As described in Valve Timing and Duration and Valve Design, optimal valve timing and duration changes across the different engine speed and load sites and plays a major role in performance, fuel consumption (BSFC)and emissions. By having the ability to alter the valve timing events these parameters can be optimised and are especially predominant in petrol engines due to the use of a throttle plate to control load, and whilst less prominent in diesel engines offer better control of NOx and particulate emissions. Under part load conditions the valve timing strategy is typically developed to target fuel consumption followed by emissions and heating of the catalytic converter, under full load conditions however the valve timing strategy is typically developed to ensure; target torque and power are achieved, engine protection from knock and preignition and that turbochargers don’t overheat and / or overspeed. Variable valve timing systems can be defined by two categories: Camshaft Phasing – Camshaft phasing systems don’t alter the valve duration or opening and closing speeds but do allow the intake and / or exhaust valve timing to be shifted, either by advancing or retarding the valves opening / closing times. Camshaft phasing systems are generally controlled by engine oil with several examples include Toyota’s VVTi and Ford’s VCT systems. Camshaft Profile Switching – Camshaft switching systems generally offer alternative camshaft profiles, therefore offering alternative opening and closing speeds, durations, opening and closing times and valve lifts. Traditionally however there would be only two differing camshaft profiles available with the engine switching between the two. Camshaft switching systems are generally also oil controlled with examples including Honda’s VTEC and Audi’s AVS systems. Part Load Example with Camshaft Phasing – 2000rpm, 2bar BMEP Following is a BSFC map and analysis of differing valve timing strategies of a spark ignition engine capable of camshaft phasing on the intake and exhaust camshafts running at 2000rpm, 2bar BMEP. Part of the analysis involves plotting cylinder pressure against cylinder volume (P-v Loops) highlighting why certain strategies suffer from higher fuel consumption and reduced combustion stability. Maximum Valve Overlap Under maximum valve overlap conditions the exhaust valve closing is retarded and the intake valve opening is advanced. As a result there was a high level of interaction between the intake and exhaust system, and due to the low load conditions and subsequent low pressure in the intake plenum there was a high backflow of exhaust into the intake system. To counteract the backflow of exhaust products the pressure in the intake plenum had to be increased by increased opening of the throttle, reducing pumping losses across the throttle (min cyl. Pressure = 0.35bar). Due to the lower PMEP, to achieve the target BMEP a lower GMEP was required (and combustion of fuel) resulting in a low BSFC. The issue with maximum valve overlap at 2000rpm, 2bar BMEP was the high levels of exhaust which remained in the cylinder for the next combustion cycle resulting in high combustion instability (measured by the standard deviation of NMEP). Combustion instability was so great to be markedly noticeable in a vehicle and would be an unfeasible valve strategy. Maximum Valve Advance Under maximum valve advance conditions both the exhaust valve closing and the intake valve opening are advanced. As a result the exhaust valve opens under a higher cylinder pressure aiding in the exhausting of the cylinder and a lower level of exhaust remaining in the cylinder. Additionally there is reduced expansion of the combustion gases reducing thermal efficiency which will need to be made up through the additional combustion of fuel to achieve the target load, BMEP. Due to the high level of cylinder exhausting, cylinder pressures are much lower (min cyl. Pressure = 0.28bar) resulting in a lower intake plenum pressure, low throttle opening angle and therefore higher pumping losses which are counteracted by increasing GMEP to achieve the target load. The low level of internal EGR resulted in very stable combustion. Minimum Valve Overlap Under minimum valve overlap conditions the exhaust valve closing is advanced and the intake valve opening is retarded. Due to the early exhaust valve opening the expansion stroke isn’t fully captured although a high level of exhausting will occur due to the high cylinder pressures. Due to the late intake valve opening however as the piston reaches TDC both intake and exhaust valves are closed requiring work by the piston to compress what contents remain in the cylinder. At the time of intake valve opening the cylinder pressure is quite low, however due to the late closing of the intake valve which occurs at approximately halfway through the compression stroke, the intake plenum pressure must counteract the increasing cylinder pressure due to the piston motion. This is achieved by increased opening of the throttle, reducing pumping losses although at the expense of the effective compression ratio. Again due to the low level of interaction between the intake and exhaust systems internal EGR levels are low and therefore combustion was stable. Maximum Valve Retardation Under maximum valve retardation the exhaust valve closing and the intake valve opening are both retarded. Due to the late opening of the exhaust valve the expansion stroke is better exploited and there’s no compression of the cylinder contents due to the subsequently late exhaust valve closing. As the exhaust valve is still open at the start of the intake stroke the piston draws exhaust back into the cylinder, raising the cylinder pressure and therefore the required intake plenum pressure and throttle angle, reducing PMEP. Due to the late intake valve closing the incoming air must act against the rising piston, therefore the intake plenum pressure must increase by increasing the opening angle of the throttle, reducing pumping losses. This type of valve strategy is also a form of the Atkinson Cycle, as the expansion ratio is greater than the compression ratio. Due to the increased levels of internal EGR there was a slight increase in combustion instability although still within limit. Camshaft Phasing at differing Engine Speed and Load Sites Below shows the BSFC maps for several different speed and load sites for a petrol engine capable of camshaft phasing on the intake and exhaust camshafts. These plots shows that: Under low speed and load conditions, a valve strategy that employed high levels of valve overlap or maximum retardation weren’t feasible due combustion instability owing to the increasing percentage of internal EGR. As speed and load increased the valve timing strategy wasn’t limited by combustion stability and could employ the full capabilities of the camshaft phasing. The best combustion stability typically resulted in the worst fuel consumption. Under lower speed and load conditions the best BSFC occurred in the region of maximum retardation and as speed and load increased this moved into the region of maximum valve overlap. At 2000rpm, the best valve strategy for optimal BSFC changed by 30°CA over 0.1 bar BMEP (≈2Nm). When calibrating the camshaft phasing for a vehicle such a change in the camshaft phasing over such a small load step isn’t feasible and therefore when calibrating the camshaft phasing the entire engine map would be a compromise and likely optimised for legislative criteria. Full Load Example with Camshaft Phasing Following is the full load valve timing for a turbocharged, direct injection petrol engine. From 1000 – 2150rpm the engine operated at maximum valve overlap to exploit scavenging of the cylinder. Under scavenging conditions, this engine exploited the boost pressure from the turbochargers (intake plenum pressure was greater than the exhaust system) to push a charge of air through cylinder and straight through to the exhaust, flushing the cylinder of combustants. As a result there was less internal EGR, the engine was less knock limited and a greater mass of air could enter the cylinder for combustion (generating even more boost pressure). Port fuel injection engines can’t exploit this strategy to the extent of direct injection engines due to fuel also bypassing the cylinder increasing fuel consumption and THC emissions. Above 3750rpm the turbocharger was temperature and speed limited due to the increasing exhaust pressure and combustion temperatures. From the previous maximum valve overlap conditions the exhaust valve closing has begun to be advanced whilst the intake valve opening has begun to be retarded. The result is that the exhaust valve is opening earlier during the expansion stroke, sacrificing expansion work however as the cylinder is under a higher pressure the pressure difference between the cylinder and the exhaust system increases the expulsion of combustants from the cylinder. Due to lower internal EGR the engine is less knock limited and will have faster combustion speeds. Due to the increasing engine speed the time available to fill the cylinder is reducing and therefore a later intake valve opening results in higher initial intake flow speeds. Due to boost pressure the intake valve can be closed later due to the difference in the intake plenum pressure and the cylinder pressure allowing more air to be rammed into the cylinder. At 6500rpm, maximum engine speed the exhaust valve closing has been further advanced and the intake valve closing further retarded. Under these conditions the turbocharger was still speed and temperature limited, the time period to intake the fresh charge and exhaust the combustants had further reduced and exhaust pressure was at the maximum recorded levels. Earlier opening of the exhaust valve aided in the exhausting of the cylinder whilst the intake valve was opened later to increase induction speeds and for reduced valve overlap due to the higher exhaust system pressures.