Fastening system for a vehicle powertrain and method for constructing the same

Abstract

A driveline mounting system for a vehicle powertrain includes a first driveline mount exhibiting a first gradual rate of change in static stiffness under load. The mounting system also includes a second driveline mount exhibiting a second gradual rate of change in static stiffness under load. The first gradual rate of change in static stiffness is related to the second gradual rate of change in static stiffness by a torque balance equation for the static equilibrium of respective displacements at the first and second driveline mounts. The torque balance equation for static equilibrium is based in part on a spatial arrangement of the first and second driveline mounts relative to the driveline.A method for designing the powertrain mounting system determines the first and second gradual rates of change of the static stiffness under load to achieve efficient energy management.

Description

TECHNICAL FIELD

[0001] The present teachings generally comprise a method according to the preamble of claim 1 for constructing a fastening system for a vehicle drive train and a drive train fastening system according to the preamble of claim 6, as is essentially known from DE 24 34 633 A1.

[0002] For further information on the state of the art, please refer to the publication DE 40 09 995 A1 and the article by M. DAHL: The coupling-free, elastic mounting of engines using rubber elements. In: Automobiltechnische Zeitschrift, Vol. 71, 1969, No. 8, pp. 266-271. - ISSN 0001-2785. BACKGROUND

[0003] Vehicle powertrains typically include a power source, such as an engine, and include a transmission that delivers power from the power source to the wheels. A powertrain mounting system is used to attach the powertrain to vehicle structural components, such as an engine mount. The mounting elements may include resilient mounts, typically including a rubber portion. In some powertrain assemblies, the mounting system may include a brace that combines hydraulic damping and elastic deformation of a rubber member to respond to driveline torque loads. SUMMARY

[0004] The invention is based on the object of providing a driveline mounting system and a method for constructing such a driveline mounting system which uses a pitch specification for the static stiffness of a mount under dynamic loads and which is constructed using the pitch specification such that the rotational stiffness of the driveline mounting system is balanced.

[0005] This object is achieved according to the invention with a drive train fastening system having the features of claim 6 and with a method having the features of claim 1.

[0006] Further, a method for designing the driveline mount system is presented. The method includes selecting a first gradual rate of change of static stiffness under load for the first driveline mount. The first gradual rate of change of static stiffness of the first driveline mount includes a first average rate of change of static stiffness in a first range of dynamic loads and a second average rate of change of static stiffness in a second range of dynamic loads. The first average rate of change of static stiffness is less than the second average rate of change of stiffness, and the loads in the first range of dynamic loads are less than the loads in the second range of dynamic loads.Optionally, the first gradual rate of change of static stiffness may also include a third average rate of change of static stiffness that is greater than the second average rate of change of stiffness and that occurs in a third range of dynamic loads. The loads in the third range are greater than the loads in the second range.

[0007] The method includes determining a corresponding second gradual rate of change of static stiffness under load for the second driveline mount using the first gradual rate of change of static stiffness in a torque balance equation for the static balance of respective torques at at least the first and second driveline mounts based on the spatial arrangement of at least the first and second driveline mounts. According to the method, a driveline mount system is then provided, including the first driveline mount having the first gradual rate of change of static stiffness and the second driveline mount having the corresponding second gradual rate of change of static stiffness.Because the second gradual rate of change of the static stiffness of the second driveline mount is based on the torque balance equation, the rotational stiffness of the driveline mount system is balanced, resulting in balanced displacements to minimize movement at the driveline center of gravity.

[0008] The driveline mounting system provided according to the method manages the energy of driveline movement to reduce loads on the driveline mounts while limiting stiffness in the mounts to isolate noise and vibration, for example, in the first range of dynamic loads. More efficient (i.e., lighter) structures of the driveline, driveline mounts, and driveline support structure, such as the body or chassis, may therefore be possible. Improved driveline response to transient load input, for example, during a garage shift, and the use of specialized driveline mounting systems, such as the three-point pendulum mount system, may be possible over a wider range of dynamic loading.The components of a known driveline mounting system are designed (i.e., "tuned") with load deflection characteristics without a pitch specification as disclosed herein, and therefore cannot provide isolation of driveline noise and vibration in certain areas and restriction of driveline movement in others (e.g., during driveline pitch or roll, such as occurs during vehicle acceleration) in the same way as driveline mounting systems constructed according to the method disclosed herein. Mounting systems constructed with the pitch specification disclosed herein are more effective at regulating dynamic torque loads.The method is adaptable to different powertrains with different powertrain mounting systems according to the spatial arrangement of the powertrain mounts that respond to the powertrain torques.

[0009] The driveline mounting system constructed according to the method disclosed herein is efficient at regulating the driveline's kinetic energy due to oscillating transient torque load input (e.g., during a garage shift, engine start-stop, tap-in / take-off, etc.). The pitch specification of the stiffness of the first and second mounts creates pseudo-damping in the system due to the continuously changing resonant natural frequency of the driveline, without requiring the higher component content of a hydraulic damper.

[0010] The foregoing features and advantages, as well as other features and advantages of the present teachings, will become readily apparent from the following detailed description of the best modes for carrying out the present teachings when taken in conjunction with the accompanying drawings. BRIEF DESCRIPTION OF THE DRAWINGS Fig. 1 is a schematic illustration in fragmentary perspective view of a portion of a first embodiment of a vehicle having a first embodiment of a powertrain (shown in phantom) including a first embodiment of a powertrain mounting system. Fig. Figure 2 is a representative graph of stiffness in Newtons per millimeter (N / mm) versus force in Newtons (N) for a first driveline mount of the driveline mount system of Fig. 1. Fig.3 is a schematic diagram of a torque balance for the driveline mounting system of Fig. 1. Fig. 4 is a schematic illustration in fragmentary perspective view of a portion of a second embodiment of a vehicle having a second embodiment of a powertrain (shown in phantom) including a second embodiment of a powertrain attachment system. Fig. 5. is a schematic diagram of a torque balance for the drivetrain mounting system of Fig. 4. Fig. 6 is a schematic diagram in fragmentary perspective view of a portion of a third embodiment of a vehicle having a third embodiment of a powertrain (shown in phantom) including a third embodiment of a powertrain attachment system. Fig.Figure 7 is a schematic diagram of a torque balance for the driveline mounting system of Fig. 6. Fig. 8 is a flowchart of a method for constructing the driveline mounting systems of Fig. 1 - 7. Fig. 9 is a schematic top view of the first driveline mount of the driveline mount system of Fig. 1. DETAILED DESCRIPTION

[0011] With reference to the drawings, in which like reference numerals refer to like components, Fig.1 shows a portion of a vehicle 10 having a powertrain 12, shown schematically in phantom. In the embodiment shown, the powertrain 12 includes an engine 14 drivingly connected to a transmission 16. In other embodiments, the powertrain may not include an engine or transmission. For example, the powertrain may utilize an electric motor instead of or in addition to a powertrain. A powertrain mounting system 18 supports the powertrain 12 relative to a load-bearing vehicle support structure, such as the vehicle body 19 and the vehicle chassis, which includes an engine mount 20. The powertrain 12 is a transversely oriented powertrain, and the powertrain mounting system 18 is a three-point pendulum mount system.

[0012] The powertrain mounting system 18 includes a first powertrain mount 22, also referred to herein as a rear torque support 22. The powertrain mounting system 18 also includes a second powertrain mount 24, also referred to herein as a first side transmission mount 24 or a left-side transmission mount 24. The powertrain mounting system 18 includes a second side engine mount 26, also referred to herein as a right-side engine mount 26. The rear torque support 22, the left-side transmission mount 24, and the right-side engine mount 26 are arranged in a spatial arrangement 28 relative to each other and to a center of gravity CG of the powertrain 12. More specifically, the powertrain 12 has a torque roll axis T1 passing through the center of gravity CG.A virtual line V1 passes through the elastic centers EC1 and EC2 of elastomeric sections 27, 29 of the left-hand transmission mount 24 and the right-hand engine mount 26. Another virtual line V2 passes through the elastic centers EC3 and EC4 of elastomeric sections of a bushing 32 (shown in . Fig. 9 and discussed with reference thereto) and an elastomeric portion 33 of a bushing 34 of the rear torque support 22. A distance S1 extends from the virtual line V1 to the torque roll axis T1, and a distance S2 extends from the virtual line V2 to the torque roll axis T1. Both distances S1 and S2 are measured along a vertical line (i.e., along the Z-axis) intersecting the virtual lines V1, V2.

[0013] The driveline attachment system 18 is constructed according to a method 500 illustrated as a flow chart in Fig. 8 and in detail related to Fig. 8. According to the method 500, the first driveline mount, i.e., the rear torque arm 22, is designed with a "stiffness slope" specification. As used herein, a "stiffness slope" specification means a gradual rate of change of static stiffness in a given direction under a dynamic load. More specifically, and as shown in the diagram 100 of Fig.2, the rear torque arm 22 exhibits a "gradual rate of change of static stiffness" under load, which is specified by the average rates of static stiffness 102 in Newtons per millimeter (N / mm) per Newton in various zones of dynamic loading, where the static load 104 is measured in Newtons (N). The different load zones are associated with the dynamic load resulting from the torque of the powertrain 12 about the vehicle's Y-axis. Load zones Z1 and ZA include such events as driving the vehicle at cruising speed; low and high acceleration maneuvers; idling the vehicle on a level or inclined surface; and also transient conditions of the powertrain 12, which include start / stop events, engage / disengage events, garage shifts, and a tap or release of the accelerator pedal.Load zones Z2 and ZB include events such as acceleration or deceleration on rough roads, acceleration or deceleration on slippery or gravel roads, or similar events that cause excessive oscillation of the vehicle's drive axle. Load zones Z3 and ZC include events such as aggressive powertrain operation at high engine speeds and / or during abrupt transmission gear shifting.

[0014] Table I lists rates of change of the static stiffness (ie the mean gradient of stiffness in (N / mm) / N) of the first driveline attachment 22 according to diagram 102 of Fig. 2 in different load regulation zones, which are described by the associated forces in Newton (N) on the first drive train attachment 22. The numerical values ​​of Fig.2 and Table I are merely non-limiting examples, and other suitable gradual rates of change in stiffness and also other load zones (i.e., ranges of dynamic loads) may be selected within the scope of the present teachings, as long as the selected load regulation zones and gradual rates of change in stiffness fall within the predetermined limits listed in Table I to ensure that the powertrain mounting system 18 provides efficient energy management. It should be understood that the powertrain mounting systems described herein are merely non-limiting examples.The stiffness gradient specification and method of designing a driveline mounting system disclosed herein can be readily applied to driveline mounting systems with mounts having different spatial arrangements than those shown herein. The values ​​shown in Table I apply equally to all such transversely oriented driveline mounting systems and also to the transversely oriented driveline mounting systems specifically described herein (i.e., driveline mounting systems 18 and 318). The values ​​for longitudinally oriented driveline mounting systems can be selected within the scope of the present teachings, as long as the selected load regulation zones and gradual rates of change in stiffness fall within the predetermined limits listed in Table I. Table I Load regulation zones Force (N) Mean gradient ((N / mm) / N) (also called mean rate of change of static stiffness under load) ZC - extreme loads in reverse direction < maximum predetermined high cycle load (e.g. -7000) ≤ |-0.50| [as small as possible without exceeding a predetermined total displacement in the fastening] ZB - moderate loads in reverse direction Maximum predetermined high cycle load (e.g. -7000) to< maximum comfort load 0.35 [greater than the ZA rate and smaller than the ZC rate, as necessary for a smooth transition between the end of the ZA range and the beginning of the ZC range] ZA - Reverse drive isolation zone Maximum comfort load up to 0 -0.25 [as large as possible without exceeding one or more predetermined vehicle insulation requirements, as determined by any other suitable method] Z1 - Forward drive isolation zone 0 to maximum comfort load (e.g. 1st gear with wide open throttle) (e.g. 7000) 0.15 [as large as possible without exceeding one or more predetermined vehicle insulation requirements, as determined by any other suitable method] Z2 - moderate loads Maximum comfort load to maximum predetermined high cycle load (e.g. 14,500) 0.25 [greater than the Z1 rate and smaller than the Z3 rate, as necessary for a smooth transition between the end of the Z1 range and the beginning of the Z3 range] Z3 - extreme loads > maximum predetermined high cycle load (e.g. 14,500) ≤ 0.50 [as low as possible without exceeding a predetermined total displacement in the fastening]

[0015] Fig. 2 shows that the rear torque arm 22 has three different mean rates of increase in static stiffness under load due to a positive torque generated in loads P 1x and P 2x in three corresponding load zones. The loading is in the fore-aft direction (the longitudinal X-axis of the vehicle 10) on the driveline mounts 22, 24, 26 and is due to the dynamic torque of the driveline 12 about the Y-axis of the vehicle for a positive torque load on the driveline 12, which is a torque in a clockwise direction when viewed in Fig. 1 along the positive Y-axis. The driveline mounts 24, 26 react together to a first axial load P 1x of the drive train 12 at a distance S1 from the torque roll axis T1. The rear torque support 22 reacts to a second axial load P2x of the drivetrain 12 at a distance S2 from the torque roll axis T1. The rear torque arm 22 is designed with a first average rate of change ΔK1 of static stiffness in a first range Z1 of dynamic loads from 0 to 7000 N, with a second average rate of change ΔK2 of static stiffness in a second range Z2 of dynamic loads from 7000 N to 14,500 N, and with a third average rate of change ΔK3 of static stiffness in a third range Z3 of dynamic loads greater than 14,500 N. It is noted that the actual rates of change of static stiffness in each of the load zones Z1, Z2, Z3 are non-linear, and that the average rates of change ΔK1, ΔK2, and ΔK3 of static stiffness are linear averages of the rates of change in these zones.

[0016] Additionally, Fig.2, that the rear torque support 22 exhibits three different mean rates of increase in static stiffness under load due to a negative torque (ie, acting at the expense of P 1x and P 2x in the opposite direction to that shown) in three respective load zones ZA, ZB, ZC, wherein the loading in the fore-aft direction (the longitudinal X-axis of the vehicle 10) on the driveline attachments 22, 24, 26 is due to a dynamic torque of the driveline 12 about the Y-axis T1 for a negative torque loading of the driveline 12, which is a torque in a counterclockwise direction when viewed in Fig. 1 along the positive Y-axis. The rear torque support 22 is provided with a first average rate of change ΔK Aof the static stiffness in a first range ZA of dynamic loads from 0 to -600 N, with a second mean rate of change ΔK B the static stiffness in a second range ZB of dynamic loads < -600 N to -7000 N with an absolute value greater than the loads in the first range of dynamic loads and with a third mean rate of change ΔK C of the static stiffness in a third range ZC of dynamic loads greater than -7000 N and an absolute value greater than in the second range of dynamic loads. The actual rate of change of the static stiffness in each of the load zones ZA, ZB, ZC is non-linear, and the mean rates of change ΔK A , ΔK B and ΔK C of static stiffness are linear averages of the rates of change in these zones.

[0017] As it is in Fig.2, the mean rate of change of static stiffness under dynamic load increases as the absolute value of the dynamic load increases, because the mean rate of change ΔK1 of the static stiffness is smaller than the second mean rate of change ΔK2 of the static stiffness, which is smaller than the third mean rate of change ΔK3 of the static stiffness. In addition, the absolute value of the first mean rate of change ΔK A the static stiffness for a torque in the counterclockwise direction is smaller than the absolute value of the second mean rate of change ΔK B the static stiffness, which is smaller than the absolute value of the third mean rate of change ΔK Cof static stiffness. The first range Z1 of dynamic loads is referred to as a "comfort load zone" because the stiffness 102 and the first mean rate of change ΔK1 of static stiffness are specifically selected to be relatively low under relatively low dynamic loads for optimal occupant comfort. The comfort zone may also include the first range ZA for dynamic loads due to a torque in the opposite direction. The second range Z2 of dynamic loads includes moderate loads that are experienced relatively frequently. The third range Z3 of dynamic loads includes even relatively extreme loads. Additionally, the ranges ZB and ZC of dynamic loads are considered load regulation zones with relatively higher, but limited, mean rates of change of static stiffness.However, the stiffness 102 and the second average rate of change ΔK2 of the static stiffness of the rear torque arm 22 in the second load zone Z2 and also the stiffness 102 and the third average rate of change ΔK3 of the static stiffness of the rear torque arm 22 in the third load zone Z3 are gradually higher than the first average rate of change ΔK1 of the stiffness in the comfort load zone to enable the driveline mounting system 18 to regulate the kinetic energy of the driveline 12 and thereby reduce the load on the driveline mounts 22, 24, 26 while limiting the overall displacement in the mount.

[0018] The rear torque arm 22 can be designed to provide gradual rates of change in static stiffness in many different ways, such as by creating the shape, arrangement, size and material properties of a main spring element 31C and stop elastomer elements 31A, 31B, 31D and 31E of the torque arm 22, which in Fig.9. The elastomeric portion 33 of the bushing 34 is disposed between and extends between the mounting structure 35 and a metal insert 37. The elastomeric stop members 31A and 31E are attached to the mounting structure 35 and face respective elastomeric stop members 31B, 31D attached to a metal insert 39. The elastomeric main spring member 31C is attached to and extends between the metal insert 39 and the mounting structure 35. This arrangement of the torque arm 22 is provided merely by way of example, and other arrangements are possible in which the bushings 34 and 32 are variably attached to driveline and chassis structures.

[0019] Now on Fig.8, according to the method 500, at step 502, a driveline mounting system is selected having a number and spatial arrangement of the mounts relative to the driveline. Next, at step 504, a pitch specification of a first of the driveline mounts is selected. In other words, at step 504, a first gradual rate of change of static stiffness under a dynamic load is selected for the first of the driveline mounts. Which of the driveline mounts serves as the first driveline mount with the selected first gradual rate of change of static stiffness depends on the driveline mounting system selected at step 502. For example, in the pendulum mounting system 18 of Fig.1, the rear torque arm 22 is the first driveline mount, and the first gradual rate of change of the static stiffness includes the gradual rates of change ΔK1, ΔK2, and ΔK3 of the static stiffness. The first gradual rates of change of the static stiffness also include the gradual mean rates of change of the static stiffness ΔK A , ΔK B and ΔK C .

[0020] The method 500 then includes step 506, in which a corresponding gradual rate of change of the static stiffness under dynamic load is determined for at least a second driveline mount of the driveline system. Which driveline mount is considered the second driveline mount depends on the driveline mounting system selected in step 502. For example, if a pendulum driveline mounting system 18 is selected, the second driveline mount is either the left-side transmission mount 24 or the right-side engine mount 26. For purposes of discussion, the left-side transmission mount 24 is Fig.1 is considered the second driveline mount for the pendulum driveline mount system 18. As discussed herein, the combined X-stiffness of the left-side transmission mount 24 and the right-side engine mount 26 balances the X-stiffness of the rear torque arm 22.

[0021] At step 506, the second gradual rate of change of the combined under-load stiffness of the left-side transmission mount 24 and the right-side engine mount 26 is determined using the selected first gradual rate of change of the static stiffness in a torque balance equation for the static balance of respective torques at at least the first and second driveline mounts 22, 24 based on the spatial arrangement of at least the first and second driveline mounts 22, 24 relative to the driveline 12. Individual static stiffnesses of the left-side transmission mount 24 and the right-side engine mount 26 may be determined from the combined static stiffness of the left-side transmission mount 24 and the right-side engine mount 26. Fig.Figure 3 shows the ideal condition where the center of gravity CG of the drive train lies in the vertical plane of the virtual line V2. Fig. 3 shows a diagram 200 of the force P 1x acting on the combined load-bearing mounts (i.e., the left-side gearbox mount 24 and the right-side engine mount 26 together) at a distance S1 of the virtual line V1 from the torque roll axis T1, as described above, and the force P 2x , which acts on the torque support 22 at the distance S2 of the virtual line V2 from the torque roll axis T1, as described above. The torque T in a clockwise direction along the Y-axis causes the force P 1x and the force P 2x which are in constant static equilibrium, and therefore P 1x with respect to the amount equal to P 2xIn the present teachings, the combined stiffness K1 is made proportional to the axial stiffness K2 of the torque arm by a constant factor (S2 / S1). Therefore, the following relationship between the axial static stiffness K1 and K2 results: K1=K2(S2 / S1); where K1 is the combined axial static stiffness of the load-bearing mounts 24, 26 in the X direction at a distance S1 for a given driveline 12 and K2 is the axial static stiffness of the rear torque arm 22 in the X direction at the distance S2 for the same given driveline 12.

[0022] As it is in Fig. 3 is the displacement D1 corresponding to the force P 1x corresponds to P 1x / K1 (ie D1 = P 1x / K1), and the displacement D2, which corresponds to the force P 2x corresponds to is equal to P 2x / K2 (D2 = P 2x / K2). This results in D1 = D2 (S1 / S2). In other words, the movement of the drive train 12 under the torque T occurs such that the center of rotation is always at or near the center of gravity CG of the drive train 12. Avoiding translation at the center of gravity CG of the drive train 12 directly leads to a reduction in the dynamic force on the attachments 22, 24, 26 in the forward / reverse direction.

[0023] The representation of the three-dimensional spatial arrangement of the drive train fastening system 18 in Fig. 1 and diagram 200 of the torque balance equation in Fig.3 are simplifications of the physics of the drivetrain mounting system 18. A more specific torque balance equation, which includes a dynamic force balance, and / or other simplifications can be used instead to determine S1, S2 and the corresponding relationship between the static stiffnesses K1, K2. For example, in Fig. 1, the center of gravity CG is on the torque roll axis T1, but it is shifted from the XZ plane by the rear torque support 22 (i.e., from the plane encompassing the virtual line V2 and extending vertically). For simplification, the torque roll axis T1 can instead be treated as a horizontal line in the Y direction through the center of gravity CG, and the distances S1, S2 can be calculated perpendicular to V1 and V2 and by such a horizontal equivalent line. The two-dimensional torque balance diagram 200 of Fig.3 uses this simplification, in which the torque roll axis T1 is treated as a horizontal line through the center of gravity CG. The displacement of the center of gravity CG with respect to the XZ plane through the virtual line V1 is relatively small, therefore any error in the distances S1 and S2 as well as in the associated calculated stiffnesses K1, K2 resulting from this simplification will be relatively small, and can be corrected for any event by system simulation and measurements on the vehicle. Another simplifying assumption used in the torque balance diagram 200 of Fig.3 is that the stiffness K2 of the rear torque arm 22 is along the X-axis of the vehicle 10. The stiffness K2 may instead be along an axis of the rear torque arm 22 that has at least a component in the X-direction, but may not be completely horizontal.

[0024] The stiffness K2 of the rear torque support 22 in the above formula, which is also given as stiffness 102 (in Fig.2), varies with the load 104 and with the first gradual rate of change of the static stiffness under load for the rear torque arm 22 (i.e., that the rear torque arm 22 has the first pitch specification). Once the first gradual rate of change of the static stiffness under load for the rear torque arm 22 is selected at step 504, a corresponding combined static stiffness K1 of the load-bearing mounts (the left-side transmission mount 24 and the right-side engine mount 26) is determined from the static balance diagram 200 of Fig.3. The corresponding static stiffness of the left-side transmission mount 24 is then known according to a selected relationship of the left-side transmission mount 24 and the right-side engine mount 26. Therefore, if the left-side transmission mount 24 and the right-side engine mount 26 are each required to respond equally to the torque load T, for example, each of them can be designed to have half the combined static stiffness K1 (i.e., K1 / 2) resulting from the above torque balance equation. A second gradual rate of change of the static stiffness under load for the left-side transmission mount 24 over the same load zones of Fig. 2 is then necessarily related to the mean rates of change ΔK1, ΔK2 and ΔK3 of the static stiffness according to the scaling of K1 / 2 for each stiffness value 102, which in Fig.2 is shown.

[0025] Once the first gradual rate of change of static stiffness under load for the first driveline mount (e.g., the rear torque arm 22) is selected at step 504 and the corresponding second gradual rate of change of static stiffness under dynamic load for the second driveline mount (e.g., the left-side transmission mount 24) is determined at step 506, the driveline mount system 18 with the rear torque arm 22 exhibiting the first gradual rate of change of static stiffness under load and with the left-side transmission mount 24 exhibiting the second corresponding gradual rate of change of static stiffness under dynamic load may be provided at step 508 with the torque arm 22, the left-side transmission mount 24, and the right-side engine mount 26 having the selected spatial arrangement as related to Fig. 1 and Fig. 3. The driveline mounting system 18 may be positioned in the vehicle 10 at step 510 according to the spatial arrangement described with respect to Fig. 1 and Fig. 3. It will be appreciated that the stiffness of the driveline mounts 22, 24, 26 in other directions, such as along a Y-axis (laterally within the vehicle 10) or along a Z-axis (up and down relative to the fore-aft X-axis), need not be graded according to method 500 and may be determined according to any suitable method.

[0026] Fig.4 shows a vehicle 310 with a powertrain 312. In the embodiment shown, the powertrain 312 includes an engine 314 and a transmission 316 (shown only schematically in phantom) attached by a powertrain mounting system 318 to a load-bearing vehicle support structure, such as the vehicle body 319 and the vehicle chassis, which includes an engine mount 320. In other embodiments, the powertrain may not include an engine or transmission. For example, the powertrain may use an electric motor instead of, or in addition to, an engine. The powertrain mounting system 318 includes a rear torque reaction mount 322, a front torque reaction mount 323, a left-side transmission mount 324, and a right-side engine mount 326.The driveline 312 is a transversely oriented driveline, and the driveline mounts 322, 323, 324, 326 are arranged in a spatial arrangement 328 which is considered to be a four-point driveline mounting system 318 with neutral torque axles.

[0027] The driveline attachment system 318 may be configured according to the method 500 of Fig. 8. For example, at step 502, the four-point driveline attachment system 318 with neutral torque axles is selected. At step 504, the rear torque arm 32 is selected as the first driveline attachment, having a Z-direction stiffness gradient specification with a first gradual rate of change of static stiffness under dynamic loading. The selected dynamic load ranges may be the same as the load zones Z1, Z2, Z3 and ZA, ZB, ZC of Fig.2 and Table I or different from them. The mean rates of change of the axial static stiffness of the rear torque mount 322 may be the same as those described in Fig. 2, or other numerical values ​​may be selected as long as the average rate of change of the static stiffness of the rear torque mount 322 increases as the absolute value of the magnitude of the ranges of loads in the load zones increases and the predetermined limitations of Table I are met.

[0028] Fig. Figure 5 shows a torque balance diagram 200A, which simplifies the axial load P 2Z in the positive Z-direction at the rear torque reaction mount 322 and the axial load P 1Zin the negative Z-direction at the front torque reaction mount 323 under dynamic loads due to a drive axle torque of the powertrain 312 in a clockwise direction about the Y-axis of the vehicle 310. With reference to Fig. 4, S1 is the distance between the elastic center EC6 of the front torque reaction mount 323 and the torque roll axis T2 in a vertical plane extending through the elastic centers EC5, EC6 of the rear torque reaction mount 322 and the front torque reaction mount 323, respectively. S2 is the distance between the elastic center EC5 of the rear torque reaction mount 322 and the torque roll axis T2 in the same vertical plane extending through the elastic centers EC5, EC6. The diagram 200A of Fig.Figure 5 represents this arrangement. Alternatively, the distances S1 and S2 can be measured in a plane through the elastic centers EC5, EC6 of the front and rear torque reaction mounts 323, 322, respectively, but perpendicular to the torque roll axis T2 instead of in a vertical plane.

[0029] As stated with reference to Fig. 1 and Fig. 2, the center of gravity CG of the drive train mounting system 318 can be Fig. 4 may actually be shifted from the vertical plane. For simplicity, the torque roll axis T2 can instead be treated as a horizontal line in the Y direction through the center of gravity CG. The two-dimensional torque balance diagram 200A of Fig.5 uses this simplification, in which the torque roll axis T2 is treated as a horizontal line through the center of gravity CG. The displacement of the center of gravity CG from the vertical plane by the elastic centers EC5, EC6 is relatively small, and therefore any error in the distances S1 and S2, as well as in the calculated stiffnesses K1, K2 associated with them, resulting from this simplification will be relatively small and can be corrected for any event by system simulation or measurements on the vehicle. Another simplifying assumption used in the torque balance diagram 200A of Fig.5 is that the stiffnesses K1 and K2 run in the direction of the vertical Z-axis. The stiffnesses K1, K2 can instead be considered along the local Z-axis of a respective corresponding mount 323, 322, which axis is commonly referred to as the "focal angle" and is not perfectly aligned with the vehicle's Z-axis. Any error associated with this simplification can, in turn, be corrected through system simulation and on-vehicle measurements.

[0030] At step 506, the second gradual rate of change of the static stiffness K1 under a dynamic load for the front torque reaction mount 323 may then be calculated using the simplified torque balance equation from the torque balance diagram of Fig. 5. In particular, K1=K2(S2 / S1); where K1 is the static stiffness in the Z direction of the front torque reaction mount 323 at a distance S1 for a given dynamic load of the drivetrain 312 and K2 is the static stiffness in the Z direction of the rear torque reaction mount 322 at a distance S2 for the same dynamic load of the drivetrain 312. As shown in Fig. 5 is the displacement D1 which corresponds to the force P 1z corresponds to P 1z / K1 (ie D1 = P 1z / K1), and the displacement D2, which corresponds to the force P 2z corresponds to is equal to P 2z / K2 (D2 = P 2z / K2). This results in D1 = D2 (S1 / S2); in other words, the movement of the drive train 312 under the torque T is such that the center of rotation is always at or near the center of gravity CG of the drive train 312. Avoiding translation at the center of gravity CG of the drive train 312 directly leads to a reduction in the dynamic force on the attachments 322, 323, 324, and 326 in the vertical direction.

[0031] At step 508, the powertrain mounting system 318 is provided and installed in the vehicle 310 at step 510. The stiffness of the mounts 322 and 323 in directions other than the Z direction, as well as the stiffness of the mounts 324, 326, may be determined by any suitable method.

[0032] Fig.6 shows a vehicle 410 with a powertrain 412. In the embodiment shown, the powertrain 412 includes an engine 414 and a transmission 416 (shown schematically in phantom) attached by a powertrain attachment system 418 to a load-bearing vehicle support structure, such as the vehicle body 419 and / or the vehicle chassis 420. The engine 414 is shown attached to the vehicle chassis 420, which includes an engine mount. The transmission 416 is shown attached to the body 419. In other embodiments, the powertrain may not include an engine, and it may not include a transmission. For example, the powertrain may use an electric motor instead of, or in addition to, an engine.The powertrain mounting system 418 includes a plurality of powertrain mounts, such as an engine mount 422 located on a first side (e.g., the left side), an engine mount 424 located on a second side (e.g., the right side), and a rear transmission mount 426. The left-side engine mount 422 and the right-side engine mount 424 are supported by the chassis 420. In a passenger car application, the rear transmission mount 426 is supported by the body 419, as shown. In a truck application, the rear transmission mount 426 is supported by the frame structure of the chassis 420.The driveline 412 is a longitudinally oriented driveline, and the driveline mounts 422, 424, 426 are arranged in a spatial arrangement 428 which is considered to be a longitudinally oriented driveline mount system 418.

[0033] The driveline attachment system 418 may be configured according to the method 500 of Fig. 8. For example, at step 502, the longitudinally oriented powertrain mount system 418 is selected. At step 504, the left-side engine mount 422 may be selected as the first powertrain mount having a Z-direction stiffness pitch specification with a first gradual rate of change in static stiffness under dynamic loading. The selected dynamic load ranges may be the same as load zones Z1, Z2, Z3 and ZA, ZB, ZC of Fig.2 and Table I or different from them. The mean rate of change of the axial static stiffness of the left-side engine mount 422 may be the same as that shown in Fig. 2, or other numerical values ​​may be selected as long as the average rate of change of the axial static stiffness of the left-side engine mount 422 increases as the absolute value of the magnitude of the ranges of loads in the load zones increases and the predetermined limitations of Table I are met. Alternatively, the right-side engine mount 424 may instead be selected as the first engine mount with a first stiffness pitch specification.

[0034] Fig. Figure 7 shows a torque balance diagram 200B, which is a simplification of the force P AZ in the Z-direction on the left-hand engine mounting 422 and the force P BZin the Z-direction on the right-side engine mount 424 under dynamic loads due to a drive axle torque of the powertrain 412 in a counterclockwise direction about the X-axis of the vehicle 410. With reference to Fig. 6, S1 is the distance between the elastic center EC8 of the right-side engine mount 424 and the torque roll axis T3 in a vertical plane extending through the elastic centers EC7, EC8 of the left-side and right-side engine mounts 422, 424, respectively. S2 is the distance between the elastic center EC7 of the left-side engine mount 422 and the torque roll axis T3 in the same vertical plane extending through the elastic centers EC7, EC8. The diagram 200B of Fig.7 illustrates this arrangement. Alternatively, the distances S1 and S2 may be measured in a plane through the elastic centers EC7, EC8 of the left and right side engine mounts 422, 424, respectively, but perpendicular to the torque roll axis T3 instead of in a vertical plane.

[0035] As stated with reference to Fig. 1 and Fig. 2, the torque roll axis T3 is treated as a horizontal line in the X-direction through the center of gravity CG. The two-dimensional torque balance diagram 200B of Fig.7 uses this simplification, in which the torque roll axis T3 is treated as a horizontal line through the center of gravity CG. The displacement of the torque roll axis T3 from a horizontal line is relatively small, and therefore any error in the distance S1 and S2 and in the associated calculated stiffnesses K1, K2 resulting from this simplification will be relatively small, and can be corrected for any event by system simulation and measurements on the vehicle. Another simplifying assumption used in the torque balance diagram 200B of Fig.7 is that the stiffnesses K1 and K2 are in the direction of the vehicle's Z-axis. The stiffnesses K1, K2 can instead be considered along the local Z-axis of each corresponding mount 422, 424, which axis is commonly referred to as the "focal angle" and is not perfectly aligned with the vehicle's Z-axis. Again, any error associated with this simplification can be corrected through system simulation or through measurements on the vehicle.

[0036] At step 506, the second gradual rate of change of the static stiffness K1 under a dynamic load for the right-side engine mount 424 may then be calculated using the simplified torque balance equation from the torque balance diagram of Fig. 7. In particular, K1=K2(S2 / S1); where K1 is the static stiffness in the Z direction of the right-side engine mount 424 at a distance S1 for a given dynamic load of the drivetrain 412, and K2 is the static stiffness in the Z direction at the distance of the left-side engine mount 422 at a distance S2 for the same dynamic load of the drivetrain 412. As shown in Fig. 7 is the displacement D1 which corresponds to the force P Bz corresponds to P Bz / K1 (ie D1 = P Bz / K1), and the displacement D2, which corresponds to the force P Az corresponds to is equal to P Az / K2 (D2 = P Az / K2). This results in D1 = D2 (S1 / S2); in other words, the movement of the drive train 412 under the torque T is such that the center of rotation is always at or near the center of gravity CG of the drive train 412. Avoiding translation at the center of gravity CG of the drive train 412 directly leads to a reduction in the dynamic force on the attachments 422, 424, and 426 in the vertical direction.

[0037] At step 508, the powertrain mounting system 418 is provided and installed in the vehicle 410 at step 510. The stiffness of the mounts 422 and 424 in directions other than the Z direction and the stiffness of the mounts 426 may be determined by any suitable method.

Claims

[1] A method of constructing a driveline mounting system (18) for a vehicle driveline (12), the driveline mounting system (18) having a plurality of driveline mounts (22, 24) including a first driveline mount (22) and a second driveline mount (24), the driveline mounts (22, 24) having a spatial arrangement relative to the driveline (12); the method comprising: selecting a first gradual rate of change of static stiffness under load for the first driveline mount (22); a second gradual rate of change of static stiffness under load for the second driveline mount (24) is determined using the selected first gradual rate of change of static stiffness in a torque balance equation for the static equilibrium of respective torques at least on the first and second driveline mounts (22, 24) based on the spatial arrangement of at least the first and second driveline mounts (24) relative to the driveline (12); and the driveline mounting system (18) is provided with the first driveline mount (22) exhibiting the first gradual rate of change of static stiffness under load and with the second driveline mount (24) exhibiting the corresponding second gradual rate of change of static stiffness under load; characterized by , that the first gradual rate of change of the static stiffness of the first driveline mount (22) comprises a first average rate of change of the static stiffness in a first range (Z1) of dynamic loads and a second average rate of change of the static stiffness in a second range (Z2) of dynamic loads; wherein the first average rate of change of the static stiffness is less than the second average rate of change of the stiffness; wherein the loads in the first range (Z1) of dynamic loads are less than the loads in the second range (Z2) of dynamic loads. [2] The method of claim 1, wherein the first gradual rate of change of the static stiffness of the first driveline mount (22) comprises a third average rate of change of the static stiffness in a third range (Z3) of dynamic loads; wherein the third average rate of change of the static stiffness is greater than the second average rate of change of the static stiffness; and wherein the loads in the third range (Z3) of dynamic loads are greater than the loads in the second range (Z2) of dynamic loads. [3] The method of claim 1, wherein the driveline mounting system (18) is a three-point pendulum mounting system; wherein the first driveline mount (22) is a rear torque arm (32); wherein the second driveline mount (24) is a first transmission side mount (24); wherein the driveline mounts further comprise a second engine side mount (26) having a third gradual rate of change in static stiffness under load; and wherein the third gradual rate of change in static stiffness is related to the first gradual rate of change in static stiffness and the second gradual rate of change in static stiffness by the torque balance equation for static equilibrium. [4] The method of claim 1, wherein the driveline attachment system (18) is a four-point driveline attachment system (318) with neutral torque axles; wherein the first driveline attachment is a front torque reaction attachment (323); wherein the second driveline attachment (24) is a rear torque reaction attachment (322); wherein the driveline attachments further comprise a first side transmission attachment (324) and a second side engine attachment (326); wherein the torque balance equation for static equilibrium comprises only the respective torques at the first and second driveline attachments (26); and wherein the torque balance equation for static equilibrium is based on the spatial arrangement of only the first and second driveline attachments (24) relative to the driveline (312). [5] The method of claim 1, wherein the powertrain mounting system (18) is a rear-wheel drive powertrain mounting system (418) having a first lateral engine mount (422), a second lateral engine mount (424), and a rear transmission mount (426); and wherein the first powertrain mount (22) is one of the first lateral mount (422) and the second lateral mount (424), and wherein the second powertrain mount (24) is the other of the first lateral mount (422) and the second lateral mount (424). [6] Driveline mounting system (18) for a vehicle driveline (12), comprising: a first driveline mount (22) exhibiting a first gradual rate of change in static stiffness under load; a second driveline mount (24) exhibiting a second gradual rate of change in static stiffness under load; wherein the first gradual rate of change of static stiffness is related to the second gradual rate of change of static stiffness by a torque balance equation for the static equilibrium of respective torques on the first and second driveline mounts (22, 24); and wherein the torque balance equation for the static equilibrium is based on a spatial arrangement of the first and second driveline mounts (24) relative to the driveline (12); characterized by , that the first gradual rate of change of static stiffness comprises a first average rate of change of static stiffness in a first range (Z1) of dynamic loads and a second average rate of change of static stiffness in a second range (Z2) of dynamic loads; wherein the first average rate of change of static stiffness is less than the second average rate of change of stiffness; and wherein the loads in the first range (Z1) of dynamic loads are less than the loads in the second range (Z2) of dynamic loads. [7] Vehicle (10) comprising: a drive train (12); a powertrain support structure; and a driveline mounting system (18) according to claim 6. [8] A vehicle according to claim 7, wherein the first gradual rate of change of static stiffness comprises a third average rate of change of static stiffness in a third range (Z3) of dynamic loads; wherein the third average rate of change of static stiffness is greater than the second average rate of change of static stiffness; and wherein the loads in the third range (Z3) of dynamic loads are greater than the loads in the second range (Z2) of dynamic loads.

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