41 provided radially from the axis of the wheel body 40, and the inflowing exhaust gas is received by the plurality of turbine blades 41 and rotated. It is configured to

  At this time, centrifugal stress generated by the rotation of the turbine wheel 20 and vibration stress generated by the plurality of variable nozzle vanes 22 are applied to each turbine blade 41. For this reason, each turbine blade 41 is set with a predetermined allowable stress based on its rigidity so as to allow a total load of centrifugal stress and vibration stress.

  However, as shown in FIG. 3, when nozzle wake resonance occurs, vibration stress increases at the resonance point, and it becomes difficult to keep the sum of centrifugal stress and vibration stress within allowable stress. The rigidity of the turbine blades 41 cannot withstand the rotational load, and each turbine blade 41 may be damaged. For this reason, in the variable capacity turbine 2 of the present embodiment, each turbine blade 41 is set to a predetermined natural frequency. Hereinafter, a series of procedures for calculating the natural frequency of each turbine blade 41 to be set will be described.

  FIG. 4 is a fatigue limit diagram of the turbine blade 41 obtained by applying material evaluation, in which the vertical axis represents the stress amplitude and the horizontal axis represents the average stress. Here, S1 is a fatigue limit line at a low cycle (with a life), and S2 is a fatigue limit line at a high cycle (a permanent life). Further, σd is an average stress at the time of cycle evaluation of the maximum rotation of the turbine wheel 20 and the minimum rotation corresponding to the idling operation of the internal combustion engine, and σe is a stress amplitude at that time. Furthermore, σb is an average stress when the turbine blade 41 is damaged, and σa is a stress amplitude when the turbine blade 41 is damaged.

Here, “σd + σe” is σmax, and “σd≈σe”. As a result, “σmax = 2σe” is obtained. On the other hand, from experience, σa is approximately half of σb, and σa is approximately half of σe. For this reason, when “σe≈2σa” and “σa = σb / 2” are substituted into “σmax = 2σe”, σb at the time of turbine blade breakage is approximately half of σmax, and “σb <(1/2) × It was found that the damage of the turbine blade 41 can be suppressed if “σmax”. In other words, in order to suppress the breakage of the turbine blade 41, nozzle wake resonance may be generated when the centrifugal stress is smaller than half of the centrifugal stress at the maximum rotation of the turbine wheel 20. Here, since the centrifugal stress is obtained by F = mω ² r, the angular velocity (the number of rotations) of the turbine wheel 20 may be reduced in order to halve the centrifugal stress. That is, the desired angular velocity ω of the turbine wheel 20 is √ (1/2) × ωmax (maximum angular velocity of the turbine wheel 20), and √1 / 2≈0.7. Nozzle wake resonance may be generated when the maximum number of rotations per second is 70% or less.

  Next, with reference to FIG. 5, the relationship between the stress which the turbine blade 41 receives and the rotation speed of the turbine wheel 20 is demonstrated. In this graph, the vertical axis represents the stress received by the turbine blade 41, and the horizontal axis represents the rotational speed of the turbine wheel 20. Here, σb is a centrifugal stress, and σa is a vibration stress at the time of nozzle wake resonance. Σc is an allowable stress that the turbine blade 41 can tolerate, and σa + σb is a total stress applied to the turbine blade 41.

  As can be seen from the graph, when the rotational speed of the turbine wheel 20 increases, the centrifugal stress σb increases accordingly. On the other hand, as the centrifugal stress σb increases, the vibration stress σa increases as the excitation force of the nozzle wake decreases. Decrease. When the centrifugal stress σb exceeds a position where the vibration stress σa is about twice, the total stress σa + σb exceeds the allowable stress σc. At this time, the position where the total stress σa + σb and the allowable stress σc intersect is the rotational speed obtained by multiplying the maximum rotational speed of the turbine wheel 20 by 0.7, so the maximum rotational speed of the turbine wheel 20 is multiplied by 0.7. At the rotational speed (70%), it is possible to suppress damage to the turbine blades 41.

  Next, a method for calculating the resonance frequency of the nozzle wake resonance will be described with reference to FIG. FIG. 6 is a Campbell diagram of the turbine blade 41, where the vertical axis represents frequency and the horizontal axis represents the rotational speed of the turbine wheel 20. The maximum resonance frequency fmax at the maximum rotation of the turbine wheel 20 is the number of variable nozzle vanes 22 Zn × the maximum rotation speed Nmax of the turbine wheel 20 per second. The maximum resonance frequency fmax multiplied by a percentage of 70% or less is the natural frequency fa of the set turbine blade 41, and nozzle wake resonance occurs at this natural frequency fa. In addition, the natural frequency fb of the turbine blade in the conventional configuration is set to fa <fb <fmax.

  On the other hand, the minimum rigidity is set for each turbine blade 41 in consideration of a collision of a foreign matter mixed therein. For this reason, the minimum allowable stress in each turbine blade 41 is determined, and considering this allowable stress, it is not necessary to make it less than 50% of the maximum rotation speed Nmax of the turbine wheel 20. For this reason, in this embodiment, it is only necessary to cause nozzle wake resonance between 50% and 70% of the maximum rotation speed Nmax of the turbine wheel 20 (see FIG. 5).

  Here, the natural frequency of the turbine blade 41 can be set by changing the shape of the turbine blade 41, for example, the thickness. Thereby, since the natural frequency of the turbine blade 41 can be set by a simple method, it can be performed at low cost.

  According to the above configuration, the nozzle wake resonance is generated at 70% or less of the maximum rotational speed of the turbine wheel 20, so that the centrifugal stress applied to each turbine blade 41 is reduced by the amount that the rotational speed of the turbine wheel 20 is reduced. Can be halved. In this state, by generating nozzle wake resonance, vibration stress can be allowed as much as the centrifugal stress is reduced by half, so that the sum of centrifugal stress and vibration stress can be kept within the allowable stress. Thereby, it is possible to suppress damage to the turbine blades 41 due to nozzle wake resonance at a low cost without largely changing the apparatus configuration by simply setting the natural frequency of each turbine blade 41 by a simple method. Note that the natural frequency of the turbine blade 41 can be set by changing the material.

  As described above, the variable capacity turbine according to the present invention is useful for a variable capacity turbine having a variable nozzle vane, and is particularly suitable when nozzle wake resonance occurs.

It is sectional drawing of the turbocharger to which the variable capacity turbine which concerns on this embodiment is applied. It is a schematic diagram around a turbine wheel viewed from the axial direction. It is a graph regarding the rotation speed of the turbine wheel in the conventional structure, and the total stress which the turbine blade received. It is a fatigue limit line figure of a turbine blade. It is a graph regarding the rotation speed of the turbine wheel in this embodiment, and the total stress which the turbine blade received. It is a Campbell diagram of a turbine blade.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Turbocharger 2 Variable capacity turbine 20 Turbine wheel 22 Variable nozzle vane 41 Turbine blade σa Vibration stress σb Centrifugal stress σc Allowable stress

Claims (4)

A turbine wheel disposed in the gas inflow passage and rotatable by receiving exhaust gas in a plurality of turbine blades;
A plurality of variable nozzle vanes disposed on the turbine wheel upstream side of the gas inflow passage and capable of changing a flow rate of the exhaust gas flowing into the turbine wheel;
Each turbine blade is set with an allowable stress that allows a load due to its rigidity,
The sum of the centrifugal stress caused by the rotation of the turbine wheel and the vibration stress caused by the nozzle wake resonance caused by the rotation of the turbine wheel is less than the allowable stress, and the nozzle wake resonance caused by the rotation of the turbine wheel is caused by the turbine wheel. A variable-capacity turbine characterized in that the natural frequency of each turbine blade is set so as to occur at 70% or less of the maximum rotational speed.   The natural frequency of each turbine blade is set to a natural frequency corresponding to a vibration mode in which the inflow side end portion of each turbine blade into which the exhaust gas flows flows becomes a vibration antinode. The variable capacity turbine according to claim 1.   The natural frequency of each turbine blade is calculated by multiplying the resonance frequency calculated by multiplying the number of variable nozzle vanes by the maximum number of rotations per second of the turbine wheel by a percentage of 70% or less. The variable capacity turbine according to claim 1, wherein the variable capacity turbine is provided.   4. The natural frequency of each turbine blade is set so that the nozzle wake resonance occurs at 50% or more of the maximum rotational speed of the turbine wheel. The variable capacity turbine described.

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