GB2565581A

GB2565581A - Improvements relating to cams for engines

Info

Publication number: GB2565581A
Application number: GB1713211.9A
Authority: GB
Inventors: Art Komuroglu Ertugrul
Original assignee: Jaguar Land Rover Ltd
Current assignee: Jaguar Land Rover Ltd
Priority date: 2017-08-17
Filing date: 2017-08-17
Publication date: 2019-02-20
Anticipated expiration: 2037-08-17
Also published as: GB2565581B; GB201713211D0

Abstract

A cam member 12 for an engine comprising a circumferential outer surface 20 defined by a width profile extending for a length in a circumferential direction and comprising an opening flank region (B, Fig. 3), a closing flank region (D, Fig. 3), a heel region (A, Fig. 3) and a nose region (D, Fig. 3); The width profile of the cam surface 20 varies along its circumferential length such that the width at a selected point along the width profile corresponds to a predicted cam load profile; the cam member provides for a significant mass reduction without compromising performance; Aspects of the present invention relate to a cam member for a camshaft of an engine, a camshaft including such a cam member, and a vehicle including such a camshaft.

Description

IMPROVEMENTS RELATING TO CAMS FOR ENGINES

TECHNICAL FIELD

The present disclosure relates to a cam, for example as would be used as part of an engine camshaft to actuate combustion chamber inlet and outlet valves. The disclosure also relates to a method for optimising the design of such a cam. Aspects of the invention relate to a system, to a method, and to a vehicle.

BACKGROUND

Internal combustion engines use cams to actuate the inlet and exhaust valves of the combustion chambers. Cam design is crucial to combustion timing and combustion chamber gas flow, and a great deal of design effort goes into optimising the cam lift profile of the opening and closing flanks of cams in order to optimise the performance and economy of an internal combustion engine.

There is also a prevailing design goal to reduce the mass of moving parts of the engine, because these marginal gains contribute to overall improvements to engine efficiency. However, reducing mass of parts must be achieved without comprising performance.

It is against this background that the invention has been devised.

SUMMARY OF THE INVENTION

Aspects of the present invention relate to a cam member for a camshaft of an engine, a camshaft including such a cam member, and a vehicle including such a camshaft.

According to an aspect of the present invention there is provided a cam member for a camshaft of an engine, wherein the cam member comprises a circumferential outer surface, that surface being defined by a width profile extending in a circumferential length, and comprises an opening flank region, a closing flank region, a heel region and a nose region, wherein the width profile of the circumferential outer surface varies along its circumferential length such that the width at a selected point along the width profile corresponds to a predicted cam load profile.

Beneficially, the invention provides a cam member having a reduced mass compared to prior art cam members without being compromised on performance. It is currently envisaged that a mass saving of around 60% can be achieved, although this is based on one particular cam member and so is merely an indication of the mass-reduction advantage that the invention provides. The reduced mass of the camshaft provides an efficiency advantage for the associated engine because it reduces parasitic losses associated with driving the camshaft and operating the tappets. Further mass reductions can be realised by optimising associated valve train components such as the tappets, tappet springs, cam chains and so on.

In some embodiments, a width of the circumferential outer surface of the opening flank region may be greater than a width of the circumferential outer surface of the heel region. Furthermore, a width of the circumferential outer surface may taper outwardly from the heel region, through an opening ramp region to the opening flank region. A width of the circumferential outer surface of the opening flank region may taper inwardly towards the nose region.

In some embodiments, the width of the circumferential outer surface of the closing flank region may be greater than a width of the circumferential outer surface of the heel region. The width of the circumferential outer surface of the closing flank region may be greater than a width of the circumferential outer surface of the nose region. The closing ramp region may be located between the closing flank region and the heel region.

In embodiments, the width of the circumferential outer surface of the nose region may be substantially constant, and the width of the circumferential outer surface of the heel region may be substantially constant.

The minimum width of the circumferential outer surface of the heel region may be between 0.5 and 2mm.

In terms of material, in some embodiments, the cam member may be made from bearing steel, for example 100Cr6 grade steel.

Within the scope of this application it is expressly intended that the various aspects, embodiments, examples and alternatives set out in the preceding paragraphs, in the claims and/or in the following description and drawings, and in particular the individual features thereof, may be taken independently or in any combination. That is, all embodiments and/or features of any embodiment can be combined in any way and/or combination, unless such features are incompatible. The applicant reserves the right to change any originally filed claim or file any new claim accordingly, including the right to amend any originally filed claim to depend from and/or incorporate any feature of any other claim although not originally claimed in that manner.

BRIEF DESCRIPTION OF THE DRAWINGS

One or more embodiments of the invention will now be described, by way of example only, with reference to the accompanying drawings, in which:

Figure 1 is a side view of a vehicle, having an engine that is equipped with a camshaft in accordance with an embodiment of the invention, wherein the camshaft is shown enlarged in the inset panel;

Figure 2 is a perspective view of a section of the camshaft in Figure 1 which shows two cam members of that camshaft in more detail;

Figure 3 is an end view of the camshaft section in Figure 2, which shows clearly the cam lift profile of the cam member;

Figure 4 is a side view of one of the cam members in Figure 2, which shows clearly its width profile, whereas Figure 5 shows the other side of the cam member;

Figure 6 is a flow chart of a process for design and manufacture of an embodiment of a cam member according to the invention;

Figure 7 is a diagram depicting the contact mechanics of two cylinders with parallel axes, which theory underpins the design process for the cam member according to the invention;

Figure 8 is a graph illustrating an example of how pressure is applied to the cam member during use; and

Figure 9 illustrates an ‘unwrapped’ circumferential cam surface of an embodiment of a cam member according to the invention.

DETAILED DESCRIPTION

With reference to Figure 1, a vehicle 2 includes an internal combustion engine 4 positioned toward a front end 6 of the vehicle in an engine bay 8. As is known, the engine 4 comprises a set of pistons that reciprocate within respective cylinders (not shown). It will be appreciated that engines for road vehicles usually have between 4 and 8 cylinders and pistons, although there may be fewer or more in some engine configurations. Gas flow into and out of the cylinders is controlled by valves, the movement of which is governed by respective tappets (not shown), for example flat or rounded tappets, or roller tappets, which are themselves driven by a camshaft 10. The camshaft 10 is shown enlarged in Figure 1 in the inset panel.

As will be noted, the camshaft 10 includes a plurality of cam members 12, one for each of the tappets. The cam members 12 may be separate components that are fixed into position on the camshaft 10, for example by a suitable welding process, or they may be formed integrally with the camshaft during the casting or the subtractive machining process. Camshafts are typically made from cast iron or billet steel and, depending on the type of engine in which the camshaft will be used, it may have a mass of between

2-1 Okg, by way of example. It will be appreciated therefore that any measure to reduce the mass of the camshaft, and thus the mass of the broader drivetrain, is desirable.

As will be discussed in more detail below, the embodiments of the invention provide a technique for designing and manufacturing cam members so that their mass can be reduced without compromising performance. It is currently envisaged that a mass saving of around 60% can be achieved, although this is based on one particular cam member and so is merely an indication of the mass-reduction advantage that the invention provides. Applied to the camshaft of a typical 4-cylinder diesel engine, it is envisaged that mass reduction of over 150g is achievable. The reduced mass of the camshaft provides an efficiency advantage for the associated engine because it reduces parasitic losses associated with driving the camshaft and operating the tappets. Further mass reductions can be realised by optimising associated valve train components such as the tappets, tappet springs, cam chains and so on.

In a broad sense, the invention involves optimising the width profile of the cam member so that the width at a given point along the cam surface corresponds to the predicted load distribution or ‘load profile’ that will be applied by the tappets on the cam member in use. That is to say, that the width around the circumference of the cam surface is selected based on the loads to which the cam will be subject, in use. More specifically, the loading applied to the cam member by the tappet during a period of rotation varies greatly depending on whether the tappet is closed, which corresponds to tappet riding over the base circle of the cam member, or whether the tappet is being actuated, corresponding to the tappet riding over the rising or closing flank of the cam. There is thus an opportunity to reduce the width of the cam member over at least the base circle when loads are relatively low, such that the reduced width enables a lower mass component to be made.

Turning now to Figures 2 to 5, a cam member 12 according to an embodiment of the invention is shown in more detail. The cam member 12 includes a circumferentiallyextending outer cam surface 20 that defines the outer periphery of the cam member

12. The cam surface 20 is the surface over which the tappet rides in use. When viewed from its end, along its rotational axis, as in Figure 3, the cam surface 20 defines a cam lift profile 22 for the cam member 12. It will be appreciated that the cam lift profile can be considered to be conventional for the purposes of this discussion.

Remaining with Figure 3, the cam lift profile 22, and therefore also the cam surface, can be partitioned into portions or regions, based on the direction of rotation ‘R’ of the cam member. Note that each region is continuous with its adjacent region as the outer surface is smooth without any step-like discontinuities. The regions of the cam lift profile 22 are named below using conventional terminology in this technical field:

a base circle or heel ‘A’ is the region of the lift profile that is concentric with the camshaft and does not have any lift height. A tappet rolling over this portion of the cam member 12 is not actuated and so remains closed;

an opening flank region ‘B’ defines the main opening phase of the cam member 12 which cause high acceleration of an associated tappet. The opening flank B region extends from the base circle A towards a nose ‘O’. Note that the nose C is centred on the cam centreline ‘L’, has the smallest radius of curvature and defines the highest lift point of the cam;

a closing flank region ‘D’ is the portion of the lift profile which extends from the nose C towards the base circle A and provides a high closing acceleration for the tappet. The opening flank region B and the closing flank region D may be symmetrical or asymmetrical depending on the required combustion characteristics and gas flow;

in order to smooth the transition between the base circle and the opening and closing flanks, the cam surface 20 may also define opening and closing ramp regions Έ’, ‘F’, immediately adjacent the base circle A. These low velocity regions smooth the transition between the base circle and the flanks on both opening and closing movements in order to avoid lash between the tappet and the cam at the start and end of a lift event.

Whereas Figure 3 shows the lift profile, Figure 2 illustrates the width profile 24 of the cam surface 20. In the same way that the lift profile defines how the height of the cam member varies along the circumferential length of the cam surface, the width profile can be considered to define or map how the width of the outer cam surface varies along its length. It is important to note that the width of the cam surface 20 varies along its circumferential length in such a way that the width at any selected point along the cam surface 30 corresponds to the predicted loading, or load distribution, that the cam member 12 will experience, in use. Expressed another way, the width profile of the cam member 12 is optimised for the load distribution that the cam member will experience, in use, such that the cam surface 20 is thicker/wider in regions where the loads on the cam are higher, such as regions where the cam acts to accelerate or decelerate the tappet and where the valve driven by the cam is exposed to pressurised gas within an associated cylinder of the engine. Note at this point that the width profile has a variable width that extends along the circumferential length of the cam member and is oriented perpendicular to the plane of the cam member.

In more detail, and also with reference to Figures 4 and 5, the width profile of the cam surface 20 is most narrow at the base circle A of the cam member 12. The cam surface 20 then starts to taper outwardly or to diverge, that is, to increase in width, when progressing through the opening ramp E and the opening flank B. This is the widest point of the cam width profile 24 for two principle reasons: firstly, it represents the region which will impart the highest acceleration on its tappet during the opening stroke thereby experiencing the highest load values, and, secondly, combustion gases act strongly on the tappet at this point, in the case of an exhaust cam, thereby applying a load to the cam member opposing the tappet opening movement. Along the nose C and the falling flank/ramp D, F of the cam width profile 24, the width of the cam surface 20 gradually tapers inwardly, that is to reduce in width, together with the lift of the cam member 12, eventually tapering to the minimum width at the base circle A. As will be discussed in more detail below, although the theoretical minimum width of the cam surface may be thinner than what is shown in the Figures, for example around 0.5mm to 2mm, the actual minimum width may be defined by manufacturing practicalities and the material from which the cam member 12 is made. For example, it may be the case that the material type, and permitted lateral positional tolerances between the cam member and interacting components such as tappets may act as a limit to how thin the heel of the cam member can be formed.

The above discussion focusses on the shape of the cam member 12, and more specifically the shape of the cam width profile 24 about the circumference of the cam surface 20. The following discussion will explain how the cam width profile 24 is determined.

Figure 6 depicts an embodiment of a process 60 by which the cam width profile 24 of the cam member 12 may be determined. This procedure would generate a data set or model which would specify the cam width profile 24, and more broadly a dimensional specification for the cam member, from which model the cam member 12 may be manufactured by way of a suitable manufacturing process. It is envisaged that the most appropriate manufacturing processes would be sintering, forging, grinding, casting, or machining from a solid, followed by suitable finishing, although other subtractive manufacturing processes may also be used. Moreover, in some applications, particularly where the cam member may be formed from plastics, moulding or additive manufacturing may be suitable techniques.

Firstly, at step 62 the maximum cam loads are identified throughout the duty cycle of the cam member. This may be achieved by suitable simulation methods that would be familiar to a skilled person. For example, the load on the cam member may be determined at a range of engine rpm values up to the maximum engine speed. From this data it is possible to then determine the cam loads at sequential crank angle values (for example in increments of 1 or 2 degrees), based on various engine speeds, the known loading and cam profile data. An example of a suitable software application that would enable such calculations is CAMEO from AVL List GmbH. CAMEO is acknowledged as a registered trade mark of its respective owner.

Thus, a plurality of cam loads in respect of a corresponding plurality of angular positions for the cam member are determined, and this data set can be considered to be a load profile. Once the cam load profile has been determined, calculations are performed, at step 64, to determine the contact pressure distribution along the circumferential outer cam surface 20. The contact pressure between the cam member 12 and the tappet is believed to be an accurate way of analysing the stress exerted on the cam during operation, and is derived from Hertzian contact stress theory. Figure 7 depicts contact mechanics between two cylinders having parallel rotational axes. As the skilled person would be aware, contact mechanics covers the interaction and deformation of solid bodies in contact with each other and under load. Hertzian contact stress relates to localised stresses that are generated during contact of two curved surfaces, in this case cylinders, as those surfaces deform slightly under the applied loads. The model uses cylinders in this case because the lower cylinder represents the cam member and the upper cylinder represents a roller of a roller tappet, or possibly also the rounded end of a rounded tappet.

The contact pressure is illustrated by the elliptical region labelled as pmax that encircles the rectangular point of contact between the two cylinders.

The half-width b of the rectangular contact area of the two parallel cylinders is derivable from Hertzian contact stress theory as:

(1)

1-!.·

In equation (1), and as depicted in Figure 7, E1 and E2 are the modulus of elasticity of the two cylinders, and v1 and v2 are the respective Poisson’s ratios of the cylinders. L is the length of the contact area, F is the applied force (N) on the lower cylinder by the upper cylinder, and R1 and R2 are the radii of the two cylinders, respectively.

From the above equation, it is possible to derive an expression for the maximum contact pressure along the centre line of the rectangular contact area:

(2)

So, based on the known load applied between the cylinders at each angular position of the cam member 12, it is possible to calculate the maximum contact pressure on the cam member 12 at incremental positions along the cam surface 20. Note that the term ‘each angular position’ may be considered to mean, for example, the load applied at 1 “ increments of cam rotational position. However a higher or lower resolution may also be used.

Figure 8 is a graph that shows a possible result of performing these calculations. The trace labelled P is the maximum static contact pressure, whereas the trace labelled L is the cam lift of the cam member 12. The Y-axis of Figure 8 represents contact stress in Mpa, for the trace P, and also cam lift in mm, for the trace L. The X-axis represents angular rotational position for the cam member in terms of crank angle. Note that the scale for each of the traces is not shown here as the trace shapes are for illustrative purposes.

It will be appreciated therefore, that the stress on the cam member 12 increases rapidly at the point where the cam lift phase begins, i.e. the opening ramp E and opening flank B, and remains relatively high until the stress reduces quickly to base levels as the cam member rotates to the closing ramp point F.

Following determination of the maximum pressure throughout the rotation of the cam member 12, this data may then be processed in order to calculate the optimum width of the cam surface 20 based on the predicted maximum dynamic stress or pressure that the cam member 12 will experience in use. The calculations may be based on the calculated contact pressure and Hertzian stress limit of the material to determine the minimum width of the cam surface for each cam angle. This is illustrated in Figure 10, in which line W illustrates the optimum width profile in mm (scale on left hand Y-axis), line L is the lift profile of the cam member (scale on right hand Y-axis) and line T represents a typical value of cam width, which in this case is approximately 10mm.

By way of further explanation, the following equations may be applied for each selected rotational angle of the cam member:

Equations 1) and 2) discussed above provide the means to calculate the Hertzian static contact stress at each angle of the cam member.

Since the context is dynamic rather than static, the dynamic contract stress may be calculated by:

3)

P_dy = pmax* (Vl.5)

So, the calculated value of dynamic stress is valid for a predetermined cylinder length, which corresponds to a known or ‘prior art’ cam member width, which may be in the region of 10mm for example. To provide more context, as an example, a practical value for the dynamic stress may be calculated to be 1300MPa. It should be appreciated that the calculated dynamic stress may be significantly less than the contact stress limit for the material (e.g 100Cr6), which may be PJim = 1800MPa, by way of example. Even factoring in a safety limit, this means that the cam member may be reduced in width from what is typical whilst not exceeding the contact stress limit.

So, using the contact stress limit, the above equations can be worked through to calculate the minimum width of the cam member, that is the contact length, L, of the cylinders in the Hertzian model discussed above.

The maximum permissible static stress based on the contact stress limit can be calculated by rearranging equation 3).

4) p_max = p_dy I (Vl .5)

Having calculated the width b of the contact area between the two cylinders, then it becomes possible to calculate the length of that contact area to satisfy the maximum permitted static contact stress.

Since the revised value of p_max is known, from equation 4), it therefore becomes possible to combine equation 2) and equation 1) to result in the following:

5)

Since the material properties and cylinder radius values remain the same as discussed above, then equation 5) can be rearranged to provide a means for calculating a value for L, that is the minimum width of the outer surface of the cam member at any given cam angle that will not exceed the maximum dynamic contact stress for that component.

As can be seen from Figure 10, the theoretical optimum width profile that is calculated in the above procedure may not be ideal from a manufacturing perspective. For example, the calculated width profile W may be too erratic to enable the cam member to be manufactured efficiently, and the calculated minimum width may be too narrow to be reproduced in a practical embodiment of the cam member 12. So, the theoretical optimum width profile may be adapted in two ways, as is expressed at step 66 and 68.

Firstly, at step 66, a value for the minimum width of the cam surface 20 may be determined. As mentioned above, this may be based on the material from which the cam member 12 is to be made and what is practical from a manufacturing perspective, but also on information about how interaction with other components such as an associated tappet, for example how lateral positional tolerance of such tappets, influences the permissible minimum with of the heel region of the cam member.

Following the calculation of the minimum width, the optimised width profile can be further processed, for example by a smoothing process, at step 68. This could be achieved in numerous ways. One example is illustrated in Figure 10 by the line labelled ‘S’. Here, the line S is generated that covers the optimised width profile. The generation of the line S can be a manual process, for example as would be undertaken by a skilled operator, or it can be implemented by a suitable software algorithm. In Figure 10, the peak cam width points are noted and a line is generated to trace the peak width points, the resulting line encompassing the entire contact pressure trace.

Another example of an optimised and smoothed cam width profile 24 is shown in Figure 9. Note that in this figure, the width profile of the cam member has been ‘unwrapped’ or ‘unfolded’ from around the circumference of the cam member so as to instead extend in a straight line. As can be seen, the base circle A has a minimum width of 1mm, and the width increases, or tapers outwardly, to a maximum of 8mm at a cam lift height of 1,332mm. Following this, the width profile narrows gradually back to the minimum width at the base circle. It should be noted that in practice the transitions between the various segments on the profile illustrated in Figure 9 may be smoothed so that a tappet will ride over the cam surface without breaking contact.

Once the cam width profile 24 has been optimised and smoothed for manufacture, then a data file may be generated, at step 70, containing a suitable model from which the cam member 12 can be fabricated. The data file may thus be suitable for loading into a computer controlled (CNC) machining apparatus, for example. Such an apparatus would be suitable for machining, at step 72, a cam member from an input steel block. Another possible option is for the cam member 12 to be produced by way of an additive manufacturing process, although this may be more suitable for light load applications.

For completeness, Figure 11 shows an example of a system 100 in which the embodiments of the invention may be implemented. The system 100 includes a processing platform 102 on which the optimisation method 60 described above may be carried out. The processing platform 102 comprises a processor 104 that is in communication with a suitable RAM/ROM memory module 106. The processor 104 is also in communication with an input/output module 108 which may also incorporate suitable communication functionality. It will be appreciated that the processor 104 is therefore operable to carry out the optimisation method as described above and uses the memory module 106 for this purpose. Simulation data 107, as discussed above, may be imported into the processing platform 102 from an external source or may also be generated by a suitable software application running on the processor 104.

Any suitable computing platform could be used to implement the method, such as a laptop or desktop configuration. The processing platform 102 is communication with a manufacturing platform 108, such as a CNC manufacturing system but it should be appreciated that this could also be an additive manufacturing system. Although in some embodiments the processing platform 102 may be in direct communication with the manufacturing platform 108, it should be appreciated that this may not be the case and, instead, the processing platform 102 may be configured to output an optimised cam member data file to an external portable memory device or another computing device, either directly or wirelessly, for later transfer to the manufacturing platform 108.

Many modifications may be made to the above examples without departing from the scope of the present invention as defined in the accompanying claims.

For example, in the above discussion it will be noted that the width profile corresponds to the entire circumference of the outer surface of the cam member. However, it is envisaged that the process described above could also be carried out in respect for only a portion of the outer surface. For example, it may be considered that the opening flank region of the outer surface of the cam member may be optimised, although it is less important to optimise the closing flank region. The ‘width profile’ of the cam 10 member should therefore be interpreted with this in mind.

Claims

1. A cam member for a camshaft of an engine, wherein the cam member comprises a circumferential outer surface, that surface being defined by a width profile extending in a length in a circumferential direction, and comprises an opening flank region, a closing flank region, a heel region and a nose region, wherein the width profile of the circumferential outer surface varies along its circumferential length such that the width at a selected point along the width profile corresponds to a predicted cam load profile.

2. The cam member of Claim 1, wherein a width of the circumferential outer surface of the opening flank region is greater than a width of the circumferential outer surface of the heel region.

3. The cam member of Claim 2, wherein a width of the circumferential outer surface tapers outwardly from the heel region, through an opening ramp region to the opening flank region.

4. The cam member of any of Claims 1 to 3, wherein a width of the circumferential outer surface of the opening flank region tapers inwardly towards the nose region.

5. The cam member of any of Claims 1 to 4, wherein a width of the circumferential outer surface of the closing flank region is greater than a width of the circumferential outer surface of the heel region.

6. The cam member of any of Claims 1 to 5, wherein a width of the circumferential outer surface of the closing flank region is greater than a width of the circumferential outer surface of the nose region.

7. The cam member of any of Claims 1 to 6, wherein a closing ramp region is located between the closing flank region and the heel region.

8. The cam member of Claim 7, wherein a width of the circumferential outer surface of the closing flank region tapers inwardly towards the heel region.

9. The cam member of any of Claims 1 to 8, wherein the width of the circumferential outer surface of the nose region is substantially constant.

10. The cam member of any of Claims 1 to 9, wherein the width of the circumferential outer surface of the heel region is substantially constant.

11. The cam member of any of Claims 1 to 10, wherein the minimum width of the heel region is between 0.5 and 2mm.

12. The cam member of any of Claims 1 to 11, wherein the cam member is made from bearing steel.

13. The cam member of Claim 12, wherein the cam member is made from 100Cr6 grade steel.

14. A cam shaft comprising one or more cam members in accordance with any one of Claims 1 to 13.

15. A vehicle including a cam shaft in accordance with Claim 14.