SUPERIOR DOWNSIZING
3.3 Problems and challenges associated with turbocharging the spark-ignition (SI) engine
3.3.4 Knock and abnormal combustion
air path as well as the significant improvement in turbocharger technology.
over the same period of time, naturally aspirated engine performance has been improving due to widespread adoption of camshaft phasing devices.
such devices can also be applied to turbocharged engines and used to shift the operating point in the compressor map and so increase mass flow (with its beneficial effect on turbine work).
A further important operational benefit of using camshaft phasing devices is that the valve overlap can be reduced at part load, thereby minimizing the amount of trapped residual gas in the cylinder. The overlap can be increased as the boost pressure builds up, so that the engine scavenging and turbine mass flow rate benefit from the positive pressure differential across the engine, cooling the combustion chamber surfaces directly. However, with port fuel injection this would result in significantly increased amounts of unburned charge flowing into the exhaust manifold with its consequent combustion and overheating of exhaust system components. Hence camshaft phasing has not been applied as readily to turbocharged engines. With the advent of direct injection systems, however, there is now a means to prevent the introduction of fuel into the air until after the exhaust valve has closed and hence to provide a means to influence engine operation by compressor map shifting and to improve driveability further. The benefits of this will be returned to in section 3.4 below.
top ring land [27, 28]. If left unchecked it can destroy an engine in a matter of a few cycles.
Consequently, knock has been acknowledged to be a significant limit to engine performance since the dawn of the internal combustion engine [29].
Ricardo was the first to postulate the process by which knock occurs [30], providing a description of the process which still holds today, and which, with the benefit of many years’ subsequent investigation, can be described as follows. Following ignition by the passage of the spark across the spark plug electrodes, there is a brief delay during which a flame kernel forms.
Providing fluid motion does not extinguish the flame, once the temperature of the flame kernel reaches a critical point, rapid inflammation of the rest of the mixture begins and proceeds by the spreading of a deflagrating flame away from the spark plug. The time taken for the flame to break out of the kernel and into the bulk of the mixture is subject to cyclic irregularity and is one of the reasons why the first 5% of the mass fraction burned (MFB) is often disregarded in analysis of combustion statistics – indeed, 10–90%
MFB is usually taken as an indicator of combustion duration (with 90% as the upper limit due to the related difficulty of establishing when all of the charge has been consumed).
Combustion of the charge within the flame front causes the gases behind it to heat up and consequently expand, both accelerating the flame front’s passage in spatial terms and compressing the unburned gas ahead of the flame. This ‘end gas’ is also heated by radiation and so, due to both processes, increases in temperature and pressure before the flame reaches it. Whilst these processes are occurring, the end gas rejects some of its heat energy to the engine structure [31], which also influences the temperature and pressure histories that it undergoes.
The end gas temperature and pressure histories (which are also modified by piston motion and surface temperature effects) cause the occurrence of low-temperature chemistry in the fuel as a form of pre-flame reaction.
This can assist in facilitating the passage of the flame but equally, as it is being heated by all of the events occurring in the chamber, the end gas will autoignite if and when the Livengood–Wu integral equation
Ú
t1 d = 1t 3.1is satisfied [32]. The parameter t in this case is the Livengood–Wu induction time representing the autoignition delay time at the instantaneous temperature and pressure of the mixture, and is defined by an Arrhenius rate equation of the form
t = Ap– exp B
n ÈT
ÎÍ ˘
˚˙ 3.2
where A, n and B are constants which vary for different fuels at different conditions in different engines. If the combustion is completed before equation 3.1 is satisfied then autoignition is avoided. If autoignition occurs and if in- cylinder conditions are correct, a flame front can then traverse the remaining end gas at 10–20 times the normal flame speed, in turn causing molecular vibration and bulk excitation of the engine structure [27, 33, 34].
The local temperature gradient which can be promoted as a result of exothermic centres (arising either from the surface or from within the charge itself due to pockets of hot retained burned gas) is important with respect to the severity of the autoignition event if it occurs [35–37]. Up to five different types of flame propagation have been identified [37], but in general if the thermal gradient in the end gas is high (>100 K/mm) a deflagrating flame ensues, if it is low (<1.25 K/mm) a thermal explosion occurs, and if it is of a medium value (~12.5 K/mm) then there is a developing detonation. The thermal explosion case is that which is actually aimed for in homogeneous charge compression ignition (HCCI) engines, since it implies a softer autoignition than a knock event; however the processes involved in knock and HCCI, while related, are different. In sI combustion, the developing detonation case is that which leads to destructive knock [38], which will occur only if sufficient end gas has not been consumed by the flame front.
Given that the flame front cannot consume the exothermic centres before they reach autoignition, the higher the proportion of charge in the end gas involved in the autoignition event, the more severe the knock amplitude and the more likely the danger of structural damage [38]. The margin for control also reduces as the amount of charge involved in the knock event increases [38].
From the foregoing it is apparent that, unless the flame front can consume any exothermic centres, thermal energy input to the charge is extremely important with regard to the likelihood and severity of any knock event as it controls the evolution of the autoignition integral. When operating on any given fuel, the turbocharged sI engine is at an immediate disadvantage in comparison to its naturally aspirated counterpart for several reasons: all of the thermal issues discussed in section 3.3.2 affect end gas temperature, and the increase in pressure due to the boost applied affects the temperature and pressure histories too. It must also be remembered in the context of knock in general that the high trapped mass (and therefore heat released per cycle) in a turbocharged engine at full load increases the severity of any knock event significantly over that which would occur in a naturally aspirated engine of the same swept volume.
as a consequence of the above, the knock limit is lower in pressure- charged engines than in naturally aspirated ones, so that often the compression ratio of the engine needs to be reduced and the engine needs to be operated with retarded ignition timing or under ‘over-fuelled’ conditions (to make
combustion less efficient and hence to cause less end gas heating to occur).6 under conditions of adverse pressure gradient across the valves, when the pressure at the exhaust valves during the valve overlap period exceeds that at the intake valves, a reduction in compression ratio necessarily increases the mass of residual gas in the chamber. This then causes the knock limit to contract further, both because of any active species present and because of the thermal effect (both global and local if the residual gases are not well mixed, causing the hot spot effects discussed above). Various technologies have been investigated to mitigate this effect in PFI turbocharged engines, including turboexpansion [39–42] and the use of a divided exhaust period [43].
Turboexpansion seeks to use over-compression of the charge air to drive an intake air expander which further removes heat energy from the charge, and so to provide a cooler charge to the engine. Between the compressor and expander stages charge-air cooling occurs at the elevated pressure condition.
The combined process has the potential to provide a given charge-air density at a lower temperature than that of a conventional intercooling process alone.
a schematic of the charge-air cycle is shown in Figs 3.6 and 3.7; this cycle is a form of air refrigeration and has also been used for aircraft air-conditioning systems. Turner et al. [40] found that, in a newly specified and designed test engine [44], the combustion benefits of turboexpansion were difficult
6 While over-fuelling is also sometimes used to reduce gas temperatures to control component temperatures, retarding the ignition works counter to this aim with regard to components in the exhaust gas path.
Compressor
Turbocharger
Air filter
Turbine
X
Y
Z I/C
EBP
Plenum
Engine Turboexpander
Energy recovery (electrical, mechanical or hydraulic, etc.)
3.6 Schematic of one possible turboexpansion system for a turbocharged engine. I/C = charge-air intercooler; EBP = exhaust back pressure; X, Y and Z refer to states in the temperature–entropy diagram shown in Fig. 3.7.
to realize because of the interactions between the different components in the charging system (further investigated by Taitt et al. [41]), though more fundamental combustion research is ongoing to assess its attractiveness.
The concept of ‘divided exhaust period’ (see the schematic shown in Fig.
3.8) divides the flow from the two exhaust valves in an engine and sends the first ‘blowdown’ pulse from the cylinder directly to the turbocharger turbine through a dedicated exhaust valve (‘a’ in Fig. 3.8) before closing that valve early and opening the second exhaust valve (‘b’) to perform final scavenging at an elevated pressure differential from intake to exhaust during the overlap period [43]. as a consequence of this, both of the staggered exhaust valve
Temperature, T
Entropy, S
Atmospheric conditions Y
Z
X Pupper
Pplenum
Patm
3.7 Temperature–entropy diagram for the turboexpansion charging system shown schematically in Fig. 3.6.
3.8 Schematic of divided exhaust period system for turbocharged engines. I/C = charge-air intercooler. Adapted from Moller et al. [43].
Exhaust valves ‘a’ and ‘b’
with separate exhaust ducts I/C
a b a b a b a b
Starter catalyst
Main catalyst
Air filter Valves ‘a’ –exhaust blowdown to turbine Valves ‘b’ – scavenging valves
events are less than 180∞ long, and there is no pulse interaction between the cylinders, though the amount of valve time–area available to flow the burnt gas from the cylinder is necessarily reduced. An additional benefit is that the high pre-turbine pressure caused by using a small turbine would not have a detrimental effect on the scavenging efficiency of the engine, because the cylinder is not exposed to this pressure in the valve overlap period. Consequently it is possible to achieve a positive pressure gradient across the cylinder during the valve overlap period higher up in the engine speed range. By fitting a valve deactivation system to the blowdown valves, the turbine can be bypassed at startup and a starter catalyst used to improve emissions performance (see next section). an illustration of the valve events used in this concept is shown in Fig. 3.9.
a more pragmatic solution to exhaust pulse interaction in many pressure- charged engines is to divide the exhaust system itself and to use two separate entries to the turbocharger turbine. For example, in a 4-cylinder engine with a firing order of 1-3-4-2, if cylinders 1 and 4 are paired together separately to 2 and 3, the blowdown pulses are separated from the next cylinder in the firing order. An example of a pulse-divided turbocharger is shown in Fig. 3.10 and the general concept is shown in Fig. 3.11. The concept was discussed by Groff et al. [45], although andriesse et al. [46] presented a slightly contradictory view, stating that with a 4-into-1 manifold connected to a single turbine housing, oxidation of carbon monoxide in the exhaust system before the turbine adds to the enthalpy available to drive it. This subject will be returned to in section 3.4.3, and the general subject of architecture and turbocharged engines in section 3.6.
In conventional turbocharged engines employing no technology specifically to mitigate the thermal and chemical issues of residuals trapped in the gas
3.9 Representation of valve events used in the divided exhaust period system for turbocharged engines. Adapted from Moller et al.
[43].
Exhaust blowdown valve
Exhaust
scavenging valve Intake valves
Valve lift
90 180 270 360 450 540 630
Crank angle (degrees at TDC firing)
exchange process, retarding the ignition is often employed and is a very powerful control on knock. This approach results in a later-phased combustion event which changes the phase of combustion heat release versus piston position and reduces work output. Importantly it also results in an increase in gas temperature at the exhaust valve opening point and thus increases the thermal input to the engine structure. overfuelling is also sometimes applied to counter knock and to control component temperatures, both in the combustion chamber and of components in the exhaust bath. reducing compression ratio is another palliative which works by decreasing the
3.10 Pulse-divided turbocharger: note dividing wall cast into housing to maintain separation of exhaust gases through the volute and up to turbine wheel entry (courtesy General Motors).
3.11 Concept of pulse division in turbocharged engines (even-firing 4-cylinder configuration).
Cylinder 1
Cylinder 4 Cylinder 3
No overlap or pulse interaction Cylinder 2
Cylinder 1 Firing order 1-3-4-2
720 crank angle Scroll for
cylinders 1 and 4 Scroll for cylinders 2 and 3
combustion efficiency to reduce the heat release rate. However, when these approaches are considered together, a destructive spiral ensues, which is at the heart of why PFI turbocharged engines historically have worse fuel consumption than naturally aspirated ones.
To illustrate one way out of this spiral, consider a turbocharged PFI racing engine versus its road-going cousin. In the racing engine, assume that the exhaust back-pressure can be removed significantly because the presence of the turbine provides nearly all of the silencing required for racing, and that a racing catalyst (with low cell density and hence pressure loss) is used. In this case the product of turbine expansion ratio and overall exhaust system back-pressure is significantly lower than in the road-going unit. as a consequence, for a given limiting temperature, either the boost or the compression ratio could be increased, and the ignition timing perhaps advanced, lowering the temperature at exhaust valve opening, and in turn permitting less fuel enrichment to be used to the benefit of fuel consumption and power output. This example is laid out in order to show how a significant improvement in one area of the engine–turbocharger system leads to a significant rebalancing of the whole, and the concept will be returned to in the discussion of the benefits of direct injection in SI turbocharged engines in section 3.4 below.