JP2014040856A - Crank shaft of multi-cylinder engine, and method for designing the same - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a crank shaft for a multi-cylinder engine, capable of reducing flexural vibration and torsional vibration and having a lighter weight by dynamic vibration evaluation.SOLUTION: A crank shaft of a multi-cylinder engine having two or more cylinders is mounted with a fly wheel on a shaft end. The flexural rigidity and the torsional stiffness of the crank arm part are higher at a position closer to the fly wheel, and are increased in three or more steps.

Description

  The present invention relates to a crankshaft mounted on a reciprocating engine such as a general-purpose engine such as an automobile engine, a marine engine, and a generator, and more particularly, a crankshaft mounted on a multi-cylinder engine having two or more cylinders, and the crank The present invention relates to a shaft design method.

  The reciprocating engine requires a crankshaft to extract power by converting the reciprocating motion of the piston into a rotational motion. Crankshafts are roughly classified into those manufactured by die forging and those manufactured by casting, and the former die-forged crankshaft, which is superior in strength and rigidity, is frequently used in multi-cylinder engines.

  FIG. 1 is a plan view schematically showing an example of a crankshaft for a general multi-cylinder engine. The crankshaft 1 illustrated in the figure is mounted on a four-cylinder engine, and includes five journal portions J1 to J5, four pin portions P1 to P4, a front portion Fr, a flange portion Fl, and journal portions J1 to J1. It consists of 8 crank arm parts (hereinafter also simply referred to as “arm parts”) A1 to A8 that connect J5 and the pin parts P1 to P4, respectively. This is a crankshaft of a 4-cylinder-8-counterweight having W1 to W8.

  Hereinafter, when the journal portions J1 to J5, the pin portions P1 to P4, the arm portions A1 to A8, and the weight portions W1 to W8 are collectively referred to, the reference numerals are “J” for the journal portion and “P” for the pin portion. ”,“ A ”for the arm portion and“ W ”for the weight portion. The pin portion P and the pair of arm portions A connected to the pin portion P are collectively referred to as “slow”.

  The journal portion J, the front portion Fr, and the flange portion Fl are arranged coaxially with the rotation center of the crankshaft 1, and the pin portion P is arranged eccentric from the rotation center of the crankshaft 1 by a distance that is half the piston stroke. . The journal portion J is supported by the engine block by a sliding bearing and serves as a rotation center shaft. A connecting rod is connected to the pin portion P by a sliding bearing, and a piston is connected to the connecting rod. The front portion Fr is one shaft end of the crankshaft 1, and a damper pulley 2 for driving a timing belt and a fan belt is attached to the front portion Fr. The flange portion Fl is the other shaft end of the crankshaft 1, and the flywheel 3 is attached to the flange portion Fl.

  As the fuel explodes (combusts) in each cylinder of the engine, the explosive force causes the piston to reciprocate and acts on the pin portion P of the crankshaft 1, and at the same time, the pin portion. It is transmitted to the journal part J via the arm part A connected to P and converted into a rotational motion. At that time, the crankshaft 1 rotates while repeating elastic deformation. Examples of the basic form of the deformation behavior are given below.

  FIG. 2 is a schematic diagram for explaining the deformation behavior of the crankshaft when the in-line four-cylinder engine crankshaft is approximated by a beam model. FIG. 2A shows an explosion in the first cylinder or the fourth cylinder. FIG. 5B shows the state when the explosion occurred in the third cylinder or the fourth cylinder. The figure shows the crankshaft deformation pattern when the explosion force F caused by the explosion generated in each cylinder is applied to the crankshaft, and the journals of the crankshaft. The state of distribution of the bearing load at the part is displayed (see Non-Patent Document 1). As shown in FIG. 2, when an explosion occurs in each cylinder, the explosion force F acts on the pin portion of the crankshaft in the cylinder, and at the same time, the arm portion via a pair of arm portions connected to the pin portion. This is mainly applied to the journal portion connected to, and at that time, it is distributed as a bearing load to all the journal portions in accordance with the elastic deformation of the crankshaft, particularly the elastic deformation of the arm portion. The crankshaft rotates while repeating such elastic deformation.

  By the way, there is lubricating oil in the bearing that supports the journal part of the crankshaft, and the oil film pressure and oil film thickness in the bearing depend on the inclination and elastic deformation of the crankshaft. It changes while correlating with the trajectory. Furthermore, depending on the surface roughness of the journal part in the bearing and the surface roughness of the bearing metal, not only the oil film pressure but also partial metal contact occurs. Ensuring the oil film thickness is important for preventing bearing seizure due to running out of oil and preventing partial metal contact, and affects fuel efficiency.

  In addition, the elastic deformation accompanying the rotation of the crankshaft and the shaft center trajectory that moves in the clearance in the bearing cause a shift of the center of rotation, which affects the engine vibration (mount vibration). It propagates through the car body and affects the noise and ride comfort in the passenger compartment.

  In order to improve such engine performance, the crankshaft is required to have high rigidity and be difficult to deform. Conventionally, loads of explosive force and rotational centrifugal force shown in FIG. 2 are applied to the crankshaft, and deformation resistance against these loads has been considered only by improving torsional rigidity and bending rigidity from a static viewpoint.

  For example, Patent Document 1 discloses a technique in which a hollow portion is provided in the central portion of the pin side surface and the journal side surface of the arm portion in order to increase the torsional rigidity and the bending rigidity while reducing the weight of the crankshaft. ing. The technique disclosed in this document focuses on weight reduction and high rigidity for one arm part, and shows a design method for each single arm part. It does not give them to the whole arm part. In other words, it is a design method that shows how to reduce the weight of a single arm when a rigidity value target in design is given, and conversely, the target value for weight reduction in design is It is a design method that shows how to increase the rigidity of a single arm part when given.

  Further, Patent Document 2 uses a three-moment method of material mechanics, approximates the crankshaft with a stepped round bar beam, and minimizes the load value applied to the journal part and the rigidity of the arm part and the arm. An optimization calculation method for obtaining an optimum distribution of the mass moment of the counterweight portion from the mass moment of the portion is shown. The technique disclosed in this document adopts an existing value for the rigidity of the arm part or determines it by another method, and then sets the mass moment of the counterweight part so that the load on the journal part is minimized. It is a method to adjust.

  In Patent Document 3, in order to reduce the vibration of the flywheel and improve the durability, in the crankshaft for a multi-cylinder engine, an arm portion disposed at a position closest to a flange portion to which the flywheel is mounted. A technique for making the wall thickness of the arm larger than that of the other arm portions is disclosed. The technology disclosed in this document focuses on the last thrown throw that is closest to the flywheel, and relates the thickness of the arm part of the last throw to the thickness of the arm part of the other throw. It is a precondition that the thickness of the arm portion of the other throw is constant.

  In other words, with regard to the thickness of the arm part, the last arm part closest to the flywheel is made larger than the other arm parts, and further, with respect to the width of the arm part, the arm part of the last throw is the other throw part. The size of the arm portion is larger than that of the arm portion, and the increase in rigidity when viewed in the total number of arm portions remains in two stages. Specifically, taking the crankshaft for an in-line 4-cylinder engine shown in FIG. 1 as an example, the first arm part A1 to the sixth arm part A6 have the same rigidity, and the seventh arm part of the last throw The rigidity of A7 is higher than that, and the rigidity of the last eighth arm part A8 is further increased.

JP 2009-133331 A Japanese Patent Laid-Open No. 10-169637 Japanese Utility Model Publication No. 5-1013

Yoshikazu Ishikawa, “Key points for designing gasoline engines for automobiles”, Sankaido, 2002

  However, even with any of the techniques disclosed in Patent Documents 1 to 3, the actual situation is that there is a limit to effectively suppressing bending vibration and torsional vibration in order to improve engine performance.

  The present invention has been made in view of such circumstances, and a crankshaft for a multi-cylinder engine capable of reducing weight while reducing both bending vibration and torsional vibration by dynamic vibration evaluation, Another object of the present invention is to provide a design method for the crankshaft.

  In order to solve the above-described problems, the present invention is summarized as a crankshaft of a multi-cylinder engine shown in (I) below and a crankshaft design method shown in (II).

(I) A crankshaft of a multi-cylinder engine having a flywheel attached to the shaft end,
The crankshaft of a multi-cylinder engine is characterized in that the bending rigidity and torsional rigidity of the crank arm part are higher as the crank arm part is closer to the flywheel, and the increase in bending rigidity and torsional rigidity is three or more.

  In the crankshaft, it is preferable that the bending rigidity and torsional rigidity of the crank arm portion are different for each crank arm portion.

(II) A method for designing a crankshaft of a multi-cylinder engine in which a flywheel is attached to the shaft end,
The crankshaft design method of the multi-cylinder engine is characterized in that the bending rigidity and torsional rigidity of the crank arm portion are increased as the crank arm portion is closer to the flywheel, and the bending rigidity and torsional rigidity are increased in three or more stages. is there.

  In this crankshaft design method, it is preferable that the bending rigidity and torsional rigidity of the crank arm portion are different for each crank arm portion.

  According to the present invention, it is possible to reduce the weight while reducing both bending vibration and torsional vibration.

It is a top view which shows typically an example of the general crankshaft for multicylinder engines. It is a schematic diagram explaining the deformation | transformation behavior of a crankshaft when the crankshaft for in-line 4 cylinder engines is approximated by the beam model, The figure (a) shows the state when explosion occurs in the 1st cylinder or the 4th cylinder. FIG. 4B shows a state when an explosion occurs in the third cylinder or the fourth cylinder, respectively. It is a figure which shows typically the representative example of the various design values of the arm part used for the balance design of a crankshaft, and a balance part. It is a schematic diagram for demonstrating the moment of inertia and the inertial product of a three-dimensional object. It is a figure which shows typically the representative example of the distributed mass used for consideration of the rotational motion of the whole crankshaft system containing a damper pulley and a flywheel. It is a schematic diagram which shows the model used by the conventional torsional vibration analysis. It is the figure which put together the outline | summary of the crankshaft of this invention. It is a figure which shows an example of the rigidity distribution of the arm part in the crankshaft of this invention. It is a figure which shows the out-of-plane deformation | transformation of a flywheel. It is a figure which shows the structure and out-of-plane deformation | transformation of a damper pulley. It is the figure which summarized the design point of the crankshaft of this invention. It is a figure which shows the bending rigidity distribution of the crankshaft in an Example. It is a figure which shows the torsional rigidity distribution of the crankshaft in an Example. It is a figure which shows the natural frequency of the bending of a crankshaft, and the natural frequency of torsion as a result of an Example. It is a figure which shows the weight of a crankshaft as a result of an Example. It is a figure which shows the moment of inertia around each axis | shaft of a crankshaft as a result of an Example. It is a figure which shows typically the bearing load of each journal part of a crankshaft as a result of an Example.

  Hereinafter, embodiments of the crankshaft of the multi-cylinder engine of the present invention and the design method of the crankshaft will be described in detail.

1. 1. Basic technologies to be considered in crankshaft design 1-1. Crankshaft balance Since the crankshaft is a rotating body, in addition to the design methods disclosed in Patent Documents 1 to 3, as a basic technology, a static balance and a dynamic balance are taken with respect to the rotating shaft, and the balance ratio is further increased. It is known that satisfying the above becomes an essential design condition for smoothing the rotational motion. Satisfying these conditions with the weight distribution of the counterweight portion is a precondition for designing the crankshaft.

(1) Static Balance As shown in FIG. 3, the crankshaft static balance will be described by taking an example of an eight counterweight crankshaft mounted on an in-line four-cylinder engine. In examining the static balance, the mass and the gravity center radius of each arm part and each weight part are defined as follows.
Arm mass: M _Aj ,
Arm center of gravity radius: R _Aj ,
Weight of weight part: M _Wj ,
Weight center of gravity radius: R _Wj .
Here, “j” is the number of the arm portion and the weight portion attached in order from the front portion side of the crankshaft, and j = 1,..., 8 in the case of a crankshaft for a four-cylinder engine.

  FIG. 3 representatively illustrates various design values of the third arm portion A3 and the third weight portion W3 that are third from the front portion side.

  In this case, taking a static balance means satisfying the following expression (1).

(2) Dynamic Balance Similarly, taking the crankshaft shown in FIG. 3 as an example, the dynamic balance of the crankshaft will be described. In examining the dynamic balance, the distance to the center of gravity of each arm part and the distance to the center of gravity of each weight part are defined as follows, with a certain point of the rotation axis of the crankshaft as a base point.
Distance to the center of gravity of the arm: L _Aj
Distance to weight center of gravity: L _Wj .
Here, “j” is j = 1,..., 8 in the case of the crankshaft for a four-cylinder engine, as in the above equation (1).

  In this case, dynamic balance means satisfying the following expression (2).

  The above formulas (1) and (2) are conventionally used static balance and dynamic balance design methods, and the sum is obtained for the center of gravity of each arm portion and each weight portion of the crankshaft. The crankshaft shape is adjusted and designed so that the sum is 0 (zero).

(3) Balance ratio The ratio of the mass moment on the weight portion side to the mass moment on the arm portion side including the pin portion of the crankshaft and a part of the connecting rod is called a balance rate. Since the balance ratio within a certain range contributes to the smooth rotation of the crankshaft, the weight of the weight portion needs to be added within a certain range not including 0 (zero). For this reason, there is a limit to reducing the weight and reducing the weight. Therefore, it is necessary to adjust the mass of the arm part in order to ensure rigidity, and in addition, it is necessary to adjust the mass of the weight part with a lower limit in order to ensure the balance ratio. The appropriate target mass is determined to some extent.

1-2. Theory of three-dimensional kinematics Next, the rotational motion of the entire crankshaft system will be considered from a wide theory of general three-dimensional kinematics. As a result, the basic shape can be grasped, and the requirements to be satisfied by the crankshaft as an ideal rotating body with less vibration fluctuation are clarified.

(1) Representation of rotational motion In the kinematics of a rotating body, when the position vector from the rotation center of the rotating body is r and the momentum vector is p, the angular momentum L is defined by the following equation (3).

  Here, “x” indicates an outer product. The momentum vector p is expressed by the following equation (4) using the mass m and the velocity vector v.

  Since the velocity vector v is a rotational motion, it is expressed by the following equation (5) using the angular velocity vector ω.

  Eventually, the following formula (6) is derived from the above formulas (3) to (5).

  When the display of the outer product of equation (6) is modified by the formula, the following equation (7) is obtained.

  If this equation (7) is written down with three-dimensional components (x, y, z), the following equation (8) is obtained.

  When this equation (8) is displayed in matrix, the following equation (9) is obtained.

  Here, I is called a moment of inertia tensor, is a symmetric matrix, and the diagonal component is a three-dimensional moment of inertia. Non-diagonal components mxy, mxz, and myz are called inertial products.

  The dynamic balance actually calculates this inertial product, and setting the dynamic balance to 0 (zero) means setting the inertial product to 0. That is, it means that the non-diagonal component of the matrix of the moment of inertia tensor I is set to 0 to form a diagonal matrix. In the three-dimensional rotating body, there is a direction called an inertial principal axis that is orthogonal, and when the X axis, the Y axis, and the Z axis coincide with the inertial principal axis, the matrix of the inertia moment tensor I is a diagonal matrix (other than diagonal) The component becomes 0). In other words, the operation of setting the dynamic balance to 0 can be said to be an operation of making the rotation axis coincide with the inertia main axis.

  Further, although the above equation (9) is displayed in a concentrated mass system, since it is possible to superimpose any number of mass points, the total moment of inertia I of a wide object is given by the following equation (10). Given.

  As shown in FIG. 4, even a distributed mass system model such as an FEM model of a three-dimensional object can be expressed by the above equation (10). Further, even when considering the mass point mass and the center of gravity position of a concentrated mass system composed of an assembly of a large number of members, it can be expressed by the above equation (10). The balance design of the crankshaft in section 1-1 described above is based on the above equation (10), considering the entire crankshaft system model as a mass system of concentrated mass composed of aggregates of each member, and extracting the X-axis system components. It corresponds to.

(2) About the ideally designed rotational motion of the crankshaft When the rotational shaft is made to coincide with the inertial main shaft, it is known that the rotating body rotates smoothly without causing blurring. Hereinafter, the rotational movement of the crankshaft will be considered using the above equation (10).

  For example, as shown in FIG. 5, the X axis is the rotation axis of the crank, and both the static balance and the dynamic balance regarding the X axis are set to zero. In addition, if the shape of the crankshaft on each axis is symmetric in the XY plane, the XZ plane, and the YZ plane, the design of the crankshaft is 0 for all mxy, mxz, and myz. In this case, the following equation (11) is established.

  Then, the following equation (12) is derived from the above equations (9) and (10).

The rotational angular velocity of the crankshaft is ω _x , and the angular momentum L _x is expressed by the following equation (13).

  And in the state without an ideal disturbance, the following (14) Formula is formed.

In this case, angular momentums L _y and L _z around the Y-axis and Z-axis become 0 as shown in the following equation (15). That is, there are no rotational components around the Y axis and Z axis.

  As described above, regarding the angular momentum, the states of the above formulas (13) to (15) represent an ideal crankshaft design, and in this case, smooth rotation around the X axis without vibration is realized.

  As another expression, the above equation (8) is time-differentiated. When the angular momentum L is differentiated with respect to time, the torque T is obtained. Moreover, since the angular acceleration (dω / dt) is obtained by differentiating the angular velocity ω with time, the motion equation of the rotational motion is expressed by the following equation (16).

  Substituting and writing the condition of the above expression (11) into the expression (16), the following expression (17) is obtained.

  In the state where there is no ideal disturbance as in the above equation (14), the following equation (18) is established.

  Then, the following equations (19) and (20) are derived from the above equations (17) and (18).

  In this case, the torque around the Y axis and the Z axis is not generated. Thus, with respect to torque, the states of the above formulas (18) to (20) represent an ideal crankshaft design, and in this case, smooth rotation around the X axis without vibration is realized.

(3) About the actual rotational movement of the crankshaft In the actual crankshaft design, the dynamic balance is not strictly zero, but there is a minute value. Further, minute displacements are generated by the elastic deformation of the crankshaft, and these minute displacements are minute inertia products [Delta] _xy , [Delta] _xz , [Delta] _yz that are not zero as shown by the following equation (21). Remains.

In addition to the rotational angular velocity ω _x around the X axis of the crankshaft, the presence of the oil film and clearance of the rotary motion bearing, the elastic vibration of the crankshaft, etc. There are minute angular velocity components δ _y and δ _z around the Z axis.

  Therefore, the above equation (12) representing the angular momentum is in a state represented by the following equation (23).

Similarly, the deviation from the ideal state is considered to have minute angular accelerations (dδ _y / dt) and (dδ _z / dt) with respect to the equation (16) representing the torque. ) Formula holds.

  From the condition of the equation (24), the following equation (25) is obtained.

  That is, in reality, the deviation from the ideal state is the cause, and the following (26) to (28) are derived from the above equation (23), and the angular momentum around the Y axis and the Z axis other than the rotation axis X axis is 0. do not become.

From the equations (27) and (28), it can be seen that the rotation angular velocity ω _{x having} a particularly large value is affected. That is, it can be seen that micro vibrations around the Y and Z axes are also caused by the rotational movement of the crank around the X axis.

  Similarly, the following (29) to (31) are derived from the above equation (25), and the torque around the Y axis and the Z axis does not become zero.

It can be seen from the equations (30) and (31) that the rotation angular acceleration (dω _x / dt) having a particularly large value is affected. These equations (26) to (31) cause the precession and vibration of the crankshaft. How small these are controlled is related to the quality of the crankshaft design and hence the engine design.

  Actually, the crankshaft bending elastic deformation or torsional elastic deformation, the bearing side engine block elastic deformation, the existence of the bearing oil film clearance and the crankshaft rotation inclination due to the oil film thickness distribution, etc. These causes overlap and a deviation from the ideal state of rotational motion occurs. As a result, a state as expressed by the above formulas (26) to (31) is obtained, and vibration noise is put on the circular rotation of the crankshaft to influence it. This hinders smooth rotation of the crankshaft and causes engine vibration and noise.

2. Conventional use of natural vibration analysis, which is a dynamic evaluation method Conventionally, natural vibration analysis is used to evaluate torsional vibration of the crankshaft. One example will be described below. As shown in FIG. 6, for each throw, the arm portion and the weight portion are replaced with a disk having an equivalent inertial mass for simplification, and the torsional rigidity between each throw is replaced with a spring constant cylinder for modeling. Then, the natural vibration analysis can be expressed by the following equation (32) (for example, see Non-Patent Document 1).

  By solving the equation (32), the natural frequency of torsion of the crankshaft system can be obtained. This natural frequency is used for order analysis of torsional vibration. This is mainly used for designing the damping effect of the torsional vibration of the damper pulley, and in particular for designing the rubber of the damper pulley and the design of the inertial mass in order to enhance the damping effect.

3. The present invention focuses on dynamic vibration of the elastic deformation of the crankshaft system, and uses the natural frequency as the rigidity and mass design of the crankshaft. As a result, it is used to improve engine performance.

  First, a general method for evaluating natural vibration will be described. The natural frequency is also called a natural vibration frequency or a resonance frequency. Considering the FEM model, for example, if the displacement vector is {δ}, the stiffness matrix representing stiffness is [K], and the mass matrix representing mass is [M], natural vibration occurs regardless of external force. Their relationship is as follows.

  The equation of motion is expressed by the following equation (33).

  If the displacement vibration is in the form of a cosine wave, the following equation (34) is established.

  Substituting this equation (34) into equation (33) leads to the following equation (35).

  In order to establish this equation (35), the determinant needs to be zero as shown in the following equation (36).

  From this equation (36), the natural frequency (f = ω / 2π) can be obtained. The corresponding δ in the above equation (34) gives the natural vibration mode.

  When the entire crankshaft system consisting of the crankshaft and flywheel, or the entire crankshaft system including the damper pulley is modeled with FEM and the above natural vibration analysis is performed, the natural vibration mode with the lower natural vibration frequency is revealed. Become. Thereby, the natural vibration frequency and natural vibration mode of bending of the crankshaft, and the torsional natural vibration frequency and natural vibration mode, respectively, are obtained.

  In the present invention, this is a design method that aims to increase the natural vibration frequency of bending and the natural vibration frequency of torsion. As a result, the bending rigidity and torsional rigidity of the crankshaft arm part monotonously increase from the front part side toward the flange part side, that is, toward the flywheel. The distribution of rigidity is the distribution in which the natural frequency is the highest when the mass is below a certain target mass.

  In the natural vibration analysis using the conventional equation (32) in Section 2 described above, the natural frequency of torsion appears, but the natural frequency of bending cannot be evaluated.

4). Crankshaft for multi-cylinder engine of the present invention 4-1. Outline The crankshaft of the present invention is intended for a crankshaft mounted on any reciprocating engine having two or more cylinders. That is, the number of cylinders of the engine may be any of 2, 3, 4, 6, 8, and 10 cylinders, or more. The arrangement of the engine cylinders is not particularly limited, such as an in-line arrangement, a V-type arrangement, or an opposed arrangement. The fuel of the engine is not limited to gasoline, diesel or biofuel. The engine also includes a hybrid engine formed by combining an internal combustion engine and an electric motor.

  FIG. 7 summarizes the outline of the crankshaft of the present invention. The crankshaft of the present invention uses the natural vibration analysis in Section 3 described above, and the crankshaft bending natural frequency and torsional eigenvalue of the entire crankshaft system including a flywheel or the entire crankshaft system including a damper pulley. The shape was designed in consideration of all the frequencies, and by examining each arm part instead of every throw, the bending rigidity and torsional rigidity of the arm part was higher in the crank arm part closer to the flywheel, and the bending rigidity and The increase in torsional rigidity is three or more. That is, in the crankshaft of the present invention, all the arm portions are numbered in ascending order in order from the front portion side to the flange portion side, and the bending rigidity and torsional rigidity of each arm portion are set to the arm portions. It increases monotonically in ascending order and increases to 3 or more levels. As long as the increase in flexural rigidity and torsional rigidity is three or more, the rigidity is allowed to be the same in some adjacent arm portions. It is shown from the Example mentioned later that the increase amount of the rigidity of each arm part shall be in the range of about 20% on the basis of the 1st arm part.

  FIG. 8 shows an example of the rigidity distribution of the arm portion in the crankshaft of the present invention. FIG. 8 illustrates a stiffness distribution pattern in the case of an in-line four-cylinder engine crankshaft. As shown in FIG. 8, the crankshaft of Conventional Example 1 has the same arm portion rigidity over all arm portions. Moreover, although the rigidity of the arm part of the crankshaft of the conventional example 2 (see Patent Document 3) monotonically increases in ascending order of the arm part number, the increase in rigidity is only two stages. On the other hand, in the crankshaft of Example 1 of the present invention, the rigidity of the arm part monotonously increases in ascending order of the arm part number, and differs for each arm part, and the increase in rigidity has seven stages. Further, the crankshaft of Example 2 of the present invention has a three-stage increase in rigidity.

In the entire crankshaft system, the flywheel directly connected to the shaft end of the crankshaft has a rotational inertia moment (mr ² : m is a distributed mass and r is a radius) that provides kinematic inertia in order to smoothly rotate the crankshaft. ) Is greatly given. This is because explosions occur intermittently in each cylinder of the engine and the rotation speed of the crankshaft is not constant, so that the rotation speed is kept as uniform as possible.

  However, the flywheel has a large radius in order to increase the rotational moment of inertia, but the plate thickness is relatively thin compared to that. This means that the out-of-plane deformation of the flywheel 3 increases as the crankshaft 1 rotates as shown in FIG. When the out-of-plane deformation of the flywheel 3 is large, a precession phenomenon, that is, a so-called swinging phenomenon occurs greatly during rotation. This also affects the deformation of the crankshaft 1, and the crankshaft 1 swings with the flywheel 3.

  The out-of-plane deformation of the flywheel is inseparable from the crankshaft deformation. That is, both the flywheel and the crankshaft influence each other, and if the deformation of the crankshaft is large, the out-of-plane deformation of the flywheel also increases. At this time, if the amplitude of vibration increases, the journal portion supported by the bearing swings around the shaft center locus, and the load increases, resulting in an increase in oil film pressure, a decrease in oil film thickness, Increased contact is provided. This situation is unpleasant and should be avoided.

  As in the present invention, it is designed by a method of increasing the frequency of natural vibration, which is a dynamic evaluation method, and if the shape is determined so that each rigidity of the arm portion monotonously increases in ascending order toward the flywheel, elastic deformation It becomes a crankshaft with excellent performance.

  In addition to the flywheel, the crankshaft with a damper pulley will be considered further. As shown in FIG. 10, the damper pulley 2 basically includes a disk portion 2a connected to the crankshaft 1, a ring-shaped vibration damping portion 2b such as rubber, and a ring-shaped inertia mass portion 2c corresponding to a weight. The damper pulley 2 has a function of absorbing torsional vibration and bending vibration at the shaft end of the crankshaft 1 and suppressing the vibration. The vibration damping effect is caused by torsional deformation or bending deformation of the vibration damping portion 2b. This induces a shift in the torsional position and a bending position in the inertial mass portion 2c that becomes the weight. The effects of these deformations on the rotation of the crankshaft 1 appear as torsional vibrations and out-of-plane deformation of the damper pulley 2.

  In this way, the damper pulley has a damping function and has the effect of suppressing vibrations when properly designed, but the crankshaft system as a whole includes the damper pulley and is designed to increase the natural frequency on it. It is preferable to incorporate.

The fact that the out-of-plane deformation at the time of rotation of the flywheel or the damper pulley is large is that the Δ _xy , Δ _xz , and Δ _{yz in} the above equations (21), (23), and (25) are not 0 (zero). Will have the value of This means that the rotation axis of the crankshaft is slightly displaced from the inertia main axis. Therefore, it is required to make Δ _xy , Δ _xz , and Δ _yz as small as possible. In this respect, increasing the resonance frequency of the natural vibration of bending and the natural vibration of torsion in the entire crankshaft system including the flywheel and the damper pulley is effective in reducing Δ _xy , Δ _xz , and Δ _yz .

4-2. Design Procedure FIG. 11 summarizes the design procedure of the crankshaft of the present invention. As described above, the distribution of rigidity of each arm portion of the crankshaft with an increased natural frequency obtained by the crankshaft design method of the present invention extends from the front portion side to the flange portion side in all arm portions. Ascending numbers are assigned in order, and then the bending rigidity and torsional rigidity of each arm part monotonically increase in ascending order of the arm part numbers, and increase to three or more levels. As long as the increase in flexural rigidity and torsional rigidity is three or more, the rigidity is allowed to be the same in some adjacent arm portions.

  Therefore, according to this design method, a crankshaft that satisfies the requirements shown in FIG. 7 can be designed. This design result can be interpreted as that the rigidity of the arm portion closer to the flywheel increases monotonically in order because the flywheel has a large mass moment and increasing the deformation rigidity increases the natural vibration frequency.

  At this time, in order to increase the natural frequency of bending and torsional natural frequency while maintaining the weight limitation of weight reduction, the rigidity distribution of each arm part of the crankshaft and the mass of each weight part It is necessary to determine the moment distribution. There are several ways to determine this, but the stiffness is determined by the optimization method of the “three-moment method” based on material mechanics and the optimization method of the natural frequency by the “force method” which is a non-parametric shape optimization method. Distribution and mass distribution can be determined. The specific selection of the optimization method is not specified here. Although the means is not specified, the design goal is to improve the natural frequency of bending and the natural frequency of torsion, and this is an exceptional design method that does not have a design policy for the rigidity of the arm portion.

The crankshaft for an in-line four-cylinder engine was taken as an example to confirm the effect of the present invention. Specifically, the numbers of 1, 2,..., 8 are assigned to the arm portions of the crankshaft in order from the front portion side toward the flange portion side, that is, toward the flywheel, The rigidity was as follows.
Bending rigidity of each arm part: Eb ₁ , Eb ₂ ,..., Eb ₈ ,
Torsional rigidity of the arm _{portions: Et 1, Et 2, ···} , Et 8.

  In the crankshaft of the example of the present invention, the rigidity of each arm portion satisfies the rigidity distribution conditions indicated by the following expressions (37) and (38).

  In the case of the crankshaft of the example of the present invention, as shown in FIG. 12, the bending rigidity of the arm portion is higher as the arm portion is closer to the flywheel, and it is not only monotonously increasing as it approaches the flywheel, The rigidity is different for each arm portion, and is increased in seven stages. Together with this, in the case of the crankshaft of the present invention example, as shown in FIG. 13, the torsional rigidity of the arm part is also higher as the arm part is closer to the flywheel and monotonously increases as it approaches the flywheel. Of course, the torsional rigidity is different for each arm portion, and is increased in seven stages. That is, the adjacent arm portions do not include the same rigidity.

  Thus, in the crankshaft of the present invention example, the rigidity distribution of the arm portion satisfies the expressions (37) and (38), and the bending rigidity and the torsional rigidity are increased in seven stages. The crankshaft satisfies the requirements shown in. Further, the crankshaft of the present invention example is designed such that the width and thickness of each arm portion are designed so that both the natural frequency of bending and the natural frequency of torsion increase in accordance with the procedure shown in FIG. .

  On the other hand, in the crankshaft of the conventional example, the rigidity of each arm portion satisfies the conditions of the rigidity distribution represented by the following equations (39) and (40).

  In the case of the crankshaft of the conventional example, as shown in FIGS. 12 and 13, both the bending rigidity and the torsional rigidity of the arm portion are the same throughout. That is, the adjacent arm portions have the same rigidity.

  14 to 17 show a comparison of various characteristics of the crankshaft of the present invention example and the conventional crankshaft.

  FIG. 14 is a diagram illustrating the natural frequency of bending of the crankshaft and the natural frequency of torsion as a result of the example. In the figure, each natural frequency of the crankshaft of the present invention is displayed in a ratio with each natural frequency of the crankshaft of the conventional example being 1. From the results shown in the figure, it is clear that both the bending natural frequency and the torsion natural frequency are higher in the crankshaft of the example of the present invention.

  FIG. 15 is a diagram showing the weight of the crankshaft as a result of the example. In the figure, the weight of the crankshaft of the present invention is indicated by a ratio where the weight of the crankshaft of the conventional example is 1. From the results shown in the figure, it is clear that the crankshaft of the example of the present invention is lighter and can be reduced in weight.

  FIG. 16 is a diagram showing the moment of inertia around each axis of the crankshaft as a result of the example. In the figure, each moment of inertia of the crankshaft according to the present invention is displayed at a ratio where each moment of inertia of the crankshaft according to the conventional example is 1. From the results shown in the figure, it has become clear that the crankshaft of the present invention is lighter and has a smaller moment of inertia, so that the effect of weight reduction can be expected from the viewpoint of rotational motion.

  FIG. 17 is a diagram schematically showing the bearing load of each journal portion of the crankshaft as a result of the example. The figure shows the result of evaluating the load by the three-moment method under the rotation of 1600 rpm. From the same figure, it became clear that the bearing load of the crankshaft of the example of the present invention is smaller particularly in the third journal portion at the center, and the bearing load is reduced. When the bearing load is reduced, there are many problems such as solving strength problems on the engine block side, reducing vibration, securing the oil film thickness of the bearing, improving fuel efficiency by reducing frictional force due to reduced partial contact at the bearing, reducing bearing wear, etc. This makes it an advantageous crankshaft.

  The present invention can be effectively used for a crankshaft mounted on any reciprocating engine having two or more cylinders.

1: Crankshaft, J1-J5: Journal part, P1-P4: Pin part,
Fr: front part, Fl: flange part, A1-A8: crank arm part,
W1-W8: Counterweight part,
2: damper pulley, 2a: disk part, 2b: vibration damping part,
2c: inertial mass part 3: flywheel

Claims (4)

A crankshaft of a multi-cylinder engine with a flywheel attached to the shaft end,
A crankshaft of a multi-cylinder engine, characterized in that the bending rigidity and torsional rigidity of the crank arm part are higher in the crank arm part closer to the flywheel, and the increase in bending rigidity and torsional rigidity is three or more.   2. The crankshaft of a multi-cylinder engine according to claim 1, wherein the bending rigidity and torsional rigidity of the crank arm portion are different for each crank arm portion. A crankshaft design method for a multi-cylinder engine with a flywheel attached to the shaft end,
A crankshaft design method for a multi-cylinder engine, characterized in that the bending rigidity and torsional rigidity of the crank arm part are increased as the crank arm part is closer to the flywheel, and the bending rigidity and torsional rigidity are increased in three or more stages.   2. The method of designing a crankshaft of a multi-cylinder engine according to claim 1, wherein the bending rigidity and torsional rigidity of the crank arm part are made different for each crank arm part.

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