A reciprocating engine requires a crankshaft for converting the reciprocating motion of pistons in cylinders to rotational motion so as to extract power. Crankshafts are generally categorized into two classes: the type manufactured by die forging and the type manufactured by casting. For multiple cylinder engines having two or more cylinders in particular, the firstly mentioned die forged crankshafts, which have higher strength and stiffness, are often employed.
FIG. 1 is a schematic side view of a common crankshaft for multiple cylinder engines. A crankshaft 1 shown in FIG. 1 is designed to be mounted in a 4 cylinder engine and includes: five journals J1 to J5; four crank pins P1 to P4; a front part Fr, a flange F1, and eight crank arms A1 to A8 (hereinafter also referred to as “crank arm”) that connect the journals J1 to J5 and the crank pins P1 to P4 to each other. The crankshaft 1 is configured such that all of the eight crank anus A1 to A8 are formed integrally with counterweights W1 to W8 (hereinafter also referred to as “counterweight”), respectively, and is referred to as a 4-cylinder 8-counterweight crankshaft.
Hereinafter, when the journals J1 to J5, the crank pins P1 to P4, the crank arms A1 to A8, and the counterweights W1 to W8 are each collectively referred to, the reference character “J” is used for the journals, “P” for the crank pins, “A” for the crank arms, and “W” for the counterweights. A crank pin P and a pair of crank arms A (including the counterweights W) which connect with the crank pin P are also collectively referred to as a “throw”.
The counterweight W may be included in each of the crank arms A, but instead may be included in at least one of the crank arms A. For example, in some of those crankshafts to be mounted in a 4 cylinder engine, of all the eight crank arms A, the leading first crank arm A1, the trailing eighth crank arm A8, and the central two crank arms (fourth crank arm A4 and fifth crank arm A5) are provided with the counterweight W. In such a case, the remaining second, third, sixth, and seventh crank arms A2, A3, A6, A7 do not include the counterweight W. Such crankshafts are referred to as a 4-cylinder 4-counterweight crankshaft.
The journals J, front part Fr, and flange F1 are arranged coaxially with the center of rotation of the crankshaft 1. The crank pins P are arranged at positions eccentric with respect to the center of rotation of the crankshaft 1 by half the distance of the piston stroke. The journals J are supported by the engine block by means of sliding bearings and serve as the central rotational axis. The big end of a connecting rod (hereinafter referred to as “connecting rod”) is coupled to the crank pin P by means of a sliding bearing, and a piston is coupled to the small end of the connecting rod. The front part Fr constitutes the front end portion of the crankshaft 1. A damper pulley 2 for driving a timing belt, a fan belt, and the like is attached to the front part Fr. The flange F1 constitutes the rear end portion of the crankshaft 1. A flywheel 3 is attached to the flange F1.
In an engine, fuel explodes within cylinders. The combustion pressure generated by the explosion causes the reciprocating motion of the pistons, which acts on the crank pins P of the crankshaft 1 and concurrently is transmitted to the journals J via each crank arm A connecting to a corresponding one of the crank pins P, so as to be converted into rotational motion. In this process, the crankshaft 1 rotates while repetitively undergoing elastic deformation.
The bearings that support the journals of the crankshaft are supplied with lubricating oil. In response to the inclination and the elastic deformation of the crankshaft, the oil film pressure and the oil film thickness in the bearings vary in correlation with the bearing load and the journal center orbit. Furthermore, depending on the surface roughness of the journals and the surface roughness of the bearing metal in the bearings, not only the variation of the oil film pressure but also local metal-to metal contact occurs. Ensuring a sufficient oil film thickness is important in order to prevent seizure of the bearings due to lack of lubrication and to prevent local metal-to-metal contact, thus affecting the fuel economy performance.
In addition, the elastic deformation caused by the rotation of the crankshaft and the movement of the center orbit in the clearances within the bearings cause an offset of the center of rotation, and therefore affect the engine vibration (mount vibration). Further more, the vibration propagates through the vehicle body and thus affects the vehicle interior noise and the ride comfort.
In order to improve such engine performance properties, there is a need for a crankshaft having high stiffness with the ability to resist deformation. In addition, there is a need for weight reduction of the crankshaft.
Crankshafts are subjected to loads due to pressure in cylinders (combustion pressure in cylinders) and centrifugal force of rotation. In order to impart deformation resistance to the loads, an attempt is made to improve the torsional rigidity and flexural rigidity. In designing a crankshaft, the main specifications such as the journal diameter, crank pin diameter, and crank pin stroke are firstly determined. After determination of the main specifications, the remaining region to be designed is the shape of the crank arm. Thus, the design of the crank arm shape for increasing both the torsional rigidity and the flexural rigidity is an important requirement.
In the meantime, crankshafts need to have a mass distribution that ensures static balance and dynamic balance so as to be able to rotate kinematically smoothly as a rotating body. Accordingly, an important requirement is to adjust the mass of the counterweight region with respect to the mass of the crank arm region determined by the requirements for the flexural rigidity and torsional rigidity in view of weight reduction while ensuring the static balance and dynamic balance.
For the static balance, the adjustment is made so that when the mass moments of inertia (the “mass” multiplied by the “radius of the center of mass”) of the crank arm region and the counterweight region are summed, the result is zero. For the dynamic balance, the adjustment is made so that, when, for each region, the product of the axial distance from the reference point to the center of mass multiplied by the mass moment of inertia (the “mass” multiplied by the “radius of the center of mass” multiplied by the “axial distance”) is determined using a point on the rotation axis of the crankshaft as the reference and the products are summed, the result is zero.
Furthermore, the balance ratio is adjusted for balancing against the load of combustion pressure within one throw (a region of the crankshaft corresponding to one cylinder). The balance ratio is defined as a ratio of the mass moment of inertia of the counterweight region to the mass moment of inertia of the crank arm region including the crank pin (also including part of the connecting rod, strictly speaking) in the crankshaft, and this balancing ratio is adjusted to fall within a certain range.
There is a trade-off between increase of the stiffness of the crank arm of a crankshaft and weight reduction thereof, but heretofore various techniques relating to the crank arm shape have been proposed in an attempt to meet both needs. Such conventional techniques include the following.
Japanese Patent No. 4998233 (Patent Literature 1) discloses a crank arm having intensively greatly depressed recess grooves in the crank pin-side surface of the crank arm and the journal-side surface thereof on a straight line connecting the axis of the journal to the axis of the crank pin (hereinafter also referred to as the “crank arm centerline”). The crank arm disclosed in Patent Literature 1 is intended to achieve weight reduction and increase of stiffness. The recess groove in the journal-side surface contributes to weight reduction by virtue of the reduced mass, and moreover, the thick region around the recess groove contributes to increasing the torsional rigidity. However, in reality, the flexural rigidity cannot be substantially increased because of the intensively greatly depressed recess grooves on the crank arm centerline.
Japanese Translation of PCT International Application Publication No. 2004-538429 (Patent Literature 2), Japanese Translation of PCT International Application Publication No. 2004-538430 (Patent Literature 3), Japanese Patent Application Publication No. 2012-7726 (Patent Literature 4), and Japanese Patent Application Publication No. 2010-230027 (Patent Literature 5) each disclose a crank arm having a greatly and deeply depressed hollow portion in the journal-side surface of the crank arm on the crank arm centerline. The crank arms disclosed in Patent Literatures 2 to 5 are also intended to achieve weight reduction and increase of torsional rigidity. However, in reality, the flexural rigidity is reduced because of the greatly and deeply depressed hollow portion on the crank arm centerline.