High Pressure Turbocharging On Gas Engines

Transcription

1 High Pressure Turbocharging On Gas Engines E. Codan S. Vögelih, C. Mathey ABB Turbo Systems Ltd Bruggerstrasse 71a, CH-5401 Baden, Schweiz Abstract High pressure turbocharging opens new development potential for diesel and gas engines. This paper describes improvements for gas engine performance and efficiency while at the same time keeping current low emission values as a priority. The specific problem of controlling gas engine power via mixture mass and equivalence ratio is discussed in detail, taking into account the increased complexity of 2-stage turbocharging. In order to achieve the demonstrated engine performance and efficiency potential, suitable turbocharging concepts are a prerequisite. However, it is of utmost importance that the partners involved (i.e. the engine builders and the turbocharger manufacturers) maintain close cooperation in order to realise overall system optimisation. Key Words: Miller Cycle, Gas Engines, 2-stage Turbocharging, Engine Efficiency

2 1 Introduction The market for gas engines is undergoing expansion at a previously unknown rate, especially in developed countries. For this there are various political and economic reasons. The availability of the fuel, involving both the material itself and the supply infrastructure, is a basic precondition. The pricing policies of the individual countries based, in part, on considerations regarding storage, availability and dependence on producers, play an equally important role. There are, however, also technical reasons for the success of the gas engine. Technological progress in the combustion as well as in the process of gas engines has already brought them to a point where power density and efficiency can be compared with the values of a diesel engine. As a third factor emissions behaviour plays an important role. By their nature gaseous fuels allow combustion at low noxious emissions values. In particular, particulate emissions are at a very low level. Good engine efficiency and the favourable hydrogen to carbon ratio of the fuel guarantee an advantage regarding CO 2 emissions. By means of lean burn technology, NO x emissions can be held at an extremely low level without appreciable penalties in terms of engine efficiency. In this way, the opportunity arises for gas engines to achieve significantly better efficiencies than diesel engines of the same output while complying with the coming round of tightened emissions limits. As with diesel engines, high pressure turbocharging will make an important contribution in the further progress of gas engines. In this report ways will be examined by which the full potential of gas engines can be exploited using high pressure turbocharging. ABB Turbo Systems Ltd (ABB) is making its own contribution to the improvement of the performance and emissions behaviour of gas engines through its own studies, close cooperation with leading manufacturers of gas engines and by making available suitable products. 2

3 2 A Short History of the Gas Engine In the area of large engines, since the 1930 s many diesel engines were also offered in gas engine versions for specific applications. The engines concerned were, in part, pure gas engines with stoichiometric combustion and spark plugs which, with turbocharging, were developed to mean effective pressure levels of 10 to 12 bar. Efficiencies were lower than with the corresponding diesel engines but higher than those of petrol engines. This was thanks to the influence of size, improved knock behaviour (methane has an octane number of about 130) and configuration for stationary operation with lower throttle losses. The main advantages were the cost and availability of the fuel, clean combustion and the possibility of recovering a great deal of heat from the hot exhaust gases. Parallel to this development, engines in the "dual fuel" category were developed which can be used in both gaseous and liquid fuel modes. The fuel injection system is configured for two operating modes: pure diesel operation up to full load and gas operation with, typically, injection of 3 to 7% of diesel fuel as an ignition pilot. The operating values of these engines lie between those of diesel and gas engines. In order to guarantee both knock-free operation and flammability of the diesel fuel compression ratios lie in the area of 11. The power density of dual-fuel engines has, in the meantime reached values of up to 20 bar mean effective pressure (pme). With pure gas engines significant progress was achieved in the 1990 s via the development of new combustion technologies, especially for lean burn operation. In this way mean effective pressures were raised to 14 to 16 bar in combination with improved efficiency. Today s engines profit from the introduction of the Miller process. This allows mean effective pressures of over 20 bar to be achieved. At the same time engine efficiencies are similar to equivalent diesel engines and in some cases even higher. The content of this report is the further development of the gas engine up to extreme Miller valve timings under the application of high pressure turbocharging. In this way it will be possible to further increase performance and eliminate the deficits versus diesel engines even in the area of power density. 2.1 Ignition System (Fig 1) Classic, spark ignited Otto cycle ignition is not limitlessly scalable. The ignition energy and the durability of the spark plug become insufficient as engine size increases. It was, however, possible to widen the application range of spark plugs by the use of pre-chambers. The prechamber can be supplied with additional gas in order to further improve the conditions for combustion (pre-chamber mixture enrichment). 3

4 Spark ignition Gas Scavenged Pre-chamber, Spark Ignition Pilot Fuel Ignition. Figure 1: Possible ignition systems for gas engines Ignition using diesel injection still remains a valuable alternative for large engines. Common rail technology has allowed a reduction in pilot injection to below 1% of the total energy (Micro pilot injection). Diesel injection can take place both in an open combustion chamber or a pre-chamber. 2.2 Gas Admission For spark ignited (Otto) engines using liquid fuels, the development of the carburettor has lead to indirect injection and then direct injection into the combustion chamber. In gas engines, by contrast, the energy needed for compressing the fuel plays a much more significant role. In the extreme case it can involve up to 15% of engine power. For this reason small gas engines, above all, are operated with a central atmospheric gas mixer. This also allows economic operation on weak gases which need to be mixed with the air in not inconsiderable quantities. Gas admission in the inlet port, generally using timed dosing valves, is widespread on large engines. This has the advantage that gas exchange and power control can be achieved in a similar way to diesel engines. Scavenging of the combustion chamber with air is possible and the variation in gas quantity within the permissible λ v fluctuation tolerance can be rapidly achieved. Direct injection into the combustion chamber is very rare on gas engines. An example for this class is represented by so-called diesel-gas engines. In this category diesel and gas are injected simultaneously, which can be achieved via two injectors or an injector with two concentric nozzle rings. These engines should be classed as genuine diesel engines with every advantage and disadvantage. However, the handling of gas at high pressure (250 bar) represents a further challenge. 4

5 2.3 The Layout Diagram The mean effective pressures achievable on gas engines is limited by the knock limit and the stability limit for very weak mixtures, (Figure 2).These two boundary areas leave a free area in which both mean effective pressure and air:fuel ratio can be increased. This is the area occupied by all contemporary gas engines in which high efficiencies and low NO x emissions are possible. Via the use of the Miller Process it is possible to shift the knock limit further upwards and thus further increase mean effective pressure and engine efficiency. Figure 2: Influence of excess air ratio (λ v ) on performance, emissions and limits of the gas engine [1]. 3 Thermodynamic Principles It has already been demonstrated [2] that the Miller Process can make a large contribution towards increasing the efficiency of combustion machines via temperature reduction. The analysis of the working process with perfect air does not help in determining the achievable potential (Figure 3). Only the red curves in the diagram were calculated using perfect air, for further curves the working process was calculated using gas properties reflecting the current state-of-the-art [3]. For combustion air and 0.70 η t ε = 16, λ V = 2.2, p mi = 30 bar ε = 14, λ V = 2.1, p mi = 30 bar P max [bar] 350 a reference, hydrocarbon based fuel the temperature influence of the chemical species was taken into account (ideal gas model). The influence of pressure (real gas model), as well as dissociation at high gas temperatures both of which would further reduce process efficiency - were not taken into account. These influences are considered negligible. While the black curve is based on a conventional process, for the green curves a Miller Process having an in-cylinder expansion ratio of 2 was assumed. This reduces the starting temperature from 80 C to 17 C Perfect air Ideal gas Ideal gas, Miller Figure 3: Ideal cycle thermal efficiency comparison. 5

6 The diagram shows that the efficiency level increases by 2 to 3 percentage points. The ideal Miller Process may result in a loss of around 1.5 % but this is more than compensated by an increase of up to 5 % due to the temperature reduction. This considerable gain results from two roughly equal contributions: the more favourable high pressure portion and a gas exchange loop with a larger positive area. The first set of curves (solid lines) refers more to diesel engines with a high compression ratio ε. For gas engines it is to be expected that the ε-area will be located in the range 13 to 15 (dashed curve). The air:fuel ratio is only slightly smaller than for a diesel engine. 3.1 Gas Exchange The reason why the process efficiency can be greatly increased via the gas exchange phase is explained in Figure 4. The p-v diagram on the left shows the idealised gas exchange process with conventional turbocharging. An increase in charging efficiency from 65% to 75% gives the possibility of slightly increasing efficiency, using more piston work in the gas exchange phase. As an alternative the pressure difference over the engine can be left unchanged resulting in more energy being exploited from a turbine (turbocompounding). 14 p [bar] Cylinder process Cylinder process Air compression Air compression Gas expansion Gas expansion η turbocharging - = T-Eff. 65% 65% Gas expansion Gas expansion - T-Eff. 75% η turbocharging = 75% 14 p [bar] Cylinder process Cylinder process Air compression Air compression Gas expansion Gas expansion η - T-Eff. 65% turbocharging = 65% Gas expansion Gas expansion - T-Eff. 75% η turbocharging = 75% 8 Gain in piston work 8 Gain in piston work 6 Gain with turbocompound Gain with turbocompound V/V d V/V d 10 Figure 4: Possibilities for converting turbocharging efficiency in power output. 6

7 The diagram on the right shows the situation with the much higher charging pressure appropriate to the Miller Process. The achievable gain via piston work has increased massively. By contrast the potential for turbo-compounding is no longer available: if the efficiency of the turbine is taken into account, the conclusion is that for 4-stroke engines with extreme Miller valve timings the advantage of turbocompounding is not available. The high pressure ratio via the engine can be exploited directly on the engine without the additional expense of turbocompounding. In practice the charging efficiency is reduced on gas engines by the control equipment. For this reason the potential gain of the improved gas exchange is less than for diesel engines. 4 Possibilities for Optimisation In order to achieve the practical realisation of these theoretical considerations different parameter variations were calculated with the help of an engine model. The engine concerned is a gas engine with atmospheric gas mixer ( premix ). The pressure ratio of the turbocharging system and the Miller effect were varied under conditions of constant mean effective pressure. In doing this, the boundary conditions from Table 1 were maintained. Table 1: Boundary conditions for the parameter variation. 1-stage 2-stage pme 24 bar 24 and 30 bar Air excess ratio constant Compression ratio Adjusted for constant compression temperature Valve timing late and early Miller early Miller Ignition timing constant Turbocharging efficiency Standard scalable characteristics derived from existing components Mixture cooling temperatures Adjusted with 10 C margin against water condensation The results (Figure 5) allow the following conclusions to be drawn: With single stage turbocharging and a mean effective pressure of pme = 24 bar the best results can be achieved at pressure ratios between 5 and 6. The differences between early and late Miller valve timing are marginal. With 2-stage turbocharging and a pressure ratio of 7 engine efficiency can be additionally improved by about 3.5% compared to single stage high pressure turbocharging. At increased Miller effect the use of early Miller is clearly more favourable. 7

8 Engine efficiency vs. pressure ratio 6 ηeng [%] stage_24 bar_early Miller 1-stage_24 bar_late Miller 2-stage_24 bar_early Miller 2-stage_24 bar_late Miller 2-stage_30 bar_early Miller η [-] Turbocharging Efficiencies π C Figure 5: Efficiencies from the simulations according to table 1. At pme = 30 bar only 2-stage turbocharging comes into consideration: the optimum pressure ratio lies at around 9. The uncertainties in the simulations lie principally in the combustion and knock behaviour of the engine. In a comparison of Miller early to Miller late, the influence of air movement on combustion cannot be quantified [4]. For this reason combustion was assumed to be constant. In addition the uncertainty relating to heat transfer is relevant, since this exerts a decisive influence on the effective compression temperature. These uncertainties can shift the effective balance in the area of moderate Miller valve timings in favour of Miller late; with the extreme Miller valve timings, however, a large mixture mass must be expelled during a large part of the compression phase. The corresponding energy losses are so large, that the advantage of Miller early can no longer be called into question. The efficiency level at pme = 30 bar is based on the extrapolation of various engine parameters. Especially conservative is the assumption that compression temperature must be reduced via the reduction of the engine compression ratio in order to maintain the same distance to the knock boundary as for pme = 24 bar. The right hand diagram shows the efficiencies of the turbocharger (η TC ) and the turbocharging process (η T ) [5]. The difference derives from the losses in the charging system and is large in the case of gas engines, especially since control organs are needed. In any case, the significance of system losses decreases as pressure ratio increases stage_24 bar_etatc 1-stage_24 bar_etat 2-stage_24 bar_etatceq 2-stage_24 bar_etat 2-stage_30 bar_etatceq 2-stage_30 bar_etat π C 8

9 4.1 Valve Overlap Engines with timed gas admission in the inlet port can profit from scavenging in the same way as diesel engines. A scavenged combustion chamber helps to hold process temperatures low. This has a positive effect on the thermal loading of components and the knock resistance of the engine. On "premix engines scavenging is not desired since it leads to increased emissions of hydrocarbons (HC) and lower engine efficiency. Nonetheless attention should be paid to valve overlap since the optimum between over scavenged gas quantity and increased gas exchange losses shifts in proportion to the charging pressure level and pressure differences over the cylinders. 5 Control of Large Gas Engines Over a wide operating range the turbocharged diesel engine can only be controlled on the basis of fuel injection quantity, since it can function in a wide λ V window. Gas engines, by contrast, always need control of gas quantity and λ V, i.e. the charging pressure must usually be controlled. Possible control interventions for single stage turbocharging are: Throttle valve (TV) Compressor bypass (BV, a compressor side wastegate for recirculation) Wastegate (WG, turbine side) Variable Turbine Geometry (VTG) Variable Valve Timing (VVT) Variable Compressor Geometry (VCG) These possibilities are represented schematically in Figure 6. Bypass Throttle VCG Receiver Intercooler Gas VTG Wastegate Figure 6: Single stage turbocharging Control possibilities 9

10 With the introduction of 2-stage turbocharging the possibilities increase, since there are up to three possibilities for different measures; e.g. the compressor bypass can affect the high pressure stage, the low pressure stage or both. Figure 7 shows schematically the different control circuits. Since no practical experience is available, the different control variants were analysed using simulation. HP+LP Bypass HP Bypass LP Bypass VCG Throttle VCG Receiver Intercooler HP TC Intercooler LP TC Gas VTG HP Wastegate Figure 7: 2-stage turbocharging Studied control possibilities As the first priority only the simpler variants were assessed. On the one hand better control behaviour is expected from 2-stage turbocharging, on the other hand complexity should not be increased unnecessarily by the use of over complicated solutions. 5.1 Stationary Simulation The following boundary conditions are fundamental to the design of gas engines: Charging pressure is defined for every operating point (pme und λ V ). Variations in efficiency are only to be expected via variations in exhaust gas pressure. The gas side should be so configured that the air side can realise a higher pressure ratio (the smaller the turbine area, the greater the control reserve) As a consequence all control variants with a constant configuration on the gas side are equivalent for stationary operation. To achieve efficiency optimisation control options with wastegate or VTG are thus of interest. The wastegate around the high pressure turbine allows a relatively broad variation in exhaust gas pressure and turbine power without serious efficiency penalties. For this reason VTG was not, initially, taken into account. 10

11 New findings can be gained from the simulations (Figure 8). All control variants which are only effective in the domain of engine plus high pressure turbocharger are virtually equivalent for the low pressure turbocharger. This exerts high requirements on the map width for the low pressure compressor since the operating curves are flat. At load rejection it would be very difficult to avoid surging at the low pressure compressor πv tot/tot HP+LP bypass HP bypass / Throttle valve πv tot/tot HP+LP bypass HP bypass / Throttle valve LP bypass LP bypass Variable valve timing (VVT) Variable valve timing (VVT) HP wastegate HP wastegate 3.0 HP wastegate + HP+LP bypass 3.0 HP wastegate + HP+LP bypass η * sv η * sv HP Compressor LP Compressor V298 [m 3 /s] V298 [m 3 /s] Figure 8: Operating lines in the compressor maps with different control options. The first phase of the investigation permitted two favoured solutions to be identified: A bypass over both stages as the best solution for moderate requirements (grid parallel operation) A combination of high pressure wastegate and bypass over both stages for more stringent dynamic requirements and better engine efficiencies (stand alone operation). 5.2 Transient Simulations The control variants dealt with above were simulated in two cases of transient operation: Grid parallel operation, i.e. load ramp from 0 to 100% in 90 seconds during which engine speed (grid frequency dependent) remains constant. Stand alone operation, i.e. load steps of 4.8 bar; at the same time the specified speed is constant but the actual speed results from the engine dynamics. In both cases engine behaviour during load rejection was investigated. 11

12 Figure 9: Speed and pme curve over time for load acceptance in island operation mode. The results show the following picture: Figure 10: Influence of speed dependent mixture control. Since the engine with minimal control reserves is configured for maximum efficiency the control variants differ only slightly as long as the turbine side remains constant (Figure 9). In practice the curves differ only in terms of control behaviour after the recovery period. The reduction in turbine area and the λ V control (mixture enrichment at load imposition) generally showed the greatest effect (Figure 10). The throttle valve alone causes large variations of the operating point in the compressor map (Figure 11). The oscillation of the governor following attainment of the setpoint speed could cause the compressor to surge. This can be avoided by integrating a load or pressure dependent limit into the control software. All bypass variants allow stable governing but only the variants encompassing both stages allow both compressors to exploit their optimum range. 12

13 πv tot/tot Transient operation Steady state operation πv tot/tot Transient operation Steady state operation η * sv η * sv HP Compressor V298 [m 3 /s] LP Compressor V298 [m 3 /s] Figure 11: Operating lines in compressor maps with throttle valve control. The high pressure variant (Figure 12) results in a good dynamic response in the high pressure stage but the stationary operating curve results in the performance maps of the high pressure compressor remaining flat and prevents the optimal exploitation of that stage. Control using only the low pressure stage reduces the feedback between the control device and the cylinder, and thus the response of the system to load changes is somewhat sluggish πv tot/tot Transient operation Steady state operation πv tot/tot Transient operation Steady state operation η * sv η * sv HP Compressor LP Compressor V298 [m 3 /s] V298 [m 3 /s] Figure 12: Operating lines in compressor maps with HP-bypass control. 13

14 πv tot/tot Transient operation Steady state operation πv tot/tot Transient operation Steady state operation η * sv η * sv HP Compressor LP Compressor V298 [m 3 /s] V298 [m 3 /s] Figure 13: Operating lines in compressor maps with LP&HP-bypass control. Control using a bypass past both compressors (Figure 13) proved, as expected, to be optimal. Both compressors can be operated in their optimum range (this is not the case in Figure 13 since in the investigation it was desired that all control systems be tested with the same specification). The response of the system and controllability are wellbalanced. It was possible to simulate the wastegate control system but it proved difficult to stabilise due to the longer response times 25 pme of the system. The wastegate thus 20 [bar] only comes into consideration as a feed-forward control and a slow feed-back control in combination with a fast, mixture side control system (throttle valve or bypass) An example of such a control was used for the simulation in Figure 14. During a load increase above a threshold value the wastegate valve is closed immediately so that load acceptance can take place as quickly as possible. Thereafter the bypass valve takes over engine output control. After reaching a quasi stationary state the wastegate is progressively re-opened until the bypass valve has reached its set point. prec pti [bar] IndDK IndBYP 0.5 IndWG [-] time [s] neng [%] Figure 14: Load acceptance with bypass and wastegate control. 14

15 5.3 Variable Compressor Geometry Regulating the swirl at the compressor intake is a well known method of improving the load acceptance of a single stage turbocharged spark ignited engine [6]. The potential of this control concept was thus also examined in relation to 2-stage turbocharging using simulation. There are three possibilities: Pre-swirl control at the intake to the low pressure compressor Pre-swirl control at the intake to the high pressure compressor Pre-swirl control at the intake of both compressors In both of the first cases it was observed that the compressor concerned delivers a very rapid increase in pressure ratio via swirl reduction following a variation in load. There is, however, not sufficient turbine energy available at the other compressor in order to deliver the increased mass flow. It therefore reacts with a reduction of pressure ratio. In sum, there is no noticeable increase in charging pressure (Figure 15). πv,hp 2.0 π T,HP [-] 1.5 πv,lp 2.5 πt,lp [-] If the pre-swirl control system is applied on 0 both compressor stages the two compressors react in the direction of the desired rapid time [s] increase in pressure ratio following the load. increase. However, the increase in revolutions in stationary operation, which corre- HP-Compressor. Figure 15: Load acceptance with VCG control on sponds to a reserve, is spread across two compressors and thus has a correspondingly lower effect. In sum these simulations demonstrated no clear advantage for the use of pre-swirl control with regard to 2-stage turbocharging. prec pti [bar]

16 5.4 Control with Variable Valve Timing The optimum valve timings for gas engines with 2-stage turbocharging result in a volumetric efficiency in the area between 0.5 and 0.6. From the standpoint of the turbocharging system, the inlet valve is then a throttle point which reduces the pressure between the receiver and the cylinder by almost half. By means of a flexible shifting of valve timings using a fully variable valve train, it thus seems that an enormous control reserve can be accessed. The following negative effects should also be taken into account: Reduction of the throttling effect of the valve results in a mismatching of the turbocharger system. Turbocharging pressure falls and the pressure difference over the cylinder collapses. Process efficiency deteriorates, due to disturbance in both gas exchange due to the reduction in p and the high pressure portion via the reduced Miller Effect. An additional difficulty results in that, via the reduced Miller Effect, the distance to the knock boundary is reduced. A change in ignition timing can become necessary, which further reduces engine efficiency. As a measure for improving the transient behaviour of the engine a shift in the inlet valve closure point was investigated. The simulation (Figure 16) showed that the improvement in load acceptance is more pronounced and that the effect is comparable with a richening of the air-fuel mixture. 5.5 Single and 2-stage High Pressure Turbocharging The transition from single to 2-stage turbocharging will take place on gas engines in variable valve timing compared to the reference Figure 16: Load acceptance with enrichment und the range between pme 23 and 25 bar. The case. system configuration of a two stage charging system is more complex. 16

17 A comparison of the two turbocharging concepts is difficult since the application ranges do not overlap. The optimum engine and turbocharging configuration for the same engine output would differ greatly, further hindering comparability. In order to undertake a comparison of load acceptance behaviour the problem was made easier by applying the same load steps to both engines ( pme = 4 bar). With single stage turbocharging pme = 20 bar is reached in five steps while with 2-stage turbocharging an additional step to pme = 24 can be applied The results of the simulation show (Figure 17), that load acceptance with 2-stage turbocharging is clearly better. The main reason lies in the fact that in transient engine Figure 17: Load acceptance with 1- and 2-stage turbocharging. operation, primarily only the high pressure turbocharger determines the reaction of the charging system while the low pressure turbocharger adjusts itself to the new conditions with a short delay. The moment of inertia of the high pressure turbocharger is considerably smaller than that of the single stage turbocharger and for this reason the engine with 2-stage turbocharging responds better to load variations 6 Requirements on Turbocharger Design In the previous chapter the possible contribution of 2-stage turbocharging to the further development of gas engines was examined. In order to access this potential, further developments will be necessary in the turbocharger components and especially in the turbochargers themselves. For the pme levels up to around 24 bar turbochargers are required for pressure ratios of 5 to 6 with high efficiencies (η TC 0.65). ABB is developing the new A100 turbocharger generation especially for these applications. In particular, the turbocharger family A100-H with radial turbine is suitable for applications on gas engines. The first sizes of A100-H turbochargers were introduced to the market in Figure 18 shows the extension of the application range of the A100-H over that of the established TPS turbocharger family. The A100 sets a new standard for the application limits of an efficient single stage high pressure turbocharging system with pressure ratios up to 5.8 [7]. 17

18 A125 TPS44-F Figure 18: A100-H Pressure ratios and volume flows As already demonstrated in 2007 [2], single stage turbocharging is not the correct technical or commercial solution for the pme range over 24 bar. In this range 2-stage turbocharging systems are needed which fulfil the requirements, especially with regard to the turbocharger stages, which can be deduced from simulations and results of first on-engine trials. The turbocharger for the high pressure stage must fulfil special requirements. Pressure level and power density have increased considerably compared to an atmospheric turbocharger, and must be taken into account in the design of the shaft, bearings and housings. The performance requirements for the components, by contrast, move more in the direction of more moderate pressure ratios but with high specific flow capacity and wide operating maps. And finally, efficiency should be very high, since especially in the case of the high pressure ratios of 2- stage turbocharging every increase in turbocharging efficiency makes an increase in engine efficiency available. For gas engines the present study has shown that the division of pressure ratios between the stages π C,ND /π C,HD in the range 1.6 to 1.8 appears appropriate. This results from requirements regarding the operating map width of the low pressure compressor, which comes more or less into play dependent on the control concept used. The resulting limitations on the pressure ratio π C,ND help to reduce the problem of carbonisation of oil residues from the intake of engine blow-by. ABB has likewise also begun developing new products for the specific requirements of two stage turbocharging. 18

19 7 Conclusion and Outlook High pressure turbocharging gives developers of gas engines the possibility to increase the power and efficiency of their engines beyond presently known limits. The application of the Miller Process with volumetric efficiency between 0.5 and 0.6 opens the way to accessing the potential of the gas engine. Via the reduction in process temperature is will be possible to shift the knock limit which limits the output and efficiency of gas engines to a point where high mean effective pressures and compression ratios are achievable. As well as the enhancement of the high pressure part of the engine process, in combination with 2-stage turbocharging the Miller Process brings a further considerable improvement in gas exchange: engine and turbocharger processes are so well tuned that the advantages of combined energy utilisation can be exploited directly on the engine without the needs for complex additional equipment (turbocompounding). High pressure turbocharging permits that the mean effective pressure level of gas engines can be raised to around 24 bar using single stage turbocharging and far beyond this value with 2- stage turbocharging. As technology leader, ABB is placing the products needed to exploit this potential at the market s disposal in good time. Only by use of 2-stage turbocharging can the full development potential be realised. This however unavoidably involves higher complexity but holds the promise of configuring the control system more flexibly and efficiently. Extensive simulations have shown which control systems promise an improvement in the load acceptance capability of gas engines. As is already known with regard to spark ignited engines in the automotive sector, there are no allembracing solutions which, on their own, promise the optimum potential. In the face of growing requirements, ever more combinations of different control options will be developed and applied. Among other capabilities, ABB Turbo Systems has longstanding experience in the simulation of turbocharged combustion engines. With the help of simulation this experience is being used to formulate system requirements into product objectives, so that the right products are available to the market in a timely manner. An important aspect in this is close cooperation with the engine builder so that the development of engine and turbocharging system converge into a joint goal i.e. total system optimisation. 19

20 References / Literature [1] POWER NEWS, Wärtsilä Diesel Group, Customer Journal 1993 [2] Codan, E. & Ch. Mathey, 2007, Hochdruckaufladung bei Grossmotoren, 12. Aufladetechnische Konferenz, Dresden (D). [3] McBride, B. J., M.J. Zehe & S. Gordon, 2002, NASA Glenn Coefficients for Calculating Thermodynamic Properties of Individual Species, NASA/TP [4] Schutting, E., A. Neureiter, C. Fuchs, T. Schatzberger, M. Klell, H. Eichlseder & T. Kammerdiener, 2007, Miller- und Atkinson-Zyklus am aufgeladenen Dieselmotor, MTZ 06/2007, [5] CIMAC - Conseil International des Machines à Combustion, 2007, Turbocharging Efficiencies - Definitions and Guidelines for Measurement and Calculation, Recommendation Nr. 27, [6] Lang, O., K. Habermann & M Wittler, 2006, Verbesserung des Betriebsverhaltens von Turbomotoren lurch Verdichtervariabilitäten, 11. Aufladetechnische Konferenz, Dresden (D). [7] Wunderwald, D., T. Gwehenberger & M. Thiele, 2008, Neue Turboladerbaureihe A100-H für die einstufige Aufladung schnelllaufender Motoren, MTZ 07-08/2008, pp