Patent Application
20070204818
The present invention relates to variable valve activation (VVA) and variable valve timing (VVT) mechanism for internal combustion engines; more particularly, to VVA/VVT mechanism for modulating the timing of compression ignited (CI) engines such as diesel engines; and most particularly, to a valve lash adjuster having electro-hydraulic lost-motion capability for varying the lift and/or timing of individual valves in a valve train of a multi-cylinder engine.
Means for varying the timing of valve actuation of internal combustion engines are well known. Such means typically take the form of a camshaft phaser or an element of a valve train, such as rocker arms, roller finger followers, hydraulic valve lifters, or hydraulic lash adjusters, provided with a mechanism for switching between a valve activating mode and a valve deactivating mode. VVA/VVT is especially well known in spark-ignited (SI) engines, in which it is an essential element of various schemes for improving fuel economy. However, camshaft phasers, while readily applied to SI engines, are not as suitable for CI engines since cam phasing introduces the risk of catastrophic valve-to-piston collisions due to the close proximity of the piston crown to the cylinder head at top dead center (TDC) at which point the valves are obliged to be closed. For this reason, an alternative technology, known in the art as “lost motion”, has found increasing favor for VVT/VVA in CI engines, since much of the functionality of existing SI VVT/VVA systems is available without danger of piston/valve collisions.
The potential importance of VVT/VVA to CI engine performance is coming to be realized within the engine industry. The firing of an SI engine is readily and accurately controlled by simply controlling the timing of the ignition spark. The firing of a CI engine, and more particularly so the firing of a controlled auto-ignition (CAI) or homogeneous charge compression ignition (HCCI) engine on the other hand, is governed by a plurality of independent or loosely-dependent variables which conspire to cause the fuel/air charge to explode at some resultant combination of temperature, pressure, and mixture.
These variables, which include, but may not be limited to, cylinder temperature, cylinder pressure, valve train timing and wear, fuel injection timing and accuracy, homogeneity of the fuel/air charge, and thermal load of the fuel/air charge, can vary from cylinder to cylinder in an individual engine and furthermore can vary for any given cylinder from one firing cycle to the next. Thus, in the prior art the exact point in the compression stroke at which the compressed charge in a cylinder will ignite cannot easily be predicted or controlled to a very high degree of certainty, and in practice the cylinders of a multiple-cylinder HCCI engine may not fire with a degree of uniformity required to meet future performance standards.
In a CI engine, the trapped air mass is the “charge volume” in the cylinder upon which compression work is done. Because adiabatic compression of the charge volume is the mechanism by which CI ignition is induced, an important ignition factor is the “Effective Compression Ratio” (ECR) within the cylinder. Thus, direct control of ECR can provide improved control of firing timing both in individual cylinders and among the cylinders in a CI engine. Controlling ECR by increasing the compression ratio can improve cold start characteristics, and by decreasing the compression ratio can improve engine performance. Other engine control strategies that can be attained by strategically controlling the opening, closing and lift of the gas valves in a CI engine, as more fully described in co-pending U.S. patent application Ser. No. 11/027,109, include in-cylinder swirl of intake gases to provide effective mixing of injected fuel and air, and controlled Exhaust Gas Recirculation (EGR) to control combustion initiation and burn rates, while lowering flame temperatures for reduced NOx emissions.
In a Type 2 engine valve train, a roller finger follower (RFF) typically is interposed between an inwardly-opening poppet valve stem tip at one end and a hydraulic lash adjuster (HLA) at the distal end, with a cam lobe providing motivation to the RFF at an intermediate point. For reasons of good dynamic performance at high speed, low friction, and convenient packaging, this mechanism is rapidly becoming the valve train of choice for many new light-duty engines today, both SI and CI.
In a Type 3, 4, or 5 valve train, a rocker arm pivots on a rocker shaft, with one end of the rocker arm being motivated by the camshaft either directly or through the medium of a follower and/or pushrod, and the other end actuating the engine valve. For reasons of valve train cost, packaging convenience, or tradition, these systems are frequently used for medium- to heavy-duty engines and may or may not use an HLA. (For simplicity of presentation hereinbelow, Type 3 should be understood to mean all central-pivot rocker arm engines, including Types 4 and 5.)
In another version of the Type 3 valve train, the rocker arm pivots on an inverted HLA instead of a rocker shaft. Since the HLA is stationary, this type of valve train offers reduced dynamic mass advantages over other Type 3 valve train.
Lost motion means in a valve train element switches the linear motion imparted to the valve train by a rotating cam between either of a valve stem/lifter/pushrod or rocker arm and a lost motion spring/piston/accumulator. In the valve activating mode, the switchable element is mechanically and hydraulically competent to transfer the motion instructions of the cam to the valve; but in the valve deactivating mode, the switchable element collapses by a controlled amount and at the appropriate time in some fashion to “lose” the motion of the cam and belay those instructions to the valve. See, for example, U.S. Pat. No. 6,883,492.
Serious drawbacks of such known VVA/VVT systems are that they employ engine lubricating oil as the hydraulic medium, which tends to be dirty, carbon-laden and relatively high viscosity, requiring relatively large passageways to prevent flow failure; they employ a relatively bulky, powerful solenoid control valve which because of its size has a relatively slow speed of response; and they introduce significant additional complexity to the cylinder head that, in so doing, creates problematic packaging and manufacturing issues.
It is highly desirable that any apparatus and control system for improved control of ECR be applicable to existing arrangements of Type 2 and Type 3 engine valve trains with a minimum of engine redesign.
What is needed in the art is an improved means for controlling engine strategies such as, for example, ECR, EGR and in-cylinder swirl in a CI engine.
What is further needed is that such improved means be applicable to, and controllable for, individual cylinders in a multiple cylinder engine.
It is a principal object of the present invention to improve control of various engine control strategies in a CI engine.
It is a further object of the invention to provide such improved control with minimum redesign requirements for Type 2 and Type 3 engines.
It is a still further object of the invention to provide such improved control through novel adaptation of existing fuel injection equipment (FIE) technologies which have been demonstrated to have the speed of response, precision, and durability required for an ECR control system.
Briefly described, in a Type 2 engine, a valve deactivation hydraulic lash adjuster (DHLA) in accordance with the invention replaces a conventional hydraulic lash adjuster in the train of a gas-exchange valve in an internal combustion engine, and especially a compression-ignited engine. In a Type 3 engine, a DHLA is disposed within an articulated rocker arm which is made selectively competent (valve activating) or incompetent (valve deactivating) thereby. In the version of the Type 3 engine described above, a DHLA is disposed above the rocker arm at a center pivot point and replaces a conventional hydraulic lash adjuster.
In each approach, an electrically controlled solenoid valve within the DHLA assembly diverts hydraulic fluid between support and non-support of a piston slidably disposed in an elongated housing and terminating in a ball head. The geometry for the valve is such that the valve pin is force-balanced, thus reducing significantly the required strength and size of solenoid relative to the prior art solenoid and valving.
The preferred hydraulic fluid is diesel fuel which is readily available in a diesel engine from the low-pressure fuel supply pump for the engine fuel rail, which fuel preferably is used to lubricate the entire “top end” of the engine. Relatively dirty, carbon-contaminated engine lubricating oil is relegated to lubricating only the engine crankshaft and cylinder bores. The continuously refreshed and very much cleaner diesel fuel is used to lubricate the camshaft bearings and valve train, with conventional seals providing segregation of the two fluids. It is understood that the lubricity of diesel fuel is less than that of lubricating oil, however there is a clearly defined minimum standard for lubricity against which the bearing surfaces may be designed. Use of diesel fuel rather than engine lubricating oil allows for substantially smaller diameter passages and cleaner operation than in prior art systems. The use of smaller passages reduces very significantly the amount of fluid to be diverted between operating modes, eliminating the need for an accumulator chamber and accumulator piston as in the prior art.
The present invention will now be described, by way of example, with reference to the accompanying drawings, in which:
a is an elevational cross-sectional view of a prior art Type 2 valve train employing the hydraulic lash adjuster shown in
a is an elevational cross-sectional view of a Type 3 valve train employing a first embodiment of a Type 3 valve deactivation rocker arm (VDRA) in accordance with the invention, shown in valve-activation mode;
b is a cross-sectional view of an embodiment showing a pushrod between the DHLA and buttress of
The benefits and advantages of a valve deactivating electro-hydraulic lash adjuster in accordance with the invention may be better appreciated by first considering a prior art variable valve deactivation system.
Referring to
In operation, rotation of camshaft 12 causes oscillation of rocker arm 14 which drives master piston 26 within main bore 28. If trigger valve 40 is closed (valve-enabling mode), slave piston 30 is displaced by the motion of master piston 26, displacing bridge 24 and actuating valves 22. If trigger valve 40 is open (valve-disabling mode), the valve springs keep valves 22 closed, and accumulator piston 46 is displaced in lost motion within accumulator chamber 44 by the motion of master piston 26.
As noted above, known VVA/VVT systems such as prior art system 10 have several serious drawbacks, which the present invention overcomes. First, they employ engine lubricating oil as the hydraulic medium, which tends to be dirty and relatively high viscosity, requiring relatively large passageways to prevent flow failure. Furthermore, the dirty fluid encourages the use of relatively low hydraulic pressures to minimize abrasive wear, which in turn requires large diameter pistons. To minimize parasitic loss of hydraulic fluid, an accumulator piston device is necessary to store and recuperate the displaced fluid. Second, they employ a relatively bulky, powerful solenoid and non-force-balanced control valve which because of their size have a relatively slow speed of response; further, they occupy an undesirably large region within and adjacent to the cylinder head. Third, they introduce significant additional complexity to the cylinder head that requires extensive and expensive redesign and modification of engine manufacturing lines.
Referring to
The objective of an HLA such as HLA 50 is to eliminate the lash that is an essential feature of traditional mechanical valve train linkages. This lash, or clearance, changes due to engine temperature, initial setting errors, and valve train wear. The lash adjuster compensates for these effects through a limited range of automatic adjustment, typically up to 5.0 mm maximum, so that intended engine operation is continuously maintained.
Referring to
Selective disabling of the motion of valve stem 94, to selectively deactivate the associated engine valve in accordance with the invention, may be achieved by providing an improved hydraulic lash adjuster means that is selectively compressible in lost motion in lieu of motion of valve stem 94. Provision of such lost-motion to the valve train requires a larger range of controlled adjustment than is required for HLA 50, and this must be equal to or greater than that of either the cam lift or the full valve lift, depending upon the system geometry. Thus a lost-motion valve train mechanism must incorporate this extended range of collapsible travel to negate the cam lift; must provide additionally the adjustment range of a lash adjuster; and also must provide a control mechanism having a speed of response high enough to provide resolution of valve event control to within ±5 crank degrees of a target value. Moreover, this functionality must be packaged in a manner that makes it compatible with the constraints of modern engines, which implies that it should be dimensioned substantially similar to existing valve train components. Thus replacing a conventional HLA for selective valve deactivation is an excellent and novel strategy that can meet these criteria.
Referring to
A steel body 108 close-fitting into well 180 includes a snout 110 having a bore 112 at its outer end into which a close-fitting piston 114 is slidably disposed. A requirement for minimal fluid leakage between piston and bore suggests that these items are preferably match-ground during manufacture as is common practice in manufacture of fuel injectors. A sleeve 116 having a spherical end 118 and a well 120 is slidably disposed on snout 110. Spherical end 118 engages with hemispherical seat 88, and piston 114 has an exposed length sufficient to accommodate both valve lost-motion travel and lash adjustment travel. Piston 114 is preferably not attached to sleeve 116 but rather engages sleeve 116 via a pressure plate 122 defined by the bottom of well 120 over which piston 114 is free to wander. Thus, side loads applied to DHLA 100 from RFF 84 are carried by sleeve 116 and snout 110, leaving piston 114 free to reciprocate within bore 112 without bending loads being imposed upon it. An external bias spring 123 is disposed around sleeve 116 between body 108 and head 118 to urge DHLA 100 into continuous contact with RFF 84.
Body 108 is preferably sealed into well 180 by first and second O-rings 124 disposed on either side of a hydraulic fluid gallery 126 formed in head 29 for supplying hydraulic fluid to DHLA 100. As discussed hereinabove, a preferred hydraulic fluid in accordance with the invention is diesel fuel which may be conveniently supplied to gallery 126 parasitically from a low-pressure pump (not shown) in an existing engine system for supplying diesel fuel to a fuel injector rail (not shown). The low-pressure stage in a typical diesel engine fuel supply system operates at about 6 bar, which pressure is quite suitable for operation of DHLA 100.
Within body 108 and communicating with bore 112 via first and second drillings 128, 130 is a spring-loaded inlet check or non-return valve 132 which may be of the ball or flute-guided conical seat type. Preferably, valve seat 134 is pressed into place to retain the valve components, although alternative constructions and orientations are fully contemplated by the invention, including but not limited to a reversed valve wherein the seat is machined directly into the body and a blanking plug is used to retain the spring; and a valve bore drilled at an angle which may allow packaging in a smaller diameter body. A shallow annular groove 136 permits access of hydraulic fluid from gallery 126 to check valve 132 at any installed rotational orientation of DHLA 100.
Further within body 108 is a lost-motion control valve 138 disposed in an axial bore 140 having a valve seat 142. Valve 138 which is spring-biased closed, is of a two-way two-position “balanced” spool design wherein opening and closing hydraulic forces are substantially equal at all times. A reduced-diameter central portion 144 of valve pin 139 is in communication with check valve 132 via third and fourth drillings 146, 148, drilling 146 being blanked by a plug 150. Hydraulic pressure created within body 108 acts upon control valve 138, but since the upper piston portion of the valve is substantially the same diameter as the diameter of seat 142, there is no resultant axial force due to pressure on the control valve; thus a low-force actuator is adequate to provide rapid valve motion.
An armature 152 disposed in an armature cavity 154 formed in body 108 is connected to valve pin 138. In opening, valve 138 is withdrawn off of seat 142 into cavity 154. Hydraulic fluid flows across seat 142 into armature cavity 154 and into a drilling 156 that breaks out into the under valve cover area 158 of the engine. Preferably, a cross drilling 160 breaks into bore 140 above valve 138 so that both ends of the valve are pressure balanced and in communication with area 158 pressure.
Within well 180 in head 29, a solenoid stator assembly 162 is disposed adjacent body 108. Assembly 162 includes a steel shell 164 providing a load path for the valve train forces into cylinder head 29. Shell 164 receives a molded-in stator 166, solenoid windings 168, and a valve return spring 170 surrounding armature 152. Electrical leads 172 extend to spade lug 106. Preferably, stator assembly 162 is assembled and tested prior to attachment to adjuster body 108 as by electron-beam or laser welding.
In normal operation, diesel fuel oil is supplied to the gallery 126 in the cylinder head 29 that feeds the deactivation hydraulic lash adjuster 100 (and all other such lash adjusters in a bank of engine cylinders). At a nominal feed pressure of 6 bar, this fuel is readily able to overcome the light spring load of check valve 132 and therefore to fill the interior drillings 128, 130, 146, 148 and thereby assist bias spring 123 in loading RFF 84 against camshaft 12. With lost-motion piston 114 extended so that the RFF is in contact with the cam on its base circle portion, all mechanical lash in the valve train is taken up, and check valve 132 then seats under coercion of its own spring. As camshaft 12 rotates in operation of the engine, an axial compressive load is applied to piston 114 that is in proportion to the dynamic load from the engine valve and spring. This load is reacted by hydraulic fluid captive within DHLA 100 and therefore, except for any undesired leakage past piston 114, check valve 132, and control valve 138, the piston and therefore the fulcrum for RFF 84 remains stable. By keeping internal fluid volume to a minimum, piston depression due to fluid compressibility is very low. In this situation, no activation of control valve 138 occurs and engine valve motion is determined by the cam profile.
If an early valve closing event is desired during a normal open cycle of an associated engine valve as just described, control valve 138 is energized via controller 139, causing release of hydraulic fluid from bore 112 which results in controlled collapse of piston 114 as the engine valve rapidly returns to its seat. Prior to the engine valve reaching its seat, the control valve may be re-energized, thus stopping further fluid release and arresting valve train motion. Engine valve seating can be closely controlled by this means in regions where the cam velocity is at or below an acceptable threshold value, typically about 0.2 m/sec at low engine idle speed. If lost-motion functionality is required over a wider range of timing authority including areas of high cam velocity, a valve seating snubber is required, as described further hereinbelow. Energizing of the control valve is managed by a computerized Engine Control Module (ECM) (not shown) in known fashion, with input from a crankshaft encoder and potentially with feedback from an accelerometer on the engine to give an indication of valve seating impact.
If total valve deactivation is desired, such as is the case if cylinder deactivation is desired, control valve 138 is energized for the whole cam lift event so that all motion of the cam and RFF is absorbed by lost-motion displacement of piston 114 within bore 112.
If it is desired to delay opening of an engine valve, control valve 138 may be energized while the RFF is on the base circle portion of the cam. Initial cam motion of the RFF on the cam eccentric portion serves to depress piston 114 in lost motion until such time as valve lift is desired, at which point control valve 138 is de-energized allowing it to close, and the resulting internal hydraulic lock provides a stable fulcrum in a new and lower position for the RFF, resulting in a later opening and lowered, centered lift event.
If, during normal operation there is a tendency for DHLA 100 to “pump up” as is known to happen occasionally for prior art HLAs due to dynamic fluctuations in the valve train, control valve 138 can be energized momentarily to normalize the internal pressure prior to a scheduled valve event, or prior to TDC if there is concern for a possible piston/valve collision.
Referring now to
Referring now to
In the prior art, a comparable Type 3 rocker arm is an inflexible unit wherein rotary motion of the cam is translated faithfully into reciprocal motion of the valve stem. Prior art lash adjustment typically is provided by either a screw head (not shown) on valve stem 94 or a hydraulic valve lifter (HVL) assembly (not shown) disposed between arm 320 and valve stem 94. In HLA embodiment 300, rocker arm assembly 314 is provided as first and second arms 320, 322 independently and rotatably mounted on shaft 316. First arm 320 is adapted to engage valve stem 94.
A well 380 is provided in second arm 322. HLA 300 is disposed in well 380 and is provided with hydraulic fluid, for example, engine oil, via an axial gallery 324 in shaft 316 and a radial passage 325 extending to a connector drilling 326 in arm 322 that communicates with HLA 300. As shown in the exemplary assembly in
First arm 320 is provided with a buttress 328 having a wear surface 330 for receiving the spherical head 318 of HLA 300. It will be seen that as HLA 300 expands in accordance with the prior art arm 322 is urged away from buttress 328, causing arm 322 and/or arm 320 to counter-rotate on shaft 316, thus changing the angular relationship between the two elements until all mechanical lash in valve train 378 between camshaft 12 and valve stem 94 is eliminated.
Referring now to
In DHLA embodiment 400, rocker arm assembly 414 is provided as first and second arms 420, 422 independently and rotatably mounted on shaft 416. First arm 420 is adapted to engage valve stem 94 as described further below, and second arm 422 is supportive of cam follower roller 417. A well 480 in second arm 422 is analogous to well 380 in
First arm 420 is provided with a buttress 428 having a wear surface 430 for receiving the spherical head 418 of DHLA 400. In an alternative embodiment shown in
Electrical connection of a pigtail to DHLA 400 may be provided via a tail bore 402 (analogous to bore 102 in
In operation in valve-activating mode, as DHLA 400 expands like a conventional HLA, arm 422 is urged away from buttress 428, causing arm 422 and/or arm 420 to counter-rotate on shaft 416, thus changing the angular relationship between the two elements until all mechanical lash in valve train 478 between camshaft 12 and valve stem 94 is eliminated. Because rocker arm assembly 478 is thus hydromechanically rigid, rotary action of camshaft 12 is faithfully translated into reciprocal action of valve stem 94.
If in normal operation there is a tendency for DHLA 400 to “pump up” as is known to occur occasionally in prior art HLAs due to dynamic fluctuations in valve train 478, control valve 138 may be energized momentarily to normalize internal pressure prior to a scheduled valve lift event or prior to engine piston TDC if there is concern for a potential piston/valve collision.
Referring now to
In some applications, it may be desirable to allow the engine valve to open fully in valve-activation mode but to close the valve early by changing assembly 478 to valve-deactivation mode on the closing slope of the valve lift cycle. Energizing of solenoid assembly 162 at this point causes immediate collapse of DHLA 400, resulting in the full closing force of the compressed engine valve spring being brought to bear on the engine valve. Such abrupt closing can cause objectionable valve clatter as well as excessive valve wear and premature failure. Accordingly, a snubber assembly 500 is preferably included in valve train 478 to arrest the motion of valve stem 94 before the valve fully closes and to provide a graduated valve closing.
A problem with typical prior art hydraulic valve seating snubbers is that they have a fixed snubbing characteristic irrespective of operating conditions other than fluid viscosity. In the present case, snubbing is desired over only the final half millimeter or so of engine valve head travel before engagement with the valve seat, and this can be very difficult to arrange for all valves In an engine under all conditions of valve seat wear, recession, and assembly tolerances.
To cater to these variables and to ensure that seating velocity is controlled over only the final portion of travel, a currently preferred embodiment of variably controllable snubber assembly 500 incorporates an eccentric snail cam adjuster 502 at the rocker tip 504 whereby the base circle portion 505 of the snail cam contacts the tip 506 of valve stem 94. The snail cam adjuster is rotated during engine assembly to engage the snubber piston 508 by a desired amount with the engine valve seated. This sets the range of action of the snubber in absorbing travel of the valve stem during the act of valve closing. Snail cam adjuster 502 is formed of a wear-resistant material such as hardened steel and is retained in a slot in rocker tip 504 by a through bolt 510.
Snubber 500 comprises a body 512 having a stepped bore therethrough comprising a first diameter region 516 and a narrower diameter region 518. A stepped snubber piston comprises a large-diameter portion 508a and a small diameter portion 508b and is slidably disposed in bores 516/518. Bore 516 is closed as by a screw plug 520, forming a closed chamber 522 above piston 508a. Bore region 516 is hydraulically pressurized dynamically to a predetermined pressure, preferably by connection to the same source of hydraulic fluid supplying fluid to gallery 424 for DHLA 400. The pressure within the snubber assembly may be controllably varied in known fashion as a function of engine speed or other desired parameter to controllably modulate the valve seating velocity.
Referring now to
DHLA 600 is disposed in well 680 formed in head 629. Well 680 is open ended at its distal end to accommodate an electrical connector (not shown) for spade lug 606.
Body 608 of DHLA 600 is close-fitting into well 680 and includes a snout 610 having a bore 612 at its outer end into which a close-fitting piston 616 is slidably disposed. Piston 616 includes spherical end 618 which engages valley 605 as described above. Piston 616 has a length sufficient to accommodate both valve lost-motion travel and lash adjustment travel.
Body 608 is preferably sealed into well 680 by first and second O-rings 624 disposed on either side of a hydraulic fluid gallery 626 formed in head 629 for supplying hydraulic fluid, preferably diesel fuel, to DHLA 600. Within body 608 and communicating with bore 612 via first and second drillings 628, 630 is a spring-loaded inlet check valve 632 which may be of the ball or flute-guided conical seat type. Preferably, valve seat 634 is pressed into place to retain the valve components. A shallow annular groove 636 permits access of hydraulic fluid from gallery 626 to check valve 632 at any installed rotational orientation of DHLA 600.
The details of lost-motion control valve 638, armature 652 and solenoid stator assembly 662 are identical to valve, armature and stator assembly 138, 152 and 162, as described in the first embodiment, and need not be described again.
In operation, in valve-activating mode, as DHLA 600 expands like a conventional HLA to provide a pivot point on which rocker arm assembly 614 pivots to cause the engine valve to open, all mechanical lash in valve train 678 between camshaft pushrod 603 and valve stem 94 is eliminated. In operation in valve-deactivating lost-motion mode of DHLA 600, control valve 638 is opened by energizing of solenoid assembly 662, allowing collapse of piston 616 into body 608 in lost motion in response to raising of pushrod 604 by the camshaft. Valve stem 94 is not actuated and the associated engine valve is not opened.
Snubber assembly 500 may be similarly provided, as described above with respect to the third embodiment, to arrest the motion of valve stem 94 before the valve fully closes and to provide a graduated valve closing.
While the invention has been described by reference to various specific embodiments, it should be understood that numerous changes may be made within the spirit and scope of the inventive concepts described. Accordingly, it is intended that the invention not be limited to the described embodiments, but will have full scope defined by the language of the following claims.
