JP2024051984A - Crankshaft - Google Patents

Abstract

To provide a crankshaft that can restrain vibration of a whole engine.SOLUTION: A crankshaft (10) comprises a plurality of journals (J), a plurality of pins (P), a plurality of arms (A), a front (Fr), and a flange (Fl). The plurality of journals (J) include a first journal (J1), a second journal (J2), and a third journal (J3) in order from the front (Fr) toward the flange (Fl). The plurality of pins (P) include a first pin (P1), and a second pin (P2) in order from the front (Fr) toward the flange (Fl). The plurality of arms (A) include a first arm (A1), a second arm (A2), a third arm (A3), and a fourth arm (A4) in order from the front (Fr) toward the flange (Fl). The lateral bending rigidity of a first throw (T1) is smaller than the lateral bending rigidity of a second throw (T2).SELECTED DRAWING: Figure 3

Description

This disclosure relates to crankshafts.

Traditionally, automobiles have used internal combustion engines as a power source. Although internal combustion engines generate power efficiently, they are prone to causing vibration and generating noise. Automobiles equipped with internal combustion engines are required to suppress vibration and noise in order to improve ride comfort and protect the environment. In particular, the maximum noise level is regulated by law. Therefore, suppressing noise in automobiles is not only a technical challenge, but is also important from the perspective of complying with the law.

In addition to automobiles, internal combustion engines are used in vehicles such as ships and heavy construction equipment, mobile machines such as push-type snow shovels, handheld machines such as chainsaws and lawnmowers, and stationary machines such as generators. Suppressing noise is also important in these machines.

One method of suppressing noise is to attach sound-absorbing materials (sound-absorbing or sound-proofing materials, etc.) to the cover surrounding the internal combustion engine. However, this method increases the weight of the entire machine due to the installation of sound-absorbing materials, and the entire machine also becomes larger due to the space required for installing the sound-absorbing materials. In particular, for automobiles, which are moving objects, increased weight leads to worse fuel efficiency. Another method of suppressing vibration is to install dampers between the machine frame and the internal combustion engine body and use the dampers to block vibrations. However, there are limits to how well dampers can block vibrations, and it is technically difficult to design dampers appropriately. Therefore, installing sound-absorbing materials to suppress noise and installing dampers to suppress vibrations is technically difficult.

In machines equipped with an internal combustion engine, vibrations generated by the engine are transmitted to the surface of the machine, and the vibrations transmitted to the machine cause the air to vibrate. This air vibration generates noise. This shows that if the vibrations generated by the internal combustion engine can be suppressed, the noise can also be suppressed.

Generally, the largest moving part in an internal combustion engine is the crankshaft. If the vibration of this crankshaft can be suppressed, it is expected that the vibration of the entire internal combustion engine will be suppressed. For these reasons, it is important to suppress the vibration of the crankshaft in order to suppress the vibration and noise generated by the internal combustion engine.

For example, Patent Document 1 discloses a crankshaft for an in-line four-cylinder engine that has eight crank shoulders (crank shoulders 1 to 8) in sequence from the accessory drive end. Each crank shoulder connects the main journal and the crank pin. Patent Document 1 states that by setting the stiffness ratio (ratio of bending stiffness to torsional stiffness) of the seventh crank shoulder to less than 1.8, it is possible to reduce crankshaft vibration and, ultimately, engine noise.

JP 2021-148281 A

The crankshaft vibrates as it rotates inside the engine. The vibration of the crankshaft can be broken down into vibration modes (hereafter simply referred to as modes). In other words, the sum of the vibrations of each mode is the overall vibration of the crankshaft.

In general, the amplitude of vibration of the crankshaft differs for each mode. Amplitude refers to the amount of energy in the vibration of each mode. The vibration of the crankshaft is transmitted to parts outside the crankshaft (such as the engine block). This vibration is measured as the vibration of the entire engine, or the vibration of the parts of the engine where the vibration is a problem (such as bearings or mounts). The effect of the vibration of each mode on the vibration of the entire engine is not uniform. In other words, the vibration of a specific mode contributes greatly to the vibration of the entire engine. Therefore, the vibration of the crankshaft can be suppressed by appropriately designing the crankshaft and reducing the amplitude of the vibration of the mode that contributes greatly to the vibration of the entire engine.

It is possible to suppress the vibration of the crankshaft by significantly improving the rigidity of each part of the crankshaft and suppressing the amplitude of vibration in all modes. However, in this case, the weight of the crankshaft increases significantly. Increasing the weight of the crankshaft is undesirable as it leads to a deterioration in fuel efficiency. Therefore, it is necessary to suppress the vibration of the crankshaft, which contributes greatly to the vibration of the entire engine, without significantly increasing the weight of the crankshaft.

The present disclosure aims to provide a crankshaft that can suppress vibrations throughout the engine.

The crankshaft for a four-cylinder engine of the present disclosure includes a plurality of journals, a plurality of pins, a plurality of arms, a front, and a flange. The plurality of pins are eccentrically disposed with respect to the plurality of journals. The plurality of arms connect the corresponding journals and pins, respectively. An engine accessory is attached to the front. A flywheel is attached to the flange. The plurality of journals include, in order from the front to the flange, a first journal, a second journal, and a third journal. The plurality of pins include, in order from the front to the flange, a first pin, and a second pin. The plurality of arms include, in order from the front to the flange, a first arm, a second arm, a third arm, and a fourth arm. The lateral bending stiffness of the first throw is smaller than the lateral bending stiffness of the second throw. The first throw is composed of the first journal, the first arm, the first pin, the second arm, and the second journal. The second throw is composed of the second journal, the third arm, the second pin, the fourth arm, and the third journal.

The crankshaft disclosed herein can suppress vibrations throughout the engine.

FIG. 1 is a front view showing an example of a typical crankshaft. FIG. 2 is a side view of an arm with a counterweight on the crankshaft shown in FIG. FIG. 3 is a front view of the crankshaft according to the embodiment. FIG. 4 is a schematic diagram showing analysis conditions for the vertical bending stiffness. FIG. 5 is a schematic diagram showing the analysis conditions for the lateral bending stiffness. FIG. 6 is a schematic diagram showing analysis conditions for torsional rigidity. FIG. 7 is a side view of the arm of the crankshaft. FIG. 8 is a front view of the crankshaft throw. FIG. 9 is a diagram showing the results of this example.

Figure 1 is a front view showing an example of a typical crankshaft 90. The crankshaft 90 is mounted on a four-cylinder engine. With reference to Figure 1, the crankshaft 90 includes five journals J1-J5, four pins P1-P4, eight arms A1-A8, eight counterweights W, a front Fr, and a flange Fl. Figure 2 is a side view of arm A with counterweight W in the crankshaft 90 shown in Figure 1. In Figure 2, the first arm A1 is shown as a representative of the arms A.

1 and 2, the crankshaft 90 rotates around the central axis X1. The first journal J1 to the fifth journal J5 each have a generally cylindrical shape with the central axis X1 as the center axis. The first journal J1 to the fifth journal J5 are arranged in this order from the front Fr to the flange Fl along the direction in which the central axis X1 extends, and constitute the main shaft portion of the crankshaft 90. The first journal J1 is connected to the front Fr. The fifth journal J5 is connected to the flange Fl. Ancillary equipment (not shown) is attached to the front Fr. The auxiliaries are, for example, pulleys for driving a timing belt and a fan belt. A flywheel (not shown) is attached to the flange Fl. Hereinafter, when there is no need to distinguish between the five journals J1 to J5, they will be collectively referred to as journals J.

The first pin P1 to the fourth pin P4 are each roughly cylindrical and are arranged alternately with the journal J along the central axis X1. The first pin P1 to the fourth pin P4 are arranged in this order from the front Fr toward the flange Fl. However, each of the first pin P1 to the fourth pin P4 is eccentric with respect to the journal J. The first pin P1 to the fourth pin P4 are arranged around the central axis X1 with a predetermined phase difference. The first pin P1 and the fourth pin P4, and the second pin P2 and the third pin P3 are arranged around the central axis X1 with a phase difference of 180°. Hereinafter, when there is no need to particularly distinguish between the four pins P1 to P4, they will be collectively referred to as pins P.

Each of the first arm A1 to the eighth arm A8 is disposed between the journal J and the pin P in the direction in which the central axis X1 extends. The first arm A1 to the eighth arm A8 each connect the journal J to the pin P. The first arm A1 to the eighth arm A8 are disposed in this order from the front Fr toward the flange Fl. Hereinafter, when there is no need to particularly distinguish between the eight arms A1 to A8, they will be collectively referred to as arm A.

The first arm A1 is disposed between the first journal J1 and the first pin P1, and connects the first journal J1 and the first pin P1. The second arm A2 is disposed between the first pin P1 and the second journal J2, and connects the first pin P1 and the second journal J2. The third arm A3 is disposed between the second journal J2 and the second pin P2, and connects the second journal J2 and the second pin P2. The fourth arm A4 is disposed between the second pin P2 and the third journal J3, and connects the second pin P2 and the third journal J3. The fifth arm A5 is disposed between the third journal J3 and the third pin P3, and connects the third journal J3 and the third pin P3. The sixth arm A6 is disposed between the third pin P3 and the fourth journal J4, and connects the third pin P3 and the fourth journal J4. The seventh arm A7 is disposed between the fourth journal J4 and the fourth pin P4, and connects the fourth journal J4 and the fourth pin P4. The eighth arm A8 is disposed between the fourth pin P4 and the fifth journal J5, and connects the fourth pin P4 and the fifth journal J5.

The first arm A1 to the eighth arm A8 are each provided with a counterweight W. Each of the counterweights W is formed integrally with the respective arm A1 to A8.

For ease of explanation, the direction in which the central axis X1 of the crankshaft 90 extends is also referred to as the axial direction. The direction in which the pin P is eccentric is also referred to as the vertical direction, and the direction perpendicular to both the axial direction and the vertical direction is also referred to as the width direction. Furthermore, in this specification, within the vertical direction, the direction in which the first pin P1 is eccentric with respect to the central axis X1 of the crankshaft 90 may be referred to as the up direction, and the direction in which the second pin P2 is eccentric may be referred to as the down direction.

The inventors decomposed the vibration of the crankshaft 90 into vibration modes and analyzed the effect of each mode on the vibration of the entire engine. Here, the vibrations decomposed into vibration modes are, for example, vertical bending vibration, lateral bending vibration, and torsional vibration. As a result of the analysis, it was found that, among the vibrations of each mode, the effects of lateral bending vibration and torsional vibration are relatively large. The lateral bending vibration of the crankshaft 90 means a vibration that displaces a part of the crankshaft 90 in the width direction, causing the crankshaft 90 to bend in the width direction. Moreover, the torsional vibration of the crankshaft 90 means a vibration that rotates a part of the crankshaft 90 around the axial direction, causing the crankshaft 90 to be twisted.

On the contrary, the analysis showed that the effect of vertical bending vibration is smaller than that of lateral bending vibration and torsional vibration. Vertical bending vibration in the crankshaft 90 refers to vibration that displaces a part of the crankshaft 90 in the vertical direction, causing the crankshaft 90 to bend in the vertical direction.

Therefore, if the amplitude of the lateral bending vibration and the torsional vibration in the crankshaft 90 is small, the vibration of the entire engine is suppressed. Furthermore, the inventors performed analysis by changing the rigidity of each part of the crankshaft in various ways, and examined which part of the crankshaft has an effect on the vibration of the entire engine. As a result, the inventors clarified that it is important to specify the ratio of the lateral bending rigidity of the first throw to the lateral bending rigidity of the second throw (lateral bending rigidity ratio) in particular in order to suppress vibration. Here, the throw is composed of one pin, arms connected to both sides of the pin, and journals connected to each of the arms. The crankshaft 90 has first to fourth throws in order from the front Fr to the flange Fl. The first throw is composed of the first journal J1, the first arm A1, the first pin P1, the second arm A2, and the second journal J2. The second throw is composed of the second journal J2, the third arm A3, the second pin P2, the fourth arm A4, and the third journal J3.

The reason why it is important to specify the lateral bending stiffness ratio of the first throw and the second throw in order to suppress vibration is because the lateral bending vibration of the first throw and the second throw contributes greatly to the vibration of the entire engine. A flywheel and a transmission (or a connecting shaft to the transmission) are connected to the flange Fl side of the crankshaft 90. The mass and inertia of these are very large compared to the crankshaft 90. Therefore, a strong restraining force acts on the flange Fl side of the crankshaft 90. Therefore, the amplitude of vibration is relatively small on the flange Fl side of the crankshaft 90. Conversely, the amplitude is larger on the front Fr side of the crankshaft 90, i.e., the first throw and the second throw, compared to the flange Fl side. It is presumed that the large amplitude vibration acting on the first throw and the second throw is transmitted to parts outside the crankshaft, such as the engine block, and increases the vibration of the entire engine.

The crankshaft according to the embodiment of the present disclosure was completed based on the above findings.

The crankshaft for a four-cylinder engine according to the embodiment includes a plurality of journals, a plurality of pins, a plurality of arms, a front, and a flange. The plurality of pins are arranged eccentrically with respect to the plurality of journals. The plurality of arms connect the corresponding journals and pins, respectively. An engine accessory is attached to the front. A flywheel is attached to the flange. The plurality of journals include, in order from the front to the flange, a first journal, a second journal, and a third journal. The plurality of pins include, in order from the front to the flange, a first pin, and a second pin. The plurality of arms include, in order from the front to the flange, a first arm, a second arm, a third arm, and a fourth arm. The lateral bending stiffness of the first throw is smaller than the lateral bending stiffness of the second throw. The first throw is composed of the first journal, the first arm, the first pin, the second arm, and the second journal. The second throw is composed of the second journal, the third arm, the second pin, the fourth arm, and the third journal (first configuration).

In the crankshaft of the first configuration, the lateral bending stiffness of the first throw is set smaller than the lateral bending stiffness of the second throw. In this way, by defining the magnitude relationship between the lateral bending stiffness of the first throw and the second throw in the crankshaft, the amplitude of the lateral bending vibration in the crankshaft can be reduced. As described above, the lateral bending vibration of the crankshaft has a large effect on the vibration of the entire engine. Therefore, the crankshaft of the first configuration can suppress the vibration of the entire engine.

In the crankshaft of the first configuration, the lateral bending stiffness of the first throw is preferably 0.95 to less than 1.00 times the lateral bending stiffness of the second throw (second configuration). In the crankshaft of the second configuration, the ratio of the lateral bending stiffness of the first throw to the lateral bending stiffness of the second throw (lateral bending stiffness ratio) is specifically set to 0.95 to less than 1.00. This ensures that vibrations throughout the engine are suppressed.

In the crankshaft of the first or second configuration, the torsional rigidity of the first throw is preferably greater than the torsional rigidity of the second throw (third configuration). As described above, not only the lateral bending vibration of the crankshaft but also the torsional vibration has a significant effect on the vibration of the entire engine. In the crankshaft of the third configuration, the magnitude relationship between the torsional rigidity of the first throw and the second throw is specified. This makes it possible to reduce the amplitude of the torsional vibration in the crankshaft, and to further suppress the vibration of the entire engine.

The crankshaft according to the embodiment of the present disclosure will be described below with reference to the drawings. The same or corresponding parts in the drawings will be given the same reference numerals, and duplicate descriptions will be omitted as appropriate.

[Crankshaft]
Fig. 3 is a front view (viewed along the width direction) of the crankshaft 10 according to this embodiment. The crankshaft 10 has substantially the same configuration as the crankshaft 90 shown in Fig. 1. Therefore, the following will describe the parts of the configuration of the crankshaft 10 that are different from the crankshaft 90.

The lateral bending rigidity of the first throw T1 is smaller than the lateral bending rigidity of the second throw T2. The lateral bending rigidity of the first throw T1 is preferably 0.95 times or more and less than 1.00 times the lateral bending rigidity of the second throw T2. In other words, the ratio of the lateral bending rigidity of the first throw to the lateral bending rigidity of the second throw T2 (lateral bending rigidity ratio) is preferably 0.95 times or more and less than 1.00. A method for calculating the lateral bending rigidity of each throw will be described later.

In this embodiment, the torsional rigidity of the first throw T1 is greater than the torsional rigidity of the second throw T2. In other words, the ratio of the torsional rigidity of the first throw to the torsional rigidity of the second throw T2 (torsional rigidity ratio) is greater than 1.00. A method for calculating the torsional rigidity of each throw will be described later.

As mentioned above, a strong restraining force acts on the flange Fl side of the crankshaft 10. Therefore, the rigidity of the portion of the crankshaft 10 on the flange Fl side, i.e., the third throw T3 and the fourth throw T4, does not have much effect on engine vibration. Therefore, the lateral bending rigidity and torsional rigidity of the third throw T3 and the fourth throw T4 are not particularly limited. Here, the third throw T3 is composed of the third journal J3, the fifth arm A5, the third pin P3, the sixth arm A6, and the fourth journal J4. The fourth throw T4 is composed of the fourth journal J4, the seventh arm A7, the fourth pin P4, the eighth arm A8, and the fifth journal J5.

In addition, among the vibration modes of the crankshaft 10, the influence of the vertical bending vibration on the vibration of the entire engine is smaller than that of the lateral bending vibration and torsional vibration. Therefore, the vertical bending rigidity of each throw is not particularly limited.

The crankshaft 10 has a first throw T1 to a fourth throw T4. Each of the first throw T1 to the fourth throw T4 has two arms A. These two arms A are typically symmetrical to each other in the axial direction. That is, the first arm A1 is symmetrical to the second arm A2, the third arm A3 is symmetrical to the fourth arm A4, the fifth arm A5 is symmetrical to the sixth arm A6, and the seventh arm A7 is symmetrical to the eighth arm A8 in the axial direction.

[How to calculate stiffness]
The following describes a method for calculating the vertical bending rigidity, lateral bending rigidity, and torsional rigidity of the crankshaft 10. To calculate the vertical bending rigidity, lateral bending rigidity, and torsional rigidity, an analytical model may be created for each throw of the crankshaft 10 and rigidity analysis may be performed. The following describes the case where the vertical bending rigidity, lateral bending rigidity, and torsional rigidity of the first throw T1 are calculated, but the vertical bending rigidity, lateral bending rigidity, and torsional rigidity of the second throw T2 to the fourth throw T4 can also be calculated in a similar manner.

Figure 4 is a schematic diagram showing the analysis conditions for the vertical bending stiffness. Figure 4 is a front view (viewed along the width direction) of the analysis model of the first throw T1 of the crankshaft 10. In the analysis of the vertical bending stiffness, a load F1 is applied downward in the vertical direction to the center in the axial direction of the first pin P1. The load F1 is applied to the upper end of the first pin P1. At that time, the axial center of the first journal J1 and the axial center of the second journal J2 of the crankshaft 10 are respectively constrained. The constrained positions of the first journal J1 and the second journal J2 are on the opposite side in the vertical direction from the side where the load F1 is applied to the first pin P1 (the lower ends of the first journal J1 and the second journal J2). As a result, the translation and rotation around the central axis X1 are constrained at the constrained points of the first journal J1 and the second journal J2. The load F1 applied to the first pin P1 causes the crankshaft 10 to bend in the vertical direction.

Under these conditions, the vertical displacement of point B of the first journal J1, point C of the first pin P1, and point D of the second journal J2 after the load F1 is applied is analyzed. Here, before the load F1 is applied to the first pin P1, point B is located at the center of gravity of the first journal J1, point C is located at the center of gravity of the first pin P1, and point D is located at the center of gravity of the second journal J2. Then, the difference (displacement difference Δ1) between the displacement of point C after the load F1 is applied and the average of the displacement of points B and D is calculated. The vertical bending rigidity is calculated by dividing the load F1 by the displacement difference Δ1.

Figure 5 is a schematic diagram showing the analysis conditions for lateral bending stiffness. Figure 5 is a top view (viewed along the vertical direction) of the analysis model of the first throw T1 of the crankshaft 10. In the analysis of lateral bending stiffness, a load F2 in the width direction is applied to the center in the axial direction of the first pin P1. At that time, the axial center of the first journal J1 and the axial center of the second journal J2 of the crankshaft 10 are each constrained. The constrained positions of the first journal J1 and the second journal J2 are on the opposite side in the width direction to the side where the load F2 is applied to the first pin P1. As a result, the translation and rotation around the central axis X1 are constrained at the constrained points of the first journal J1 and the second journal J2. The load F2 applied to the first pin P1 causes the crankshaft 10 to bend in the width direction.

Under these conditions, the displacement in the width direction of point B of the first journal J1, point C of the first pin P1, and point D of the second journal J2 after the load F2 is applied is analyzed. Then, the difference between the displacement of point C after the load F2 is applied and the average of the displacement of points B and D (displacement difference Δ2) is calculated. The lateral bending stiffness is calculated by dividing the load F2 by the displacement difference Δ2.

Figure 6 is a schematic diagram showing the analysis conditions for torsional rigidity. Figure 6 is a front view of an analytical model of the first throw T1 of the crankshaft 10. In the analysis of torsional rigidity, a torque T about the central axis X1 is applied to the second journal J2. At that time, the end of the first journal J1 on the front Fr (Figure 3) side in the axial direction is restrained. As a result, the end face of the first journal J1 on the front Fr side is completely restrained. The torque T applied to the second journal J2 causes the crankshaft 10 to be twisted about the central axis X1.

Under these conditions, the angle θ formed by the line connecting points E and F of the first journal J1 after torque T is applied and the line connecting points G and H of the second journal J2 when viewed along the axial direction is analyzed. Here, points E and F are each located at the center of the first journal J1 in the axial direction. Before torque T is applied to the second journal J2, point E is located at the upper end of the first journal J1, and point F is located at the lower end of the first journal J1. Points G and H are each located at the center of the second journal J2 in the axial direction. Before torque T is applied to the second journal J2, point G is located at the upper end of the second journal J2, and point H is located at the lower end of the second journal J2. The torsional rigidity is calculated by dividing torque T by angle θ.

[Crankshaft shape]
In the crankshaft 10 according to this embodiment, the lateral bending rigidity of the first throw T1 is smaller than the lateral bending rigidity of the second throw T2. Meanwhile, the torsional rigidity of the first throw T1 is greater than the torsional rigidity of the second throw T2. A specific shape of the crankshaft 10 that satisfies these conditions for lateral bending rigidity and torsional rigidity will be described below.

The lateral bending rigidity and torsional rigidity of the crankshaft 10 can be increased by increasing the width L and thickness t of the arm A. Here, the width L of the arm A refers to the dimension in the width direction of the upper end of the arm A. The thickness t of the arm A refers to the dimension in the axial direction of the upper end of the arm A. However, the lateral bending rigidity of the crankshaft 10 is more likely to be increased by increasing the thickness t of the arm A than by increasing the width L of the arm A. On the other hand, the torsional rigidity of the crankshaft 10 is more likely to be increased by increasing the width L of the arm A than by increasing the thickness t of the arm A.

Figure 7 is a side view of arm A of crankshaft 10. In Figure 7, a first arm A1 is shown as a representative of arm A. With reference to Figure 7, first arm A1 may include a thickened portion 11. The thickened portion 11 is provided to increase the width L of arm A. In short, arm A including thickened portion 11 has a larger width L compared to arm A not including thickened portion 11. Thickened portion 11 is provided on both sides of arm A in the width direction near pin P.

Figure 8 is a front view of the throw of the crankshaft 10. In Figure 8, the first throw T1 is shown as a representative throw. With reference to Figure 8, the first arm A1 and the second arm A2 of the first throw T1 may include a thickened portion 12. The thickened portion 12 is provided to increase the thickness t of the arm A. In other words, the arm A including the thickened portion 12 has a larger thickness t than the arm A not including the thickened portion 12. The thickened portion 12 is provided on the surface of the arm A near the pin P. Specifically, the thickened portion 12 is provided on the side of the arm A that is connected to the journal J that is connected to the arm A in the axial direction.

In the crankshaft 10, the first arm A1 and the second arm A2 of the first throw T1 may include the thickened portion 11, and the third arm A3 and the fourth arm A4 of the second throw T2 may not include the thickened portion 11. In this case, the width L of the first arm A1 and the second arm A2 is greater than the width L of the third arm A3 and the fourth arm A4. Therefore, the first throw T1 has a greater torsional rigidity than the second throw T2. Also, in the crankshaft 10, the first arm A1 and the second arm A2 may not include the thickened portion 12, and the third arm A3 and the fourth arm A4 may include the thickened portion 12. In this case, the thickness t of the first arm A1 and the second arm A2 is smaller than the thickness t of the third arm A3 and the fourth arm A4. Therefore, the lateral bending rigidity of the first throw T1 is smaller than that of the second throw T2. In other words, if the first throw T1 and the second throw T2 are shaped in this way, the above-mentioned conditions for lateral bending rigidity and torsional rigidity are satisfied.

[effect]
The lateral bending vibration of the crankshaft 10 has a large effect on vibration of the entire engine. Therefore, in the crankshaft 10 according to this embodiment, the lateral bending rigidity of the first throw T1 is set to be smaller than the lateral bending rigidity of the second throw T2. In this way, by defining the magnitude relationship between the lateral bending rigidities of the first throw T1 and the second throw T2 in the crankshaft 10, it is possible to reduce the amplitude of the lateral bending vibration in the crankshaft 10. Therefore, with the crankshaft 10 according to this embodiment, it is possible to suppress vibration of the entire engine.

In the crankshaft 10 according to this embodiment, the ratio of the lateral bending stiffness of the first throw T1 to the lateral bending stiffness of the second throw T2 (lateral bending stiffness ratio) is specifically set to 0.95 or more and less than 1.00. This ensures that vibrations throughout the engine can be suppressed.

Torsional vibration of the crankshaft 10, like lateral bending vibration, has a large effect on vibration of the entire engine. In the crankshaft 10 according to this embodiment, the magnitude relationship between the torsional stiffness of the first throw T1 and the second throw T2 is specified. This makes it possible to reduce the amplitude of the torsional vibration in the crankshaft 10, and to further suppress vibration of the entire engine.

To confirm the effect of the crankshaft according to the embodiment, the lateral bending rigidity and torsional rigidity were changed for each of the first and second throws of the crankshaft, and the magnitude of vibration (vibration level) of the entire engine was evaluated. In this example, a model was created that included the crankshaft, engine block, engine mount used to fix the engine to the automobile, flywheel and transmission attached to the flange of the crankshaft, etc., and the vibration of the entire model was investigated by numerical analysis. Specifically, the vibration level at the engine mount was investigated. If the vibration level at the engine mount is small, the vibration transmitted to the automobile will be small, improving the ride comfort of the automobile and reducing the noise emitted from the automobile. However, since it is only necessary to check vibrations of frequencies that humans can perceive, vibrations with frequencies of 1000 Hz or less were evaluated.

Figure 9 shows the results of this embodiment. In Figure 9, the vertical axis represents the vibration level (G) of the crankshaft, and the horizontal axis represents the ratio of the lateral bending rigidity of the first throw to the lateral bending rigidity of the second throw (lateral bending rigidity ratio). Referring to Figure 9, when the lateral bending rigidity ratio is 1.00 or more, the vibration level is greater than 8 (G). On the other hand, when the lateral bending rigidity ratio is less than 1.00, the vibration level is less than 7 (G). Among them, when the lateral bending rigidity ratio is less than 1.00 and the torsional rigidity ratio is greater than 1.00, the vibration level is particularly small. From the above, it can be seen that if the lateral bending rigidity ratio is less than 1.00, that is, if the lateral bending rigidity of the first throw is smaller than the lateral bending rigidity of the second throw, the vibration of the entire engine can be suppressed.

The above describes the embodiments of the present disclosure. However, the above-described embodiments are merely examples for implementing the present disclosure. Therefore, the present disclosure is not limited to the above-described embodiments, and can be implemented by modifying the above-described embodiments as appropriate within the scope of the spirit of the present disclosure.

10, 90: Crankshaft J: Journal P: Pin A: Arm Fr: Front Fl: Flange T1: 1st throw T2: 2nd throw

Claims (3)

1. A crankshaft for a four-cylinder engine, comprising:
Multiple journals and
A plurality of pins disposed eccentrically with respect to the plurality of journals;
a plurality of arms each connecting a corresponding one of the journals and the pin;
a front where the engine accessories are mounted;
a flange to which the flywheel is attached;
the plurality of journals include, in order from the front toward the flange, a first journal, a second journal, and a third journal,
The plurality of pins include, in order from the front toward the flange, a first pin and a second pin,
the plurality of arms include, in order from the front toward the flange, a first arm, a second arm, a third arm, and a fourth arm,
a lateral bending stiffness of a first throw consisting of the first journal, the first arm, the first pin, the second arm, and the second journal is smaller than a lateral bending stiffness of a second throw consisting of the second journal, the third arm, the second pin, the fourth arm, and the third journal. 2. The crankshaft according to claim 1,
The crankshaft, wherein the lateral bending stiffness of the first throw is greater than or equal to 0.95 and less than 1.00 times the lateral bending stiffness of the second throw. 3. The crankshaft according to claim 1 or 2,
The torsional stiffness of the first throw is greater than the torsional stiffness of the second throw.

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