Method for making a bearing for drive system of concrete mixer vehicle and drive system for concrete mixer vehicle and concrete mixer vehicle

Abstract

Method for making an output bearing for an output pinion for a drive for a concrete mixer vehicle, in which the output pinion is provided with a flange for putting in rotation a drum of the concrete mixer vehicle, the output bearing being orientable and comprising an internal ring and an external ring coaxially placed with interposition of first sliding rollers and second sliding rollers.

Description

[0001] DESCRIPTION

[0002] METHOD FOR MAKING A BEARING FOR DRIVE SYSTEM OF CONCRETE MIXER VEHICLE AND DRIVE SYSTEM FOR CONCRETE MIXER VEHICLE AND CONCRETE

[0003] MIXER VEHICLE

[0004] Technical Field

[0005] The present invention relates to a method for making a bearing for a drive system of a concrete mixer vehicle according to the characteristics of the pre-characterizing part of claim 1.

[0006] The present invention also relates to a drive system for concrete mixer vehicle incorporating such bearing and a concrete mixer vehicle equipped with such drive system according to the characteristics of the respective sub-claims.

[0007] Prior art

[0008] In the field of the construction of drives and in particular of drive systems for the rotation of a drum of a concrete mixer vehicle, it is known to combine a drive system with a drum hydraulic or electric motor to put in rotation the drum of the concrete mixer vehicle within which the continuous mixing occurs of the concrete contained in the drum itself. The rotation occurs in the two directions, one direction being suitable for mixing and one direction being suitable for discharging the concrete contained in the drum.

[0009] The drum usually is operated with a rotation speed between 0.1 rpm and 20 rpm, with a maximum rotation speed of about 14-20 rpm and a maximum torque ranging from 30000 Nm up to 60000 Nm and above, depending on the mixing capacity of the drum itself. The drum may be 3 to 4 m long, for example and without limitation for the application of the present invention, and it is fixed at its ends to a frame with a tractor for road transport of the concrete. A fundamental characteristic to be respected is that the rotation speed of the drum must be independent of the operating speed of the tractor motor. Since the tractor does not have a power take-off capable of supplying the speed and torque necessary for the drum to operate, the solution usually adopted consists of interposing a system consisting of a hydraulic pump and motor between the tractor motor and the drum. The motor of the axial piston drum is the most commonly used as it represents the best compromise between power output, size and costs. A drum motor of this type is capable of supplying maximum power at speeds between 1800 and 2500 rpm. Alternatively, as explained above, an electric motor can be used. Given the rotation speeds of the drum and of the drum motor, it is necessary to place a mechanical gear-reduction device between them that guarantees a minimum reduction ratio of approximately 1 :100 or even higher, as for example in the case of an electric motor, in which case a higher reduction ratio is likely to be used, even in the order of 1 :300, since the electric motor works better at high rpm. For this kind of application, an epicyclic multistage gear-reduction device is usually used for reasons of size and reliability. The gear-reduction device constituting the drive system is usually installed coaxially with the drum. The gear-reduction device constituting the drive system can be arranged partially inside the drum or, as in the case of the present invention, in an external position with respect to the internal compartment of the drum. There are also solutions with the drum motor arranged in axis with respect to the drive system as well as solutions with the drum motor arranged out of axis with respect to the drive system. The installation takes place on the opposite side of the drum with respect to the concrete introduction and extraction opening.

[0010] Technical solutions are known based on the use of a bell joint equipped with a grooved coupling between a series of teeth and grooves in which the teeth themselves can slide by rotation around a rotation center for the drive fixing flange. In this way, a misalignment is allowed between the final drive stage and the drive fixing flange on which the drum is fixed. The Patent Application JP 2004 100776 describes a bearing with improved rigidity in which a sliding surface of the rollers is made of a combination of arched parts with a radius such as to correspond to an external curvature radius of the roller and rectilinear parts, in which a contact area is created between the roller and the sliding surface formed by a combination of two semi-ellipsis and a rectangle, the rectilinear parts being tangent to the arched parts at both sides of the arched parts.

[0011] The Patent Application US 3 910 656 describes an orientable roller bearing for heavy loads comprising external and internal rings, each with a concave curvature of the rolling rail, and a plurality of convex rollers placed between the external and internal rings. The convex surface of the rollers, that in the central part of the roller presents a convex curvature radius substantially equal to the concave curvature radius of the rolling rail of the internal coupling ring, presents a modified curvature at each end for increasing orthogonal oscillation clearance between the roller and its rolling rails, in order to reduce the load on the edge of the roller and the consequent concentration of stresses at the ends of the roller. The provided oscillation clearance is correlated to the calculated elastic deformation of the roller and rolling rail in contact under the nominal maximum load. It is moreover described a method for the manufacturing of the roller.

[0012] The Patent Application US 3 990 753 describes a roller bearing comprising rollers with a positive inclination angle. The bearing includes external and internal rings having external and internal rolling rails which are spaced to define an annular space, a plurality of rolling rollers in the annular space between the rails. The rolling rollers are adapted to engage the internal and external rings in the contact zones of the external and internal rails. Actuating means are present for actuation during the operation of the bearing in order to cause at least the rolling rollers in the loaded region to assume a not-negative inclination angle within a predetermined angular range. The rolling rollers have an inclination angle when the axis of the rolling roller and the axis of the bearing are not coplanar and said inclination angle is positive when the components of the frictional force in the axial direction that are generated in each contact zone between a rolling roller and said rails and that act on the rolling roller are oriented in such a was as to sum with each other or in such a was as to be co-directional to the axial component of the load of the orthogonal contact force supported by said rolling roll in that contact zone. Furthermore a bearing is described with an inclination angle selected in such a way that the friction forces deriving from the inclined position of the rollers act to reduce the entity of the orthogonal contact forces between rollers and rolling rails.

[0013] The Patent Application DE 83 26 270 describes a gear-motor for a concrete mixer in which the transmission of the power occurs at least by means of a first stage composed by an internal sun gear, epicycloid gears and epicycloid support, the rotary movement of the epicycloid support being transmitted to an exit joint by means of a joint with curved teeth with a spherical bearing, in which the hub of the inwardly-directed protruding shaft of the exit joint presents a convex curvature radius adapted at the center of the pin of the spherical bearing and the supports are axially supported in a direction in correspondence of the convex curvature radius of the hub and in the other direction by means of rolling bearings.

[0014] Problems of prior art

[0015] The classic bearings used in concrete mixer vehicle drives do not optimally handle the high axial loads found in this application. The high ratio between axial and radial load causes non- optimal contact between rollers and rails. Indeed, with the experience gained over the years, it has been realized that with the classic configuration the life of the orientable bearing varied greatly between the various pieces even in the same production batch.

[0016] Furthermore, prior art solutions often require part of the drive to be housed within a space created in the drum of the concrete mixer vehicle, which leads to both problems related to drive maintenance and to the volume of the drum, which in order to guarantee the necessary capacity must be increased to take into account the volume occupied by the portion of the drive that is arranged within the volume of the drum itself.

[0017] Furthermore, it should be noted that the drum is mounted on roller supports that allow it to rotate and that support the weight of the drum and of the concrete contained in it. Part of the weight is also discharged onto the drive system. A further drawback is due to the fact that, during the movement of the concrete mixer vehicle on the road, the drum, the frame and the drive system are subject to strong stresses with the consequence that torsion phenomena are triggered of the frame of the concrete mixer vehicle itself. Consequently, there are relative motions between the rotating drum and the support frame on which the drive system and the drum motor are also fixed. To solve this problem, the solutions of prior art involve the use of a complex known with the name of bell joint, which allows the drive of the rotational motion and at the same time it allows the reciprocal movement of the drum and the part of the drive system that is fixed to the frame. The bell joint must also be able to support and discharge at least a part of the weight of the drum and of the concrete contained in it.

[0018] Some solutions of the prior art involve the use of a drive system that is coupled to the vehicle frame by means of elastic means, such as for example high thickness rubber pads. Through the elastic deformation of these support pads, reciprocal movements of the drum with respect to the vehicle frame are allowed. These systems, although they appear simple, present problems with regard to the limitation of the movements allowed by the elastic means to prevent excessive movement from leading to excessive stress on the components with possible detachments of the same. Furthermore, these elastic means may be subject to premature aging with consequent loss of their functionality, given their exposure to temperature changes, sun, weather, etc. The need to provide for sizing of the pads suitable for the application and sizing of the mechanical components to withstand the induced vibrations, furthermore, compromises at least in part the economic advantages obtained from the adoption of these systems, so much so as to make the solution more expensive than traditional systems, in addition to the fact that it does not provide for any reduction in the weight and overall dimensions of the group composed of the drive and drum motor.

[0019] The Patent Application JP 2004 100776 does not contain any consideration about the radial clearance between the roller surface and the external ring surface or about the radial clearance between the roller surface and the semi-surfaces of the internal ring which, as explained in the description is an essential parameter to obtain the results that form the objective of the present patent application.

[0020] The Patent Application US 3 910 656 describes an orientable roller bearing that presents a modified curvature at each end of the rollers for increasing orthogonal oscillation clearance between the roller and its rolling rails, in order to reduce the load on the edge of the roller and the consequent concentration of stresses at the ends of the roller. The provided oscillation clearance is correlated to the calculated elastic deformation of the roller and rolling rail in contact under the nominal maximum load. In other words, the Patent Application US 3 910 656 suggests to adopt a configuration in which is increased the clearance reducing the contact surface for reducing the load on the edge of the rollers, i. e. is suggested a solution that is contrary to what is instead claimed in the present patent application.

[0021] The Patent Application US 3990753 provides the presence of a recess on the external guide, stating that an external guide is provided which has a zone having a lower curvature radius, which is in conflict with claimed subject matter in the present patent application. Furthermore, the Patent Application US 3990753 does not contain any consideration about the radial clearance between the roller surface and the external ring surface or about the radial clearance between the roller surface and the semi-surfaces of the internal ring which, as explained in the description is an essential parameter to obtain the results that form the objective of the present patent application.

[0022] Aim of the invention

[0023] The main aim of the invention is to improve the life of the bearing supporting the drum that undergoes the stresses of the drum in the case of a gear-reduction device external to the drum.

[0024] Aim of the present invention is to provide a bearing for a concrete mixer vehicle drum drive which allows for effective discharging of the weight of the concrete mixer vehicle and at the same time ensures a high degree of alignment between the drive members and the concrete mixer vehicle drum even in conditions where the drum is in an inclined condition.

[0025] The aim of the present invention is also to solve the aforementioned problems and disadvantages of the prior art, providing a compact and reliable drive for a concrete mixer vehicle drum that allows a reduction in overall dimensions and that has a longer life in relation to the maintenance interventions foreseen, especially with reference to the verification and replacement of the drum support bearing.

[0026] A further aim of the present invention is to allow an effective discharge of the portion of the weight of the concrete mixer vehicle drum that weighs on the drive while ensuring a reduction of the drive itself while maintaining its reliability unchanged.

[0027] Concept of the invention

[0028] The aim is achieved with the characteristics of the main claim. The sub-claims represent advantageous solutions. Advantageous effects of the invention

[0029] The solution according to the present invention, through the considerable creative contribution the effect of which constitutes an immediate and not negligible technical progress, presents several advantages.

[0030] Specific advantages obtained by means of the new solution concern an increase in the average useful life of the bearing which is advantageously increased by 80%.

[0031] It allows to obtain a greater compactness of the drive for concrete mixer vehicle with the additional advantage of eliminating the need to have a portion of drive housed within a space obtained within the volume of the drum of the concrete mixer vehicle. This aspect also simplifies the maintenance operations of the drive also in relation to the possible inspection and replacement of the drum support bearing.

[0032] Description of the drawings

[0033] An embodiment solution is described below with reference to the attached drawings to be considered as a non-limiting example of the present invention in which:

[0034] Fig. 1 schematically represents a concrete mixer vehicle comprising the drive made according to the present invention.

[0035] Fig. 2 schematically represents the drive for concrete mixer vehicle made according to the present invention.

[0036] Fig. 3 is a schematic exploded view of the drive for concrete mixer vehicle made according to the present invention.

[0037] Fig. 4 is a sectional view of the drive for concrete mixer vehicle made according to the present invention.

[0038] Fig. 5 schematically represents a sectional detail of a portion of the drive for concrete mixer vehicle made according to the present invention.

[0039] Fig. 6 is a perspective view of the concrete mixer vehicle drive bearing made according to the present invention.

[0040] Fig. 7 is a sectional view of a first embodiment of the concrete mixer vehicle drive bearing made according to the present invention. Fig. 8 is a sectional view of a second embodiment of the concrete mixer vehicle drive bearing made according to the present invention.

[0041] Fig. 9 is a sectional view of one of the components of the bearing of Fig. 7.

[0042] Fig. 10 is a sectional view of another of the components of the bearing of Fig. 7.

[0043] Fig. 11 is a sectional view of one of the rollers of the bearing of Fig. 7.

[0044] Fig. 12 schematically illustrates some parameters of the concrete mixer vehicle drive bearing made according to the present invention.

[0045] Fig. 13 is a sectional view of the concrete mixer vehicle drive bearing made according to the present invention illustrating an inclined condition of the bearing components.

[0046] Fig. 14 schematically represents, in a sectional view, the reciprocal movement permitted between the components of the drive made according to the present invention.

[0047] Fig. 15 represents a chart showing the test results that highlight the considerable increase of useful life of the inventive bearing with respect to the bearings of prior art in use previously to the introduction of the present innovation.

[0048] Definitions

[0049] In the present description and in the annexed claims the following terms must be intended according to the here provided definitions.

[0050] With the term “contact degree” it is indicated a parameter known to an expert in the art of the bearings which is calculated as the ratio Rr / Ra between the roller radius and the radius of the respective internal and external rolling rail in the conditions of normal use of the bearing with the maximum provided load as nominal load. For the purpose of this calculation the internal and external rolling rails radiuses can be considered as having the same radius except for construction tolerances used to obtain the advantages of the inventive solution which will be explained in the following of the present description, assuming, therefore that Rae = Rai = Rail = Rai2. The more this ratio comes closer to 1 , the greater becomes the useful contact surface between the roller and the respective rolling rail and, as a result, the greater becomes the load capacity of the bearing or, with the same load acting on the bearing, the greater becomes the useful life of the same. With reference to the term “clearance” of the bearing, it indicates the possibility of a relative movement of the external and internal rings, between each other, in a condition of the bearing non mounted on the respective shaft. When one between the internal rings and the external ring is constrained, for example by means of fixing the internal ring on the shaft or fixing of the external ring on the respective seat, the other ring, i. e. the one which is not fixed is still free to move thanks to the clearance which is present and it can move both in axial and radial direction, from which the definitions of the terms “radial clearance” and “axial clearance”.

[0051] With the term “radial clearance” it is indicated a known parameter to an expert in the art of the bearings which is an intrinsic constructive parameter of all the bearings and defines the radial movement admitted between rails and rolling bodies. In practice, the radial clearance is the measured value of the total radial movement of the external ring with respect to the internal ring on a plane perpendicular to the axis of the bearing. The radial clearance is determined by the real diameters of the rails and by the diameter of the rolling bodies, i. e. in case of the bearing of the present application, by the diameter of the rollers. In practice, with an equal diameter of the rails, the less the diameter of the rollers is the greater the radial clearance will be and vice-versa. The declared radial clearance as given as a nominal parameter of a bearing has to be considered with the bearing in a free and unloaded conditioning, without angular variation, or before being assembled, therefore without load. For the purpose of this invention the parameter of the radial clearance is calculated in the condition in which the deviation angle, i. e. the inclination angle between the external ring of the bearing and the internal ring of the bearing is equal to 0 degrees and the bearing is in a condition of absence of load.

[0052] With the term “axial clearance” it is indicated a known parameter to an expert in the art of the bearings which is an intrinsic constructive parameter of all the bearings and defines the axial movement admitted between rails and rolling bodies. In practice, the axial clearance is the measured value of the total axial movement of the external ring with respect to the internal ring parallel to the axis of the bearing. The axial clearance is determined by the real diameters of the rails and by the diameter of the rolling bodies, i. e. in case of the bearing of the present application, by the diameter of the rollers. In practice, with an equal diameter of the rails, the less the diameter of the rollers is the greater the axial clearance will be and vice-versa. The declared axial clearance as given as a nominal parameter of a bearing has to be considered with the bearing in a free and unloaded conditioning, without angular variation, or before being assembled, therefore without load. For the purpose of this invention the parameter of the axial clearance is calculated in the condition in which the deviation angle, i. e. the inclination angle between the external ring of the bearing and the internal ring of the bearing is equal to 0 degrees and the bearing is in a condition of absence of load.

[0053] With the term “radial load” it is indicated a known parameter to an expert in the art of the bearings which is an intrinsic constructive parameter of all the bearings and defines the design radial load for the bearing expressed in newton, the radial load being the force acting on the bearing in the radial direction, or perpendicularly to the rotation axis of the bearing in the working conditions.

[0054] With the term “axial load” it is indicated a known parameter to an expert in the art of the bearings which is an intrinsic constructive parameter of all the bearings and defines the design axial load for the bearing expressed in newton, the axial load being the force acting on the bearing in the axial direction, or parallel to the rotation axis of the bearing in the working conditions.

[0055] Description of the invention

[0056] With reference to the figures (Fig. 1), the concrete mixer vehicle (8) consists of a tractor (6) driven by a respective motor of the tractor (4) and a frame (7) on which a rotating drum (3) resting on supports (5) is applied. The rotation of the drum (3) occurs (Fig. 2, Fig. 3) by means of a drive system (1) driven by a respective drum motor (2), which puts in rotation the drum (3) of the concrete mixer vehicle (8) within which the continuous mixing occurs of the concrete contained in the drum (3) itself. The drum motor (2) can be of the hydraulic type or of the electric type. In the case of a drum motor (2) of the hydraulic type, it is preferably, but not necessarily, an axial piston motor.

[0057] The rotation occurs in the two directions, one direction being suitable for the mixing and one direction being suitable for the discharging of the concrete contained in the drum. Since the drum (3) must be operated with a rotation speed included approximately but not limitedly in the range of 0.1 rpm to 20 rpm, in general with maximum rotation speed values between 14 rpm and 20 rpm and a maximum torque ranging from 30000 Nm up to 60000 Nm and since the drum motor (2) is able to provide the maximum power at speeds ranging from 1800 rpm to 2500 rpm, it is necessary to interpose a mechanical gear-reduction device between them that guarantees a minimum reduction ratio of approximately 1 :100. The solution according to the present invention provides for the interposition of a drive system (1 ) consisting of an epicyclic multistage gear-reduction device. The gear-reduction device constituting the drive system (1 ) is preferably installed coaxially to the drum (3) at the opposite side of the drum (3) with respect to the concrete introduction and extraction opening.

[0058] The drive system (1 ) is made (Fig. 3, Fig. 4) according to a configuration with planetary- gears and planetary-gear supports. Although in the illustrated embodiments the configuration includes 2 stages of planetary-gears and planetary-gear supports, the present invention must be understood as applicable also to solutions in which there are a greater number of stages, such as drive with a configuration of 3 stages of planetary-gears and planetary-gear supports.

[0059] In particular, in the illustrated embodiment, the configuration provides (Fig. 4, Fig. 5) that the last stage of the drive system (1), in the case shown the second stage, is a stage configured with a flange (23) which is a single body together with the output pinion (15) of the drive system (1) in which the output pinion (15) is connected to the last stage of the drive system (1) with the interposition of a bell (21) suitable for allowing the drive of the rotational motion to the flange (23) and, therefore, to the drum (3), even in the conditions in which the previously described relative motions occur between the rotating drum (3) and the support frame (7) on which the drive system (1) and the respective drum motor (2) are also fixed. The bell (21) couples with the planetary-gear support of the second stage (14), as will be explained later.

[0060] The output pinion (15) comprises (Fig. 5) a first portion (16) for coupling with an output bearing (26), which is a bearing of orientable type, and a second portion (17) for coupling with a grooved coupling (38). The drive of the rotary motion to the drum (3) is obtained by means of a grooved coupling (38) which comprises grooves (37) and teeth (36) reciprocally coupled to allow an inclination (Fig. 14) according to an inclination angle (30) of a few degrees between the axis of the drive (22) and the axis of the drum (33) of the concrete mixer vehicle (8). The grooved coupling (38) preferably comprises:

[0061] - a series of grooves (37) arranged radially externally on the second portion (17) of the output pinion (15) for transmitting the rotational motion from the last stage of the drive system (1 ), to the flange (23) for fixing to the drum (3);

[0062] - a series of teeth (36) arranged radially internally on the bell (21).

[0063] Each tooth (36) of the series of teeth (36) slides within a respective groove (37) of the series of grooves (37) in such a way that it guides the reciprocal movement of inclination (Fig. 14) between the output pinion (15), which is integral with the flange (23), and the bell (21 ) which constitutes the element that transmits the rotational motion from the last stage of the drive system (1 ).

[0064] The maximum inclination (Fig. 14) allowed is equal to the inclination angle (30) between the drive axis (22) and the drum axis (33), the inclination angle (30) being at least equal to + / -3 degrees, even more preferably being at least equal to +1-6 degrees. This allows to take into account the bendings of the frame (7) of the concrete mixer vehicle (8) that occur during road transport, avoiding that the stresses induced on the drive system (1) may lead to its blocking or to its premature wear, as well as avoiding that anomalous stresses may occur on the supports of the rotating drum (3).

[0065] Advantageously, the solution according to the present invention provides that no component of the drive system (1 ) is housed within a chamber obtained at the head of the drum (3), i.e. at the side of the drum (3) which is the opposite side with respect to the side on which the opening for the introduction and withdrawal of the concrete is present. This is possible thanks to the compact shape of the drive obtained thanks to the described configuration.

[0066] Although reference is made in the figures and in the description to a two-stage epicyclic gear-reduction device, it will be clear that solutions with three reduction stages can also be envisaged without departing from the scope of the present invention, the obvious extension to the case of an epicyclic gear-reduction device with a different number of stages being evident to an expert in the sector. For example, in the case of a drive system (1) for a concrete mixer vehicle with a two-stage epicyclic gear-reduction device, the box (19) of the drive system (1) houses inside it only (Fig. 3, Fig. 4) both stages of the gear-reduction device:

[0067] - the first stage with the respective first-stage planetary-gear supports (11), first-stage planetary-gears (10) and first-stage pinion (9);

[0068] - the second stage with the respective second-stage planetary-gear supports (14), second- stage planetary-gears (13) and second-stage pinion (12).

[0069] The second stage planetary-gear support (14) couples with the bell (21) by means of a radial coupling gear (48) so that the bell (21 ) rotates together and at the same speed as the second-stage planetary-gear support (14).

[0070] The flange (23) for fixing or coupling with the drum (3) of the concrete mixer vehicle (8) includes the output pinion (15) which couples with the second stage as previously described. The fixing flange (23) is blocked in the assembly position by means of an annular fixing element (20).

[0071] The drive system (1 ) is fixed on the frame (7) of the concrete mixer vehicle (8) by means of (Fig. 1 , Fig. 2) fixing means (25) obtained in correspondence with the base of the box (19) of the drive system (1) itself.

[0072] The drum motor (2) is fixed on the box (19) of the drive system (1) at a cover (18) that closes the inside of the box (19) containing the first stage (9, 10, 11) and the second stage (12, 13, 14) of the drive system (1 ). The cover (18) is located on an opposite side of the drive system (1) with respect to the side of the drive system (1) at which the flange (23) is present. The cover (18) includes a driving input (24) for the passage of a shaft (43) of the drum motor (2) in such a way that the shaft (43) of the drum motor (2) is coupled to the first stage (9, 10, 11) of the drive system (1 ).

[0073] In particular, the first stage (9, 10, 11) of the drive system (1) comprises a pinion of the first stage (9) which is supported on the cover (18) by means of an input bearing (27). The pinion of the first stage (9) comprises on its head a toothed recess (44) into which the shaft (43) of the drum motor (2) is inserted to operate the drive system (1). In this case the shaft (43) of the drum motor (2) is machined to obtain a counter-toothing for coupling with the toothed recess (44) of the pinion of the first stage (9).

[0074] The pinion (9) of the first stage couples (Fig. 3, Fig. 4) with corresponding planetary-gears (10) of the first stage, which in turn transmit motion to the planetary-gear support (11) of the first stage. The planetary-gears (10) of the first stage are supported on the planetary-gear support (11 ) of the first stage by means of bearings of the first stage (28). The planetary-gear support (11) of the first stage in turn transmits motion to the pinion (12) of the second stage. The pinion (12) of the second stage couples (Fig. 3, Fig. 4) with corresponding planetary- gears (13) of the second stage which in turn transmit motion to the planetary-gear support (14) of the second stage. The pinion (12) of the second stage is rotatably supported by the planetary-gears of first and second stage and by both planetary-gear supports. The planetary-gears (13) of the second stage are supported on the planetary-gear support (14) of the second stage by means of bearings of the second stage (29). The planetary-gear support (14) of the second stage in turn transmits motion to the bell (21) which is coupled, as previously described, with the second portion (17) of the output pinion (15) to allow angular inclination according to the inclination angle (30) between the output pinion (15) and the last stage of the drive system (1), in this case, the second stage (12, 13, 14) of the drive (1).

[0075] In particular, angular inclination is allowed according to the inclination angle (30) between the last stage of the drive system (1) and the assembly composed of output pinion (15) and bell (21). Indeed, the bell (21) follows the output pinion (15) and, therefore, varies its inclination, the radial coupling gear (48) being made in such a way as to ensure rotation between the last stage of the drive system (1 ) and the assembly composed of output pinion (15) and bell (21).

[0076] In order to allow the described angular inclination between the output pinion (15) and the last stage of the drive system (1), the box (19) comprises (Fig. 4, Fig. 5, Fig. 14) a housing (46) which is arranged on an output side of the box (19) opposite to an input side of the box (19), wherein the input side is the side of the box (19) at which the driving input (24) is present (Fig. 2) and wherein the output side is the side of the box (19) at which the flange (23) is present (Fig. 2).

[0077] The output bearing (26) is inserted into the housing (46), which is a bearing of orientable type and is blocked in place by the annular fixing element (20). The housing (46) is closed by a deformable gasket (45) which is applied to the end of the internal surface of the bell (21) and which can be deformed as a result of the permitted angular inclination between the output pinion (15) and the last stage of the drive system (1), at the same time maintaining a sealed condition of the drive (1 ).

[0078] In the inventive solution the flange (23) incorporates the output pinion (15) of the drive system (1) and is supported thanks to the output pinion (15), allowing the achievement of a more compact and less complex solution, with also a lightening of the drive system (1). This solution also allows the discharge of a part of the weight of the drum (3) and of the stresses through the box (19) of the drive system (1) without causing unwanted stresses to the mechanical components of the drive system (1). The discharge of the weight occurs towards the frame (7) through the box (19) of the drive system (1) via the output bearing (26).

[0079] Advantageously, the output bearing (26) is shaped to allow (Fig. 14) an inclination according to the inclination angle (30) of a few degrees between the drive axis (22) and the drum axis (33) of the concrete mixer vehicle (8) by means of a sliding movement with inclination of the flange (23) and respective output pinion (15) with respect to the bell (21) of the drive system (1). In the preferred embodiment of the present invention (Fig. 6, Fig. 7, Fig. 8, Fig. 9, Fig. 10, Fig. 11 , Fig. 12), the output bearing (26) is a roller bearing with two rows of rolling elements in the form of first rollers (41) and second rollers (42) housed in a containment cage (35) in which the first rollers (41) and the second rollers (42) are arranged according to a condition of reciprocal inclination with respect to each other within the locking cage (35). The inclination according to the inclination angle (30) can occur (Fig. 14) in both directions, i.e. with an inclination angle (30) between the drive axis (22) and the drum axis (33) of the concrete mixer vehicle (8) preferably within a range of at least -3 degrees to +3 degrees, even more preferably within a range of -6 degrees to +6 degrees. The output bearing (26) comprises an external ring (39) equipped (Fig. 7, Fig. 8, Fig. 9) with a first coupling surface (31) arranged radially externally with respect to the output bearing (26). The output bearing (26) comprises an internal ring (40) equipped (Fig. 7, Fig. 8, Fig. 10) with a second coupling surface (32) arranged radially internally with respect to the output bearing (26). The internal ring (40) is inserted (Fig. 7, Fig. 8) inside the external ring (39) in such a way that the first rollers (41) and the second rollers (42) are arranged radially between the internal ring (40) and the external ring (39) and are held in position by the blocking cage (35).

[0080] The housing (46) of the box (19), inside which (Fig. 5) the output bearing (26) is inserted, has a circular shape delimited by a wall (47) for the insertion of the output bearing (26) and the corresponding annular fixing element (20). The output bearing (26), in the condition (Fig. 5) of insertion into the housing (46), is arranged with the first coupling surface (31) of the external ring (39) in a resting and contact condition with the wall (47) of the housing (46) and is further arranged with the second coupling surface (32) of the internal ring (40) in a resting and contact condition with the first portion (16) of the output pinion (15) in such a way that the output bearing (26) allows free rotation of the flange (23) with the respective output pinion (15) with respect to the box (19) and at the same time, thanks to the configuration of orientable type of the output bearing (26), it also allows the described inclination between the drive axis (22) and the drum axis (33).

[0081] Radially internally (Fig. 9), the external ring (39) comprises an external ring surface (Sae) which constitutes an external sliding rail for the first rollers (41) and the second rollers (42), the external ring surface (Sae) being characterised by a concave internal curvature facing towards the inside of the body constituting the external ring (39), where the internal curvature has a curvature radius which in this description will be defined as the external ring radius (Rae). The external ring surface (Sae) develops axially symmetrically with respect to the centre of the external ring (Ce).

[0082] Radially externally (Fig. 10), the internal ring (40) comprises an internal ring surface (Sai) which is composed of a first semi-surface (Sail ) and a second semi-surface (Sai2) reciprocally axially symmetrical with respect to the centre of the internal ring (Ci). The first semi-surface (Sail) constitutes a first internal sliding rail for the first rollers (41), the first semi-surface (Sail) being characterised by a first external curvature which is concave and faces the outside of the body constituting the internal ring (40), where the first external curvature has a curvature radius which in this description will be defined as the first radius of the internal ring (Rail). The second semi-surface (Sai2) constitutes a second internal sliding rail for the second rollers (42), the second semi-surface (Sai2) being characterised by a second external curvature which is concave and facing the outside of the body constituting the internal ring (40), in which the second external curvature has a curvature radius that in this description will be defined as the second radius of the internal ring (Rai2). The first radius of the internal ring (Rail) and the second radius of the internal ring (Rai2) are equal to each other to obtain the described configuration of reciprocal symmetry of first semi-surface (Sail) and second semi-surface (Sai2) with respect to the centre of the internal ring (Ci). In the remainder of this description, the expression radius of the internal ring (Rai) will be used to indicate indifferently the value of first radius of the internal ring (Rail) and second radius of the internal ring (Rai2) which are equal to each other.

[0083] The radius of the external ring (Rae) and the radius of the internal ring (Rai) are such that in the assembled condition of the output bearing (26), the first rollers (41) and the second rollers (42) are arranged inclined (Fig. 7) with respect to a vertical axis of the output bearing (26), according to an inclination angle (I) which may be of the order of 110°, such as between 100° and 130°.

[0084] Each of the first rollers (41) and each of the second rollers (42) has (Fig. 11) a conformation with a roller surface (Sr) which is convex with respect to a roller center (Cr), the roller surface (Sr) constituting the contact surface of the rollers (41 , 42) with the external sliding rail, i.e. with the surface of the external ring (Sae) which is located on the external ring (39), and with the internal sliding rail, i.e. with the first semi-surface (Sail) and the second semi-surface (Sai2) of the internal ring (40). The roller surface (Sr) being characterised by a roller curvature which is convex and facing the outside of the body constituting each of the rollers (41 , 42), where the roller curvature has a curvature radius which in this description will be defined as the roller radius (Rr). The roller surface (Sr) develops axially symmetrically with respect to the center of each of the rollers (41 , 42). In general, the rollers of an orientable roller bearing consist of a spherical shell and of two rectilinear side surfaces. Spherical shell and rectilinear side surfaces are connected by means of a joined surface. The latter is used to prevent that a sharp edge remains between the spherical shell and rectilinear side surfaces. The sharp edge would represent a big problem with bearing in rotation and under load conditions because the roller could get blocked using the sharp edge as a fulcrum and the result would be the breakage of the bearing. The joined surfaces in fact eliminate this possibility and favour the rolling of the rollers with respect to the rails in any speed condition and of load.

[0085] As previously explained, the classic bearings used in concrete mixer vehicle drives do not optimally handle the high axial loads found in this application. The high ratio between axial and radial load causes non-optimal contact between rollers and rails. Indeed, with the experience gained over the years, it has been realized that with the classic configuration the life of the orientable bearing varied greatly between the various pieces even of the same production batch. This high variability in the life of concrete mixer vehicle bearings remained inexplicable for a long time as there were no variability factors between the various production batches that could justify the differences found and even more inexplicably the phenomenon occurs even in the case of bearings from the same production batch that show very different life times between them. Following multiple bench tests with precise axial and radial load values, it was discovered that the problem of the variability of the life of concrete mixer vehicle bearings is due to the high ratio between axial load and radial load typical of the concrete mixer vehicle application as a result of which there is a non-optimal contact between the rollers and the rails of the concrete mixer vehicle bearing even when the concrete mixer vehicle bearing is made according to the typically required specifications.

[0086] To find a solution to this problem, which is only found in the particular application of orientable bearings for concrete mixer vehicles, we tried to reduce the radial clearance range given by the construction tolerances of the components of the output bearing (26). Furthermore, we tried to optimize the contact between the rails and the rollers as much as possible. To do this, an optimal range was identified in which the ratio between the radius of the rails and the radius of the rollers, and consequently also its reciprocal, guarantees obtaining the maximum possible useful life for the bearing with uniformity between bearings of the same batch and between bearings of different batches. The optimization of the ratio of the radii allows to maximize the contact area between the rollers and the rails by decreasing the surface pressure. All this was done taking into account the production limits.

[0087] In particular, following studies, design and numerous tests, it was found that to obtain the desired uniformity in the duration or average life between bearings of the same batch and between bearings of different batches, it is essential to respect very precise construction parameters. In particular, in order to obtain the advantages described, in order to improve both the average life of the single output bearing (26) for drive system (1) for drum (3) of the concrete mixer vehicle (8) and to obtain a uniform average life parameter for the output bearings (26) of the same production batch, the applicants have discovered that it is necessary to produce the output bearing (26) for the drive system (1) according to characteristics that are different from those prescribed by the applicable standards, which, in fact, constitute for an average technician in the sector an indication contrary to what is the teaching of the present invention. Indeed, contrary to what is provided for by the standards on the production of this type of bearing, the present invention provides that the output bearing (26) for drive system (1) for drum (3) of the concrete mixer vehicle (8) is characterised by:

[0088] - the curvature of the roller surface (Sr) and the curvature of the surface of the external ring (Sae) constituting the external sliding rail of the rollers (41 , 42) are made with a level of precision such that the ratio between the roller radius (Rr) and the radius of the external ring (Rae) is equal to a Fext value which is between 98.90% and 100%, preferably between 99.00% and 99.80%, even more preferably between 99.35% and 99.60%;

[0089] - the curvature of the roller surface (Sr) and the curvature of the semi-surfaces of the internal ring (Sail , Sai2) constituting the internal sliding rails of the rollers (41 , 42) are made with a level of precision such that the ratio between the roller radius (Rr) and the radius of the internal ring (Rai) is equal to a Fint value which is between 99.00% and 100%, preferably between 99.35% and 99.60%.

[0090] More preferably, Fext and Fint are at least equal to 99.00%, preferably at least equal to 99.60%, the preferred value being 100%. With reference to the upper limit, although reaching 100% is preferable, depending on the application and the manufacturing costs, values of Fext and Fint lower than 99.80%, preferably lower than 99.70%, may be acceptable.

[0091] Even more preferably, the radial clearance between the roller surface (Sr) and the surface of the outer ring (Sae) is between 0.040 and 0.080 mm and the radial clearance between the roller surface (Sr) and the surface of the internal ring (Sai), i.e. the semi-surfaces of the internal ring (Sail , Sai2), is between 0.040 and 0.080 mm.

[0092] As previously explained, the radial clearance is a a known parameter to an expert in the art of the bearings which is an intrinsic constructive parameter of all the bearings and defines the radial movement admitted between rails and rolling bodies. Obviously, the radial clearance cannot be zero, otherwise would constitute an obstacle to the rolling motion. For the application concerning a gear reducer for a concrete mixer vehicle and for this bearing in particular, reducing the radial clearance at minimum equates to maximizing the contact surface between the rollers and the respective rolling rails in the load condition. Therefore, the obtainment of a high contact degree and of a minimum radial clearance represent two ways to maximize the contact surface between the rollers and the respective rolling rails in the load condition. The bearing according to the present invention is made with the combination of these two strategies. As the contact degree increases, both the contact surface between the roller surface (Sr) and the external ring surface (Sae) and the contact surface between the roller surface (Sr) and the internal ring surface of (Sai) increase. Obviously, given a determined contact degree as a design parameter, as the applied load increases, the real contact surface in the use conditions of the bearing increases, that is increase both the contact surface between the roller surface (Sr) and the external ring surface (Sae) and the contact surface between the roller surface (Sr) and the internal ring surface (Sai) increase. This is due to the deformations to which are subject the components of the bearing in the load conditions, obtaining as an effect that the sections of the rollers and of the rails tend to reciprocally align, this occurring in any use condition at each inclination angle between the external ring of the bearing and the internal ring of the bearing.

[0093] Only with these characteristics can it be ensured that the contact between the rollers (41 , 42) and the corresponding sliding rails is in the order of 97%, contrary to the solutions of the prior art in which a contact between the rollers (41 , 42) and the corresponding sliding rails of less than 92% was obtained. It is precisely thanks to these characteristics that the surprising effect of lengthening the useful life of the output bearing (26) is obtained as well as the desired uniformity between bearings of the same batch and between bearings of different batches.

[0094] In particular, the result is obtained using a particular conformation of the curvature of the roller surface (Sr) and of the corresponding curvature of the external ring surface (Sae) constituting the outer sliding rail of the rollers (41 , 42) and the curvature of the semi-surfaces of the internal ring (Sail , Sai2) constituting the internal sliding rails of the rollers (41 , 42). Indeed, the indicated curvatures are calculated using the following formula:

[0095] The formula is applicable to all the bearings of the same type, that is orientable with rollers, used for this application in which the main bearing supports the exit flange of a gear- reducing device for a concrete mixer vehicle. As a result, for the purpose of the design of a bearing with different sizes, remain valid the ratio between the radius of the rollers and the radius of the sliding rails, in addition to the radial clearance.

[0096] In the formula (1), Rr indicates the curvature radius of the roller, Rp indicates the radius of the rails. The radius of the rail, according to what is defined in this description, can be the external ring radius (Rae) relative to the external ring surface (Sae) or the internal ring radius (Rai) relative to the first semi-surface (Sail) which constitutes a sliding internal rail for the first rollers (41) and relative to the second semi-surface (Sai2) which constitutes a sliding internal rail for the second rollers (42).

[0097] The parameter Kp is an experimental coefficient defined on the base of the matured experience during multiple tests performed. This parameter is obtained by the correlation between the ratio Rr / Rp of the real tested bearing and the duration in terms of hours of the latter during the test at the bench. As the parameter Kp increases, one obtains that Rr / RpMiN decreases and therefore also the bearing life decreases. As the parameter Kp decreases, one obtains that Rr / RpMiN increases and therefore also the bearing life increases.

[0098] The theoretic ideal parameter Kp used in the calculation is preferably in the range between 72,335 and 72,344. The effective parameter Kp obtained for the bearing has to be considered acceptable if it is in the range between 72,302 and 72,404.

[0099] The multiplicative value “0.67” is a fixed value coefficient used in the base formula for the calculation of the life of a bearing of this type, which assumes this value in the case of calculation of the equivalent load for the roll bearings of the similar type to the inventive one, according to the formula:

[0100] Both for the formula (1) and for the formula (2), the parameter Fr indicates the design radial load for the bearing expressed in newton, the radial load being the force acting on the bearing in the radial direction, or perpendicularly to the rotation axis of the bearing in the working conditions. The parameter Fr relative to the radial load contributes to the definition of the cycle of useful life of the bearing. Indeed, starting from a determined radial load condition provided as design specification, applying a greater radial load a decrease of the useful life of the bearing occurs, while applying a lower radial load an increase of the life of the bearing occurs. A realistic data for the application of bearing for a concrete mixer vehicle having a size of 12 m3, with a fully loaded drum, is Fr = 150 kN. A realistic data for the application of bearing for a concrete mixer vehicle having a size of 8 m3, with a fully loaded drum, is Fr = 105 kN. A realistic data for the application of bearing for a concrete mixer vehicle having a size of 14 m3, with a fully loaded drum, is Fr = 180 kN.

[0101] Both for the formula (1) and for the formula (2), the parameter Fa indicates the design axial load for the bearing expressed in newton, in which the axial load is the force acting on the bearing in the axial direction, or parallel to the rotation axis of the bearing. The parameter Fa relative to the axial load contributes to the definition of the cycle of useful life of the bearing. Indeed, starting from a determined axial load condition provided as design specification, applying a greater axial load a decrease of the useful life of the bearing occurs, while applying a lower axial load an increase of the life of the bearing occurs. A realistic data for the application of bearing for a concrete mixer vehicle having a size of 12 m3, with a fully loaded drum, is Fa = 55 kN. A realistic data for the application of bearing for a concrete mixer vehicle having a size of 8 m3, with a fully loaded drum, is Fa = 45 kN. A realistic data for the application of bearing for a concrete mixer vehicle having a size of 14 m3, with a fully loaded drum, is Fa = 60 kN.

[0102] Both for the formula (1) and for the formula (2), the parameter Y2_max is a coefficient for maximum axial load, directed to consider a worst condition. It is an intrinsic parameter of the bearing defined by the manufacturer. The value of the parameter Y2 depends from the constructive geometry of the bearing, it is defined by the manufacturer and is a part of the design data of the bearing itself. From the matured experience the coefficient Y2 can be in the range 2 - 3,5. For the application for the present invention, in the formula (1 ) it is preferable to use a value Y2 between 2.2 and 3.7, preferably a value Y2 equal to 3.5, corresponding to a worst condition of axial load, considering the fact that the greater Y2 is, the greater the weight of the axial load is in the determination of the equivalent load (Peq) acting on the bearing. Indeed, starting from a determined axial load condition provided as design specification, applying a greater real value of the parameter Y2real a decreasing is obtained of the useful life of the bearing, while applying a lower real value of the parameter Y2real an increase of the life of the bearing is obtained.

[0103] The parameter L10h indicates the life, that is the duration, of the bearing expressed in hours and it is calculated according to the formula:

[0104] 10

[0105] In practice, the parameter L10h indicates the theoretic duration expressed in hours of the bearing in a determined load condition, defined by radial load, axial loaded and rotation speed. The notation L10h, or fatigue nominal duration, in the bearings refers to the average duration in hours that a group of bearings, with a breakdown probability of 10%, can reach before breakdown due to fatigue. In other words, 90% of the bearings should last longer. This duration is calculated on the basis of different factors, among which the applied load, the rotation speed and the characteristics of the bearing itself. The select values are at discretion of the result that one wants to obtain. In our case, an objective L10h value has been defined with greater life objective with respect to all the tests made in precedence with the constructive solution of prior art and based on this the design procedure has been started. By a purely mathematical point of view, if the target L10h increases, a greater ratio Rr / Rp will be obtained, if the target L10h decreases, a lower ration Rr / Rp will be obtained. The exponent 10 / 3 is the exponent of raising to a power applicable for the formula of the calculation of the parameter L10h valid for the case of the roll bearings.

[0106] The C parameter indicates the dynamic load factor expressed in newton. It is an intrinsic parameter of the bearing supplied by the manufacturer and it is part of the design data. The grater C is, the greater is the life of the bearing with the same load conditions, that indicatively can be between 300 kN and 700 kN. The parameter n indicates the rotational speed of the shaft supported by the bearing expressed in rpm, that is the central shaft of the flange (23). Usually, the drum is operated with a rotation speed between 0.1 rpm and 20 rpm, with a maximum rotation speed of about 14-20 rpm.

[0107] For the purposes of the present invention and for the shown roller bearing in the case of application on ratio-motor for concrete mixer vehicle with a capacity of 12 m3, and with a fully loaded drum, real values for the use in the indicated formulas can be Fr = 150 kN, Fa = 55 kN, n = 15 rpm. For the purposes of the present invention and for the shown roller bearing in the case of application on ratio-motor for concrete mixer vehicle with a capacity of 8 m3, and with a fully loaded drum, real values for the use in the indicated formulas can be Fr = 105 kN, Fa = 45 kN, n = 15 rpm. For the purposes of the present invention and for the shown roller bearing in the case of application on ratio-motor for concrete mixer vehicle with a capacity of 14 m3, and with a fully loaded drum, real values for the use in the indicated formulas can be Fr = 180 kN, Fa = 60 kN, n = 15 rpm.

[0108] In practice, starting from the formulas (2) and (3) which are known to an expert in the art of the bearings, making in the formula (3) explicit the value of Peq given by the formula (2) and successively resolving the so obtained formula with respect to the C parameter, ones obtains:

[0109] At this point, the C dynamic load factor is put in relation with the parameter that the applicants have figured out be the responsible of the performances of the bearing in terms of duration in case of the specific application of a concrete mixer vehicle, that is the ratio between the radiuses Rr / Rp, according to the following formula: The connection between the C dynamic load factor and the ration Rr / Rp is represented by the experimental coefficient kp, defined following conspicuous experimental investments, research and development.

[0110] In this way, the equation (1) is obtained which allows to calculate the ratio [Rr / Rp] as minimum project value of the bearing for which one wishes to have a greater useful life with respect to the useful life of the bearings of prior art. [Rr / Rp]min is the minimum value of the ratio Rr / Rp to be used as constructive parameter of the bearing in order to be able to obtain the objective duration of the bearing inserted in the formula, with the inserted load parameters in the formula. Once the minimum is obtained, the maximum is calculated according to the productive limits of the manufacturer of the bearing.

[0111] To set up a real example, in case of the inventive bearing [Rr / Rp]min equal to 99,50% has been obtained, a theoretic upper bound was imposed adding 2 hundredths of percentage obtaining an ideal range [Rr / Rp]teor between 99,50% and 99,52%.

[0112] About the ratio [Rr / Rp] the theoretic ideal range is between 99,50% and 99,52%, a range however acceptable is between 99,37% and 99,59%. The indicated values are effective for all the bearings of this type and for this application. Setting in the calculation phase lower values of the parameter [Rr / Rp] with respect to what indicated, a bearing will be obtained with a minor contact degree, with the consequence that, with the same loading conditions applied the bearing will have a time of shorter useful life. Setting in the calculation phase higher values of the parameter [Rr / Rp] with respect to what indicated, one runs a chance to increase the costs of construction of the bearing without obtaining any additional benefit in terms of increase of the time of useful life. Furthermore, setting in the calculation phase value of the parameter [RR / Rp] beyond what indicated, particularly with values close to 1 , there is the risk that under load conditions the rails-rollers contact is not optimal and as a result one obtains a contrary effect with respect to what is desired, that is a shorter useful life of the bearing with the same loading conditions applied, because of premature wearing due to not-optimal contact between rollers and sliding rails.

[0113] A verification phase follows of feasibility carried out on the base of the available machines and manufacturing process, in which the theoretic range [Rr / Rp]teor indicated is corrected obtaining the reference value [Rr / Rp] for the construction of the bearing, which in the case of the example, has carried to a reference value [Rr / Rp] in a range from 99,37% to 99,59%.

[0114] Once Rr / Rp is defined, absolute values of Rr and Rp have been select on the basis of the total sizes of the bearing and in such a way as to respect the range of the reference value [Rr / Rp] for the construction of the bearing.

[0115] It is necessary to put in evidence that once the design "constrain" has been determined represented by the reference value [Rr / Rp] for the construction of the bearing, the choice of the values of the radius of rollers and rails is something belonging to the constructive choices of the physical manufacturer of the bearing and does not have any influence on the result of the present invention.

[0116] It is necessary to put in evidence that the obtained formula and the described results are not the result of a mere arbitrary selection which can be identified by an expert in the art. All the bearings of all the suppliers that have been tested during the years have shown the same limit in terms of duration; as a result one can deduce that all the manufacturers have always adopted similar design choices and different with respect to the solution according to the present invention. Following conspicuous investments and experimental tests has been figured out in detail the reason of this duration limit, determining the limiting factor and formulating and putting in practice an inventive project solution thanks to which a considerable result is obtained in terms of increase of the useful life of the bearing for the specific application characterized by a very low bearing rotation speed, between 0.1 rpm and 20 rpm as in the case of bearing of exit of a gear-reducing device for concrete mixer vehicle. In substance, summing up, applicants have been concentrating experience and inventiveness to solve a problem in relation to a niche application, that up to now had represented a limit of the state of the art. If a mere selection was sufficient on the part of an expert in the art to get to an improvement, it is difficult to comprise why no manufacturer up to now has never adopted this solution.

[0117] Specifically for the bearing for concrete mixer vehicle made according to the present invention, the following values have been selected which respect the reference value [Rr / Rp] for the construction of the bearing calculated according to the described method:

[0118] - Rr / Rp = 99,37% - 99,59%

[0119] - Rr = 83,871 mm

[0120] - Rae or Rp_est = 84,288 mm corresponding to Rr / Rae equal to 99,50%

[0121] - Rai or Rp_int = 84,393 mm corresponding to Rr / Rai equal to 99,38%

[0122] Successively, prototype bearings have been manufactured, tests on the bench have been carried out and it has been vefied that the proposed reference range [Rr / Rp] is acceptable in terms of obtained bearing duration.

[0123] Since it is a particular application, that is a bearing for the rotation of the drum of a concrete mixer vehicle that rotates with a substantially small speed in the range between 0.1 rpm and 20 rpm, in general with maximum rotation speed values between 14 rpm and 20 rpm, it is important to always have clear that the bearing according to the present invention is specially designed for the related application to the rotary support with possible inclination of the output shaft of the flange (23) that is connected to the drum of the concrete mixer vehicle or for any other application subjected to the same load conditions. The essential characteristic that has allowed to be able to push at most on the maximization of the relation Rr / Rp and minimization of radial clearance is represented by the low rotational regimen of this bearing in the range between 0.1 rpm and 20 rpm. Substantially it is a semi-static application and as a result the influence of the concepts differential slip, frictional resistance and dynamic slip is minimal to a point to be non-influential. In general, though holding true that maximizing Rr / Rp and minimizing the radial clearance the obstacle to the reciprocal sliding is increased between the rollers and the sliding rails and the sliding friction is increased, it is furthermore true, and still most essential, that being the rotational state of this bearing so low with maximum speed 20 rpm, such aspects are absolutely negligible. And the obtained result constitutes a proof: if we would have introduced obstacles to the sliding and incremented the friction, the duration of the bearing would be reduced at the same load conditions, instead with the described method this duration has been incremented by far. The formula (1) allows to determine Rr / RpMIN, then, as explained, a range is established Rr / RpMIN - Rr / RpMAX. Therefore, no constraint is introduced between the radiuses of the internal and external rails. The important thing is that, for each rail, either external or internal, the constraint is valid that the ration Rr / Rp falls inside the determined range. Furthermore, it is provided that the radial clearance of the bearing is selected by design and the assembly seats of the bearing are dimensioned, that is the internal shaft for the internal ring and the housing of insertion for the external ring, on the gear-reducing device in such a way as to obtain a radial clearance in work, that is under loaded conditions, such that the duration of the bearing is maximized. As explained, the method is applicable to orientable bearings for the application on concrete mixer vehicles, keeping in consideration that the applied parameters vary according to the capacity of the drum of the concrete mixer vehicle, that influences the applied loads on the bearing, as shown in the following table 1.

[0124] TABLE 1 Furthermore it is pointed out that the described method, applying the indicated parameters, is independent from the size of the bearing which is selected according to the capacity of the drum of the concrete mixer vehicle according to the following table 2. TABLE 2

[0125] For example, with a roller (41 , 42) having a width of approximately 33 mm, a diameter of approximately 18 mm, the roller radius (Rr) may have values between 80 mm and 90 mm, such as for example 83.6 mm. Correspondingly, the external ring (39) may have a body width of the order of 84 mm + / - 0.2 mm, a body diameter of the order of 180 mm + / - 0.03 mm, with the external ring radius (Rae) which may have values between 80.4 mm and 90.4 mm, such as for example 84 mm. Correspondingly, the internal ring (40) may have a body width of the order of 75 mm + / - 0.2 mm, a body diameter of the order of 110 mm + / - 0.02 mm, with an internal ring radius (Rai) that may have values between 80.4 mm and 90.4 mm, such as for example 84 mm.

[0126] In general, the bearings of the inventive solution allow to obtain an 80% longer bearing life compared to the bearings of the prior art. The data from some tests carried out that demonstrate this notable improvement are reported below, in addition to the fact that the variability in the average life is significantly reduced even between bearings belonging to the same production batch.

[0127] For example (Fig. 15), considering the results of a test of the cycle of useful life applied to bearings for concrete mixer vehicle of prior art and applied to inventive bearings for concrete mixer vehicle, having equal test conditions (applied load, rotational speed, inclination), it is evident the difference that we can obtain with the inventive bearing which shows a prolonged useful life of 80% more with respect to the bearings of prior art.

[0128] With the bearings of the prior art, there is a high variability in the average life between bearings belonging to the same production batch. For example, compared to the total average life of the bearings of the same batch of prior art bearings, it is possible to have bearings with a life equal to -40% of the average life of the bearings of the batch and bearings with a life equal to +40% of the average life of the bearings of the batch. With the described solution, instead, the inventive bearings have very low variability in average life. For example, compared to the total average life of the bearings of the same batch of inventive bearings, it is possible to have bearings with a life equal to -10% of the average life of the bearings of the batch and bearings with a life equal to +10% of the average life of the bearings of the batch. These values are indicative of the significant improvement that is obtained and that has a significant impact on the performance of the bearing, whose average life is thus a value that can be defined with greater precision and which therefore allows for efficient planning of maintenance and replacement interventions, avoiding unpredictable failures due, for example, to the presence in the same batch of bearings with an average life of -40% and bearings with an average life of +40%.

[0129] The rolling surface of the rollers (41 , 42), i.e. the roller surface (Sr), is preferably treated to give this rolling surface increased wear resistance by means of, for example, surface hardening treatments, preferably surface hardening treatments in the form of induction hardening at the rolling surface.

[0130] Ultimately, the present invention relates to (Fig. 5, Fig. 6, Fig. 7) an output bearing (26) for an output pinion (15) for (Fig. 1 , Fig. 2, Fig. 3, Fig. 4) a drive system (1) for a concrete mixer vehicle (8), wherein the output pinion (15) is provided with a fixing flange (23) protruding externally with respect to the drive (1) for putting in rotation a drum (3) of the concrete mixer vehicle (8), the output bearing (26) being of the orientable type and comprising (Fig. 6, Fig. 7, Fig. 8, Fig. 9, Fig. 10) an internal ring (40) and an external ring (39) arranged coaxially with interposition of sliding rollers (41 , 42) consisting of first sliding rollers (41) and second sliding rollers (42) wherein each of the rollers (41 , 42) has a roller surface (Sr) which is curved and convex with a curvature radius according to a roller radius (Rr), where a center of the internal ring (Ci) and a center of the external ring (Ce) are coincident with a center of the bearing (C). The output bearing (26) is configured in such a way as to allow a condition of free rotation of the second coupling surface (32) by reciprocal rotation between the internal ring (40) and the external ring (39) around a longitudinal axis (49). The output bearing (26) is configured in such a way as to allow a condition of free reciprocal inclination between the internal ring (40) and the external ring (39) with respect to a longitudinal axis (49) according to (Fig. 13) an inclination angle (30) by inclination of the second coupling surface (32) with respect to the first coupling surface (31). A first set of first rollers (41) and a second set of second rollers (42) are arranged between the internal ring (40) and the external ring (39) on opposite sides of the output bearing (26) with respect to (Fig. 6) a central plane (P) of symmetry on which said transverse axis (34) lies. The external ring (39) includes radially internally an external ring surface (Sae) which constitutes an external sliding rail for the first rollers (41) and the second rollers (42), the external ring surface (Sae) being a curved surface with a curvature radius according to a radius of the external ring (Rae). The internal ring (40) comprises radially externally a first semi-surface (Sail) which constitutes an internal sliding rail for the first rollers (41) and a second semi-surface (Sai2) which constitutes an internal sliding rail for the second rollers (42), the first semi-surface (Sail ) and the second semi-surface (Sai2) of the internal ring being curved surfaces with a curvature radius according to a radius of the internal ring (Rai). Between the roller surface (Sr) and the internal rails and the external rail there is a degree of contact greater than 95%. Preferably between the roller surface (Sr) and the internal rails and the external rail there is a degree of contact of at least 97%.

[0131] Under the conditions of use, the radial load supported by the output bearing (26) is such that the radial force supported is in a range between 70 kN and 180 kN.

[0132] In the conditions of use, the axial load supported by the output bearing (26) is such that the supported axial force is in a range between 30 kN and 60 kN.

[0133] The inclination angle (30) for reciprocal inclination between the internal ring (40) and the external ring (39) with respect to the transverse axis (34) is between -10 degrees and +10 degrees, preferably between -6 degrees and +6 degrees, it being evident that multiple configurations will be possible in which the inclination angle (30) can assume different values depending on the size of the drive (1 ) and the drum of the concrete mixer vehicle.

[0134] Furthermore, the present invention also relates to (Fig. 2, Fig. 3, Fig. 4) a drive (1) for (Fig. 1) a drum (3) of a concrete mixer vehicle (8) comprising an epicyclic multistage gear-reduction device, wherein the drive (1) comprises (Fig. 4) a structural box (19) containing the components of the epicyclic gear-reduction device, the drive (1) being provided with a driving input (24) for connection of a drum motor (2) for driving a first stage of the reduction stages of the epicyclic gear-reduction device, the drive (1) comprising an output flange (23) for fixing to the drum (3) and the drive of rotational motion from the drum motor (2) to the drum (3), the output flange (23) comprising an output pinion (15) which is coupled with the last reduction stage of the epicyclic gear-reduction device of the drive (1), the output pinion (15) being supported internally to the drive (1) by a respective output bearing (26) made as described, so that the output bearing (26) allows a condition of free rotation of the flange (23) with the respective output pinion (15) with respect to the box (19) and an inclination between a drive axis (22) and a drum axis (33) according to said inclination angle (30).

[0135] In particular (Fig. 4, Fig. 5, Fig. 9) the external ring (39) of the output bearing (26) comprises radially externally a first coupling surface (31) for assembly and internal coupling to the drive (1), the drive (1) comprising a housing (46) obtained on the box (19), the output bearing (26) being arranged within the housing (46), the housing (46) having a circular shape delimited by a wall (47) for coupling with the first coupling surface (31) of the output bearing (26), in such a way that there is a condition of discharge of radial loads applied to the output bearing (26) from the output pinion (15) towards the box (19) of the drive (1). Furthermore, the output pinion (15) comprises (Fig.4, Fig. 5) a first portion (16) for coupling with the output bearing (26), the internal ring (40) of the output bearing (26) comprising radially internally a second coupling surface (32) for assembly and coupling with the output pinion (15) and the output pinion (15) comprises (Fig.4, Fig. 5) a second portion (17) for coupling with a grooved coupling (38) shaped in such a way as to allow said inclination according to said inclination angle (30), by means of a bell (21 ) which constitutes an element of interconnection and drive of the rotational motion between the last stage of the drive system (1) and the second portion (17) of the output pinion (15) via the grooved coupling (38) with drive of the rotational motion independent of the value of the inclination angle (30).

[0136] The described method can be used for any dimension of the bearing. The result which is obtained from the application of the method is the ideal ratio Rr / Rp for the obtainment the desired result, that is the increase of the mean life of the bearing. The ratio ideal Rr / Rp is, as indicated, between 99.50% and 99.52%, with acceptable values that are however between 99.37% and 99.59%. The method and the individuated values are valid for any orientable roller bearing for the use on an exit shaft of a gear-reducing device for concrete mixer vehicle. The method and the individuated values, instead, are not valid for an orientable roller bearing for a different use with respect to the indicated one in which the rotational speeds of the bearings are very greater, as for example greater then 100 rpm because the described method is valid exclusively for rotational low-speed bearings, as for example less than 50 rpm.

[0137] The present invention also relates to a (Fig. 1 ) concrete mixer vehicle (8) comprising a mixing drum (3) wherein the concrete mixer vehicle (8) comprises a drive (1) for the drum (3) of the concrete mixer vehicle (8) according to what described, i.e. comprising the output bearing (26) having the characteristics previously set forth.

[0138] In the end, the present invention relates to a production method of an output bearing (26) for an output pinion (15) for a drive (1 ) for a concrete mixer vehicle (8), in which the output pinion (15) is provided with a flange (23) protruding outside with respect to the drive (1 ) for putting in rotation a drum (3) of the concrete mixer vehicle (8), the output bearing (26) being orientable and comprising an internal ring (40) and an external ring (39) coaxially placed with interposition of first sliding rollers (41) and second sliding rollers (42) in which each of the rollers (41 , 42) has a roller surface (Sr) curved and convex with curvature radius according to a roller radius (Rr), the output bearing (26) being configured for a free rotation condition between internal ring (40) and external ring (39) around a longitudinal axis (49) and for a free reciprocal inclination condition between the internal ring (40) and the external ring (39) with respect to the longitudinal axis (49) according to an inclination angle (30), a first series of first rollers (41) and a second series of second rollers (42) being placed between the internal ring (40) and the external ring (39) on opposite sides of the output bearing (26) with respect to a central plane (P) of symmetry on which a symmetry transverse axis (34) lies, the external ring (39) comprising radially internally an external ring surface (Sae) which constitutes a sliding external rail for the first rollers (41) and the second rollers (42), the external ring surface (Sae) being a curved surface with curvature radius according to an external ring radius (Rae), the internal ring (40) comprising radially externally a first semi-surface (Sail) which constitutes a sliding internal rail for the first rollers (41) and a second semi-surface (Sai2) which constitutes a sliding internal rail for the second rollers (42), the first semisurface (Sail) and the second semi-surface (Sai2) of the internal ring being curved surfaces with curvature radius according to an internal ring radius (Rai), the roller surface (Sr) constituting the contact surface of the rollers (41 , 42) with the external rolling rail consisting of the external ring surface (Sae) that is on the external ring (39), and with the internal rolling rail consisting of the first semi-surface (Sail) and of the second semi-surface (Sai2) of the internal ring (40), in which the output bearing (26) is configured for a rotational speed of the output pinion (15) between 0.1 rpm and 30 rpm, the method comprising a calculation process of constructive parameters of the bearing comprising the following phases:

[0139] Phase (A): calculation of a parameter corresponding to the ratio Rr / Rp minimum between the roller radius (Rr) and a rail radius (Rp) which is indicative of corresponding values of the external ring radius (Rae) and of the internal ring radius (Rai), the calculation occurring by means of the formula: in which:

[0140] - kp is a correlation parameter with respect to a dynamic load factor, the parameter kp being between 72,335 and 72,344;

[0141] - the parameter Fr indicates the design radial load expressed in newton, the parameter Fr being between 100 kN and 200 kN, preferably between 120 kN and 180 kN, the preferred value being 150 kN:

[0142] - the parameter Fa indicates the design axial load expressed in newton, the parameter Fa being between 40 kN and 70 kN, preferably between 45 kN and 65 kN, the preferred value being 55 kN:

[0143] - the parameter Y2_max is a coefficient for maximum axial loaded, the parameter Y2_max being between 2.2 and 3.7, the preferred value being 3.5;

[0144] - the parameter n indicates the rotation speed of the output pinion (15), the parameter n being between 10 rpm and 30 rpm, preferably between 15 rpm and 25 rpm;

[0145] - the parameter L10h indicates the theoretic duration of the bearing expressed in hours, expressed as average duration in hours that a group of bearings, with a breakdown probability of 10%, can reach before a breakdown, the parameter L10h being between 200 hours and 1000 hours;

[0146] Phase (B): following the calculation of the parameter Rr / Rp minimum, selection of a value Rr / Rp maximum within the range:

[0147] In practice, a maximum Rr / Rp value is imposed, increased by an amount between 1.5 cents and 3 cents with respect to the minimum Rr / Rp, this procedure being valid regardless of the size of the bearing.

[0148] Phase (C): following the calculation of the parameter Rr / Rp minimum and of the parameter Rr / Rp maximum, selection of values of construction of the bearing Rr' and Rp' in such a way that that the ratio Rr' / Rp' is:

[0149] - greater or equal to with respect to the calculated ratio Rr / Rp minimum reduced by a first value of admissible lower deviation equal to 0.2%;

[0150] - lower or equal to with respect to the selected ratio Rr / Rp maximum increased by a second value of admissible upper deviation equal to 0.2%.

[0151] Phase (D) final verification that the ratio [Rr' / Rp1] for the construction of the bearing is between Rr / Rp minimum and Rr / Rp maximum and if not repetition of the phase (C).

[0152] The description of the present invention has been made with reference to the attached figures in a preferred embodiment thereof, but it is evident that many possible alterations, modifications and variations will be immediately clear to those skilled in the art in the light of the foregoing description. Thus, it should be emphasized that the invention is not limited by the foregoing description, and includes all alterations, modifications and variations in accordance with the attached claims.

[0153] NOMENCLATURE USED

[0154] With reference to the identification numbers shown in the attached figures, the following nomenclature has been used:

[0155] 1. Drive system or drive

[0156] 2. Drum motor

[0157] 3. Drum

[0158] 4. Tractor motor

[0159] 5. Support

[0160] 6. Tractor

[0161] 7. Frame

[0162] 8. Concrete mixer vehicle

[0163] 9. Pinion of the first stage

[0164] 10. Planetary-gear of the first stage

[0165] 11. Planetary-gear support of the first stage

[0166] 12. Pinion of the second stage

[0167] 13. Planetary-gear of the second stage

[0168] 14. Planetary-gear support of the second stage

[0169] 15. Output pinion

[0170] 16. First portion

[0171] 17. Second portion

[0172] 18. Cover

[0173] 19. Box

[0174] 20. Annular fixing element

[0175] 21. Bell

[0176] 22. Drive axis

[0177] 23. Flange 24. Driving input

[0178] 25. Fixing means

[0179] 26. Output bearing

[0180] 27. Input bearing

[0181] 28. First stage bearing

[0182] 29. Second stage bearing

[0183] 30. Inclination angle

[0184] 31. First coupling surface

[0185] 32. Second coupling surface

[0186] 33. Drum axis

[0187] 34. Transverse axis

[0188] 35. Cage

[0189] 36. Tooth

[0190] 37. Groove

[0191] 38. Grooved coupling

[0192] 39. External ring

[0193] 40. Internal ring

[0194] 41. First roller

[0195] 42. Second roller

[0196] 43. Shaft

[0197] 44. Recess

[0198] 45. Gasket

[0199] 46. Housing

[0200] 47. Wall

[0201] 48. Coupling gear

[0202] 49. Longitudinal axis

[0203] C. Bearing center

[0204] Ce. External ring center Ci. Internal ring center

[0205] Cr. Roller center

[0206] I. Inclination

[0207] P. Plane Rae. External ring radius

[0208] Rai. Internal ring radius

[0209] Rail . First radius of the internal ring

[0210] Rai2. Second radius of the internal ring

[0211] Rr. Roller radius Sae. External ring surface

[0212] Sai. Internal ring surface

[0213] Sail . First semi-surface

[0214] Sai2. Second semi-surface

[0215] Sr. Roller surface

Claims

CLAIMS1. Production method of an output bearing (26) for an output pinion (15) for a drive (1) for a concrete mixer vehicle (8), in which the output pinion (15) is provided with a flange (23) protruding outside with respect to the drive (1) for putting in rotation a drum (3) of the concrete mixer vehicle (8), the output bearing (26) being orientable and comprising an internal ring (40) and an external ring (39) coaxially placed with interposition of first sliding rollers (41) and second sliding rollers (42) in which each of the rollers (41 , 42) has a roller surface (Sr) curved and convex with curvature radius according to a roller radius (Rr), the output bearing (26) being configured for a free rotation condition between internal ring (40) and external ring (39) around a longitudinal axis (49) and for a free reciprocal inclination condition between the internal ring (40) and the external ring (39) with respect to the longitudinal axis (49) according to an inclination angle (30), a first series of first rollers (41) and a second series of second rollers (42) being placed between the internal ring (40) and the external ring (39) on opposite sides of the output bearing (26) with respect to a central plane (P) of symmetry on which a symmetry transverse axis (34) lies, the external ring (39) comprising radially internally an external ring surface (Sae) which constitutes a sliding external rail for the first rollers (41) and the second rollers (42), the external ring surface (Sae) being a curved surface with curvature radius according to an external ring radius (Rae), the internal ring (40) comprising radially externally a first semisurface (Sail) which constitutes a sliding internal rail for the first rollers (41) and a second semisurface (Sai2) which constitutes a sliding internal rail for the second rollers (42), the first semisurface (Sail) and the second semi-surface (Sai2) of the internal ring being curved surfaces with curvature radius according to an internal ring radius (Rai), characterised in that the output bearing (26) is configured for a rotational speed of the output pinion (15) between 0.1 rpm and 30 rpm, the method comprising a calculation process of constructive parameters of the bearing for obtainment of a output bearing (26) with characteristics of improved mean life, the calculation process comprising at least one phase (A) of calculation of a parameter corresponding to the ratio Rr / Rp minimum between the roller radius (Rr) and a rail radius (Rp) which is indicative of corresponding values of the external ring radius (Rae) and of the internalring radius (Rai), the calculation occurring by means of the formula:in which:- kp is a correlation parameter with respect to a dynamic load factor, the parameter kp being between 72,335 and 72,344;- the parameter Fr indicates the radial load expressed in newton;- the parameter Fa indicates the axial load expressed in newton:- the parameter Y2_max is a coefficient for maximum axial loaded, the parameter Y2_max being between 2.2 and 3.7, the preferred value being 3.5;- the parameter n indicates the rotation speed of the output pinion (15), the parameter n being between 10 rpm and 30 rpm, preferably between 15 rpm and 25 rpm;- the parameter L10h indicates the theoretic duration of the bearing expressed in hours, expressed as average duration in hours that a group of bearings, with a breakdown probability of 10%, can reach before a breakdown, the parameter L10h being between 200 hours and 1000 hours.

2. Production method of an output bearing (26) according to the previous claim, characterised in that the parameter Fr is between 100 kN and 200 kN, preferably between 120 kN and 180 kN, the preferred value being 150 kN.

3. Production method of an output bearing (26) according to any of the previous claims, characterised in that the parameter Fa is between 40 kN and 70 kN, preferably between 45 kN and 65 kN, the preferred value being 55 kN.

4. Production method of an output bearing (26) according to any of the previous claims, characterised in that it further includes a phase (B) which, following the calculation of the parameter Rr / Rp minimumin the phase (A), is a selection phase of a value Rr / Rp maximum within the range:

5. Production method of an output bearing (26) according to the previous claim, characterised in that it further includes a phase (C) which, following the calculation of the parameter Rr / Rp minimum and of the parameter Rr / Rp maximum, is a selection phase of values of construction of the bearing Rr' and Rp' in such a way that that the ratio Rr' / Rp' is- greater or equal to with respect to the calculated ratio Rr / Rp minimum reduced by a first value of admissible lower deviation equal to 0.2%;- lower or equal to with respect to the selected ratio Rr / Rp maximum increased by a second value of admissible upper deviation equal to 0.2%.

6. Production method of an output bearing (26) according to the previous claim, characterised in that it further includes a phase (D) which is a verification phase that the ratio [Rr' / Rp1] for the construction of the bearing is between Rr / Rp minimum and Rr / Rp maximum and if not repetition of the phase (C).

7. Output bearing (26) for an output pinion (15) for a drive (1) for a concrete mixer vehicle (8), in which the output pinion (15) is provided with a flange (23) protruding outside with respect to the drive (1) for putting in rotation a drum (3) of the concrete mixer vehicle (8), the output bearing (26) being orientable and comprising an internal ring (40) and an external ring (39) coaxially placed with interposition of first sliding rollers (41 ) and second sliding rollers (42) in which each of the rollers (41 , 42) has a roller surface (Sr) curved and convex with curvature radius according to a roller radius (Rr), the output bearing (26) being configured for a free rotation condition between internal ring (40) and external ring (39) around a longitudinal axis (49) and for a free reciprocal inclination condition between the internal ring (40) and the external ring (39) with respect to the longitudinal axis (49) according to an inclination angle (30), a first series of first rollers (41) and a second series of second rollers (42) being placed between theinternal ring (40) and the external ring (39) on opposite sides of the output bearing (26) with respect to a central plane (P) of symmetry on which a symmetry transverse axis (34) lies, the external ring (39) comprising radially internally an external ring surface (Sae) which constitutes a sliding external rail for the first rollers (41) and the second rollers (42), the external ring surface (Sae) being a curved surface with curvature radius according to an external ring radius (Rae), the internal ring (40) comprising radially externally a first semi-surface (Sail) which constitutes a sliding internal rail for the first rollers (41) and a second semi-surface (Sai2) which constitutes a sliding internal rail for the second rollers (42), the first semi-surface (Sail) and the second semi-surface (Sai2) of the internal ring being curved surfaces with curvature radius according to an internal ring radius (Rai), the roller surface (Sr) constituting the contact surface of the rollers (41 , 42) with the external rolling rail consisting of the external ring surface (Sae) that is on the external ring (39), and with the internal rolling rail consisting of the first semisurface (Sail) and of the second semi-surface (Sai2) of the internal ring (40), characterised in that the output bearing (26) is configured for a rotational speed of the output pinion (15) between 0.1 rpm and 30 rpm and further characterised in that constructive parameters of the output bearing (26) including the ratio Rr / Rp between the roller radius (Rr) and a rail radius (Rp) which is indicative of corresponding values of the external ring radius (Rae) and of the internal ring radius (Rai), are calculated according to the method in accordance with an any of the previous claims 1 to 6.

8. Output bearing (26) according to the previous claim characterised in that the inclination angle (30) for reciprocal inclination between internal ring (40) and external ring (39) with respect to the transverse axis (34) is between -10 degrees and +10 degrees, preferably between -6 degrees and +6 degrees.

9. Drive (1) for drum (3) of concrete mixer vehicle (8) comprising an epicyclic multistage gearreduction device, in which the drive (1) includes a structural box (19) containing the components of the epicyclic gear-reducing device, the drive (1) being provided with a driving input (24) for connection of a drum motor (2) for driving a first stage of reduction stages of the epicyclic gear-reduction device, the drive (1 ) comprising an output flange (23) for the fixing to the drum (3) and for the transmission of rotational motion from the drum motor (2) to the drum (3), the output flange (23) comprising an output pinion (15) which is coupled with a last reduction stage of the epicyclic gear-reduction device of the drive (1), the output pinion (15) being supported inside the drive (1) by a respective bearing, characterised in that the support bearing of the output pinion (15) is an output bearing (26) configured for a rotational speed of the output pinion (15) between 0.1 rpm and 30 rpm and further characterised in that constructive parameters of the output bearing (26) including the ratio Rr / Rp between the roller radius (Rr) and a rail radius (Rp) which is indicative of corresponding values of the external ring radius (Rae) and of the internal ring radius (Rai), are calculated according to the method in accordance with an any of the previous claims 1 to 6.

10. Drive (1) for drum (3) of concrete mixer vehicle (8) according to the previous claim, characterised in that the external ring (39) of the output bearing (26) includes radially externally a first coupling surface (31) for mounting and coupling inside the drive (1), the drive (1) comprising a housing (46) located on the box (19), the output bearing (26) being placed inside the housing (46), the housing (46) having a circular shape delimited by a wall (47) for coupling with the first coupling surface (31 ) of the output bearing (26), in such a way that a condition is obtained for transmission of radial loads applied to the output bearing (26) from the output pinion (15) towards the box (19) of the drive (1).

11. Drive (1 ) for drum (3) of concrete mixer vehicle (8) according to any of the previous claims 9 to 10, characterised in that the output pinion (15) includes a first portion (16) for coupling with the output bearing (26), the internal ring (40) of the output bearing (26) comprising radially internally a second coupling surface (32) for mounting and coupling with the output pinion (15).

12. Drive (1) for drum (3) of concrete mixer vehicle (8) according to the previous claim, characterised in that the output pinion (15) includes a second portion (17) for coupling with a grooved coupling (38)shaped in such a way as to allow said inclination according to said inclination angle (30).

13. Drive (1) for drum (3) of concrete mixer vehicle (8) according to the previous claim, characterised in that the grooved coupling (38) includes: - a series of grooves (37) radially externally located on the second portion (17) of the output pinion (15) for rotational transmission of motion from the last stage of the drive (1) to the output pinion (15) and to the fixing flange (23) of the drum (3);- a series of teeth (36) radially internally located on a bell (21); the bell (21) constituting an interconnection and rotational motion transmission element between the last stage of the drive (1) and the second portion (17) of the output pinion (15) by means of the grooved coupling (38) with rotational motion transmission independent with respect to the value of the inclination angle (30).

14. Concrete mixer vehicle (8) comprising a mixing drum (3) characterised in thatIt includes a drive (1 ) for drum (3) of concrete mixer vehicle (8) according to any of the previous claims 9 to 13.

Images