Honda has successfully brought a range of gasoline engines with turbocharging and direct-injection as the VTEC Turbo series to the market since 2015. Introduced to the European market in spring 2017, the newly developed 1 l three-cylinder engine is the smallest product in the series so far. Downsizing and state-of-the-art friction reduction technologies have enabled the 10th generation Civic model to improve fuel economy by 26 % compared to its predecessor.
OVERVIEW OF THE NEW THREE-CYLINDER ENGINE
A new 1 l three-cylinder engine with turbocharging and direct-injection has been developed as a series of Honda VTEC Turbo engines [1, 2, 3], enhancing both fuel economy and driving performance. A rapid combustion concept realised using high tumble intake port design, combined with a side-mounted multi-hole injection system and an intelligent turbocharger boosting system with an electronically controlled wastegate, are common to all VTEC Turbo engines. In addition to dual-VTC (Valve Timing Control System), VTEC (Variable Valve Timing and Lift Electronic Control) has been used in the intake valve system in order to realise an Atkinson cycle for reduced fuel consumption in low valve lift mode. Also, state-of-the-art technologies for friction reduction have been adopted to achieve best-in-class fuel consumption characteristics, improving fuel economy by 26 % over the previous engine in the New European Driving Cycle (NEDC) mode, (1).
SPECIFICATIONS & ENGINE STRUCTURE
This new 1 l turbocharged gasoline direct-injection engine has been developed as a successor to the 1.5 to 1.8 l class NA (Naturally Aspirated) engines. (2) shows the newly developed engine’s major specifications in comparison to those of 1.5 and 1.8 l NA engines, for reference. Basic engine specifications such as bore diameter, bore spacing, crankcase height, cylinder offset, valve system configuration and injector location are derived from the 1.5 l NA engine for manufacturing efficiency and parts commonality.
(3) shows cross and longitudinal section-views of the engine, as well as the coolant flow inside the engine. The crankcase is made of aluminium alloy using a high-pressure die-cast method, with cast iron liners moulded in. The coolant flow is divided into two flows in a sub-jacket created in front of the crankcase. One flow enters the crankcase water jacket and eventually goes up to the cylinder head. Another flow goes up directly to the cylinder head. This allows the balance of flow quantity between the crankcase and cylinder head to be easily optimised through passage hole sectional area adjustment of the cylinder head gasket, thus ensuring appropriate thermal conditions.
Cooling of the combustion chamber was improved through a combination of the above-mentioned coolant flow and a vertically split two-piece water jacket in the cylinder head, which makes cold coolant flow across the lower water jacket of the cylinder head enabling the combustion chamber of each cylinder to be evenly cooled. The two-piece water jacket also cools the exhaust port efficiently from both the top and bottom.
The cylinder head has an integrated exhaust manifold in order to minimise heat mass. This contributes to a quick activation of the catalyst and eventually leads to lower exhaust emissions and fuel consumption for catalyst heating control during warm-up. At the same time, the exhaust gas temperature has been reduced in full-load conditions, which allows the use of standard-grade heat- resistant material for the turbocharger. This also helps to reduce thermal stress in the cylinder head itself, enabling a reduction in fuel enrichment in the high-power area and contributing to improved fuel economy during such use.
Like the previously released 1.5 and 2 l four-cylinder VTEC Turbo engines, the combustion concept of the new engine involves rapid combustion. One of the challenges of this engine was to achieve the high targets defined for performance and exhaust emissions with a relatively small bore diameter of 73 mm and side mounted direct injection.
The combustion chamber configuration is derived from the existing, naturally aspirated engine in terms of valve angle, spark plug angle and injector location. However, the shape of the intake port and piston top were designed to increase the tumble ratio of the new engine for rapid combustion, and intake valve diameter was reduced to avoid fuel adhesion. The in-cylinder flow simulation results in (4) show that the intake air flows along the pentroof and shallow dish shape of the piston top, and that a much stronger tumble flow is created compared to the reference 1.5 l naturally aspirated engine.
The tumble flow of the engine is maintained during the compression stroke and finally converted to higher turbulent kinetic energy before top dead centre, which is important for enabling higher combustion speeds promoted by fast flame propagation. The burn delay characteristic shows the positive influence of reinforced tumble flow, which was confirmed in actual firing testing as well as in calculations. The burn delay characteristic is at the lower end of the scatter band of existing engines as shown in (4).
CRANK DRIVE SYSTEM
(5) shows the overview and configuration of the crank drive system. Vibration rooted in the first-order inertia couple is a phenomenon that is a challenge to avoid in a three-cylinder engine. However, a balancer-less system was used to avoid increased friction and weight. As a countermeasure for this vibration, the crankshaft reciprocating mass balancing ratio was optimised though counterweight configuration in order to reduce the vertical vibration at the engine mount points, which makes a stronger contribution to body vibration than horizontal vibration. Different balancing ratios were compared when developing this engine and finally a balancing ratio of 75 % of reciprocating mass was selected as it showed lower vertical vibration.
In addition, exceptionally narrow diameters for the main journals and crankpin journals were chosen to reduce the rotational friction of the engine – a choice that was enabled by using strong nitride steel material. The comparison of reciprocating parts FMEP (friction mean effective pressure) in (5) shows that the new engine has almost minimal friction characteristics in the respective scatter band. On the other hand, reductions in journal diameter tend to increase torsional vibration, and hence a worsening of noise and vibration of the engine. To cope with this problem, crankshaft damper ring inertia mass and Eigen-frequency was carefully optimised.
The piston has a cooling oil gallery, reducing the piston crown temperature by more than 30 K with oil jet cooling. This contributes to an improvement in knocking behaviour and enables a higher compression ratio and optimised ignition timing, resulting in a 0.5 % fuel efficiency increase in NEDC mode. Additionally, this piston crown temperature decrease contributes to reduced wear of the piston ring groove, so surface treatment like anodising was not necessary even in such a high power density engine. The piston-cooling jet is equipped with a built-in check valve and set to operate only when the variable displacement oil pump is switched to high pressure mode; it is controlled electronically by a solenoid valve depending on the operating condition.
VALVE TRAIN SYSTEM
(6) shows the valve train system configuration as well as the valve lift curves in different regions of the engine map. The valve train uses VTC for both intake and exhaust, with VTEC rocker arms on the intake side that can switch valve lift and timing between low lift mode and high lift mode. In low lift mode operation, the valve lift and timing are controlled to realise an Atkinson cycle with early closure, and the rocker arms are switched to high lift mode operation when high output is demanded. Thus, fuel economy can be improved under low-load conditions while still delivering high torque under high-load conditions. This operation strategy reduces pumping losses by selecting early valve-close timing under low-load conditions. A benefit of 5 % in BSFC (Brake Specific Fuel Consumption) has been achieved by the low lift mode resulting in a 2 % improvement in vehicle fuel consumption (NEDC mode). Optimal VTC and VTEC control is performed depending on operating conditions such as engine load and engine speed, as shown in the VTEC rocker arms switching map in (6).
A three-cylinder engine with a short overall length has limited layout space, especially when installing a turbocharger, so an efficient layout for the VTC system is needed. The new engine uses a VTC structure that positions the VTC control solenoid and oil control valve aligned to the axis of the camshaft. This simplifies the oil circuit, and the effect of the built-in check valve helps to deliver a high responsiveness of the VTC system.
TIMING BELT DRIVE SYSTEM
The engine drives the camshaft and oil pump using timing belts located inside the engine, (7). As the timing belts are located inside the engine, the belt noise is lowered through insulation, and engine friction is also reduced by 1.8 %, thanks to the elimination of sliding guide parts. As a consequence, a 0.6 % fuel consumption improvement was observed in NEDC mode compared to a conventional timing chain. Timing belt materials were developed with specifications that provide sufficient durability for use in the oil environment. A comparison of the belt in oil durability with those of a conventional timing belt in air are also shown in (7), and it can be observed that the belt has much higher durability performance in oil conditions even when compared with conventional belt in air conditions.
VARIABLE DISPLACEMENT OIL PUMP
A variable displacement oil pump is used to reduce engine friction and heat loss. (8) shows the sectional view of the oil pump, which is located in the lower interior of the engine, and is driven by the crankshaft via a timing belt in oil. Variable oil pump capacity is achieved by using a structure that changes the eccentricity by swinging the cam ring forming the outer ring of the vane pump. Oil pressure applied to the chambers above (chamber 1) and below the cam ring (chamber 2) controls this eccentricity. An electronically controlled solenoid valve is used to switch the oil pressure between two levels in accordance with the engine load and speed, thereby maintaining the target oil pressure allowing it to fall. Also, a pilot valve was adopted to regulate oil pressure at the target value. This helps to prevent the oil pressure from rising excessively even during warm-up conditions at low oil temperature and high viscosity in low pressure mode. Additionally, the piston cooling jets are deactivated to minimise heat loss.
Oil pressure is switched to high pressure mode when the engine operating conditions exceed a certain threshold, as it is necessary to reduce temperature of the pistons through a piston cooling jet, or the temperature of the con-rod bearings at high load or high engine speed. The analysis of fuel efficiency benefits of the variable oil pump compared with conventional oil pump without oil pressure control shows that, in the early stage of NEDC mode, friction reduction is the main source of the benefit, whereas in the latter half heat loss reduction contributes considerably to the benefit. Through reducing friction and the heat loss of piston cooling jets, NEDC mode fuel economy is enhanced by a total of one per cent.
An electronically controlled thermostat, (9), was used to improve fuel economy through friction reduction and to ensure engine thermal reliability. The valve opening temperature for the wax element in the electronically controlled thermostat located at the outlet of the engine was set to 103 °C, 20 °C higher than that of a conventional mechanical thermostat. During low-load engine operation, the coolant and the oil are also kept at high temperatures, leading to mechanical friction reduction. The piston in the wax element contains a ceramic heater, and the thermostat can be opened by heater operation at a lower coolant temperature than the temperature set for the wax element.
As an effect of the electronically controlled thermostat system, it can be observed that engine friction clearly declines depending on the increase in coolant and oil temperature. As a result of the reduced engine friction, fuel consumption has been reduced by 0.6 % in NEDC mode. During high load engine operation, the heater is operated and the coolant is kept at a lower temperature, in order to ensure engine thermal conditions as in a conventional thermostat system. Furthermore, in order to prevent overheating in the event of an increase in transient load caused by sudden acceleration, the thermostat is guaranteed to respond quickly enough.
OUTPUT & EFFICIENCY
The maximum output of 95 kW is comparable to that of a 1.5 l NA engine, and the maximum torque of 200 Nm is higher than that of a 1.8 l NA engine. It is notable that the maximum torque is already generated at 2,250 rpm and maintains a 90 % level up to 4,500 rpm, enabling excellent acceleration behaviour in stop-go mode in urban driving. (10) shows the resulting BSFC map and the position of the engine within the scatter band for part load BSFC. As a result of the application of the technologies outlined, BSFC of 231 g/kWh could be achieved together with a broad range of excellent BSFC below 240 g/kWh. Using BSFC at 1,500 rpm and BMEP of 2 bar as representative points, the new engine has best-in-class BSFC among other engines [4, 5, 6].
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TOMONORI NIIZATO is Senior Chief Engineer in the Engine Development Division at Honda R&D Co Ltd Automobile Center in Haga-gun, Tochigi (Japan).
MITSUHIRO SHIBATA is Chief Engineer in the Gasoline Engine Development Division at Honda R&D Co Ltd Automobile Center in Haga-gun, Tochigi (Japan).
DR-ING MICHAEL FISCHER is Department Manager, Powertrain & Material Research at Honda R&D Europe (Deutschland) GmbH in Offenbach (Germany).
ULF REINSCHMIDT is Section Manager, Powertrain Engineering at Honda R&D Europe (Deutschland) GmbH in Offenbach (Germany).