**2. Materials and Method**

#### *2.1. Description of Engine Bench*

The tests were carried out on a three-cylinder spark-ignition engine. It is a turbocharged direct-injection engine. Its characteristics are specified in Table 1. It was braked by a dynamic HORIBA HT 250 bench managed by a SPARC control unit. The engine was instrumented with type K temperature sensors, 0–3 bar static pressure sensors, two Kistler type 4049B dynamic pressure sensors for the intake and the exhaust (frequency = 60 kHz), one HBM type 40 torquemeter (frequency = 10 kHz), three AVL type ZI33 in-cylinder pressure sensors (frequency = 150 kHz), and one ETAS ES430 air-fuel ratio sensor (frequency = 2 kHz).


**Table 1.** Engine main features.

The INCA Software (version 7.1.10/3) controlled the engine ECU.

All the sensors were connected to a National Instrument fast acquisition box. The STARS software from HORIBA made it possible to automatically control the dynamic bench, and it could create different cycles. The PR-L804 fan from Dynair cooled the engine radiator, and the Fumex FB110 fan cooled the intake heat exchanger. Both fans were controlled by STARS software, which modified their rotation speeds according to the speed of the vehicle during a cycle. This equipment created test conditions close to the real ones. The pollutant emissions CO and HC were measured with a 3200 CAPELEC device, the NOx emission with an ECM NOxCANt sensor, and particle emission with a Pegasor Particle Sensor (PPS). The setup is shown in Figure 2.

**Figure 2.** Experimental setup.

## **3. Blowby Simulation in Steady State**

*3.1. Building a Blowby Simulation Model*

As a first step, it is necessary to model the blowby gases of the engine described in Table 1. In this way, the Gamma Technology simulation software, GT-suite 2021, which has a "RingPack" module (specially created to model piston/cylinder/ring friction and to model the pressures and flows of gases circulating through the rings and piston grooves) was used. Only a single cylinder simulation model was created. Indeed, the objective was to create a simulation model of the phenomenon of oil scavenging by the blow-by gases according to the value of the final gap of the first two rings of one of the cylinders of the engine used. This single cylinder has all the technical characteristics of the engine used, such as the cylinder and piston diameters, the length of the connecting rod, the depth, height, and location of the piston grooves and the rings. Some of these characteristics are shown in Table 1. The input parameters required for this model were the temperature and pressure inside the cylinder during a cycle, see Figure 3. The latter was derived from experimental tests, while the temperature was approximated by the extensive literature on the subject. Indeed, temperature is a key factor, as it determines the proportion of an object that expands according to its physical properties and, in particular, its coefficient of expansion. Piston, ring and cylinder wall temperatures are proportional to the specific power developed by the engine and therefore depend on the engine load and speed [42]. On a single-cylinder diesel engine at 25% load and a speed of 3600 rpm, Abril, et al., [43] measured a compression ring temperature close to 200 ◦C during the combustion and expansion phases. On a single-cylinder, spark-ignition engine at 75% load and 3500 rpm, Thiel, et al., [42] determined that the temperature of the compression ring could reach 210 ◦C during these same phases. Husberg, et al., [44] found piston crown temperature variations of up to 75 ◦C between 25% and 50% load on a single-cylinder diesel engine. Taking all these elements into account, a coefficient of thermal expansion of the compression ring equal to 1 × <sup>10</sup>−<sup>5</sup> <sup>K</sup>−<sup>1</sup> with a maximum temperature of 250 ◦C at 6000 rpm in full load is assumed. The temperature of the compression ring changes linearly with the engine speed.

**Figure 3.** Simulation model development with GT-suite.

Model calibration:

The model calibration was based on experimental tests perform by a car manufacturer development department. These tests were carried out under full load and steady state on an engine strictly identical to the one used, see Table 1. The blowby rate was measured with an AVL 442 blowby meter. This device determines a flowrate from the measurement of a pressure difference generated by the blowby gases through a change in cross-section in the blowby gas intake duct. The results are displayed by the device in l.min−<sup>1</sup> for a pressure of 1000 mbar and a temperature of 25 ◦C [45], see Table 2.

**Table 2.** Blowby flow measurement for the engine at full load.


Study of the position of the compression ring under full load

Considering only the axial motion of the ring and applying the fundamental principle of dynamics, the equation is:

$$m\frac{d2h}{dt2} = F\_p + F\_i + F\_{fr} + F\_{oil} + F\_a \tag{3}$$

where *m* is the ring mass. The oil film pressure forces (*Foil*) and adhesion forces (*Fa*) are not significant compared to the intensity of the forces pressure forces (*Fp*), inertial forces (*Fi*), and friction forces (*Ffr*) [46]. The position of the ring in the piston groove (*h*) is therefore summarised by the variation of the forces shown in Figure 4. At low rotational speeds, i.e., between 1500 rpm and 3500 rpm, the pressure forces, *Fp*, are greater than the inertial forces, *Fi*, and therefore the compression ring remains in a low position in the piston groove throughout the cycle. From 4000 rpm, the inertial forces take advantage over the pressure forces *Fp*, which makes the ring move in its groove.

**Figure 4.** Set of Forces supported by the ring.

When the cylinder pressure is lower than the crankcase pressure, the ring may move in its groove. This is the case at 1000 rpm, since the pressure in the cylinder is lower than the atmospheric pressure when the exhaust gases are removed.

Study of the position of the sealing ring under full load

The same analysis can be made on this ring and, given the lower pressures involved, it moves axially at each cycle and whatever the rotation speed.

Simplification of the calibration

Usually, the endgap value of the sealing ring is much higher than that of the compression ring. In this case, it is twice as large (0.4 mm versus 0.2 mm). As a result, the blowby gas flow is mainly related to the position of the compression ring (Ring 1). Between 1500 rpm and 3500 rpm, the compression ring stays fixed on its base. In this speed range, the only possible path for the blowby gases is through the endgap of the compression ring. Therefore, the calibration was focused on this speed range and then extended to the entire engine speed range.

Analysis at 2000 rpm and full load

Figure 5 shows the simulation results at 2000 rpm and full load. The pressures between the rings are visible in the upper and middle graphs. The pressure above the compression ring (Land 1) is identical to the pressure in the cylinder. Its maximum is 77 bars at 400◦ CA.

**Figure 5.** Evolution of the pressures and axial positions of the piston rings at 2000 rpm and full load.

The maximum pressure between the compression and sealing rings (Land 2) is 29 bars. The pressure between the oil seal and scraper rings (Land 3) varies between 1025 mbar and 1045 mbar, while the pressure in the crankcase was stabilised at 1025 mbar. From this data and the inertial effects due to the speed of the piston displacement, it is possible to characterise the axial evolution of the compression and sealing rings, which is visible on the lower graph of Figure 5. As expected, the compression ring stays fixed on the lower part of the groove during the whole cycle, while the sealing ring is in the upper position of its groove during the end of the exhaust gas expulsion stroke and during the whole intake stroke. Outside this range, it stays in contact with the lower part of its groove. The evolution of the blowby flowrate can be seen in Figure 6. This one becomes significant from the end of the compression stroke to the end of the expansion stroke. A maximum was reached as 40◦ after TDC and amounted to 1.45 g.s<sup>−</sup>1. Finally, the cumulative blowby flowrate for one cycle reached 20.26 mg.cycle<sup>−</sup>1. This result is very close to the experimental tests where the cumulative blowby flowrate for this same operating point was measured at 19.90 mg.cycle−<sup>1</sup> (cf., Table 2). This simulation model allows us to better understand the ring dynamics and to estimate the evolution of the pressures that prevail in this ring/piston/cylinder area. Above all, thanks to the characterisation of the blowby flowrate over a complete cycle, it is possible to determine which engine strokes are preponderant in the phenomenon of oil sweeping towards the oil pan.

**Figure 6.** Evolution of blowby flowrate at 2000 rpm and full load.

Simulation over the entire engine load

From this model, and by adding different engine operating points performed on the engine bench, it is possible to draw a complete map of the blowby flowrate as a function of engine speed and engine load.

The result is shown in Figure 7. It is very clear that the flowrate is higher than 18 mg.cycle−<sup>1</sup> in a restricted area, i.e., at low speed and high load. Conversely, outside this zone, the flow rate remains below 12 mg.cycle<sup>−</sup>1.

#### *3.2. Analysis of Oil Flow Carried by Blowby Gases*

From these results and Equation (1), it is possible to determine the amount of oil swept by the blowby gases as a function of the position of the rings. To do this, it is necessary to estimate the amount of oil present in the Land 2. This quantity of oil depends on two factors: the first is the thickness of the oil film deposited by the ring, which varies by a few micrometers [47]. The second is the oil that is pumped out of the piston grooves and back into this area. This quantity of oil depends on the engine load and inertial effects, e.g., the rotation speed. As an example, Thirouard [48] measured an oil film height of up to 40 μm at 3500 rpm and half load on a single cylinder spark ignition engine and 20 μm at 2800 rpm and low load on a single cylinder diesel engine. In order to quantify the results, it is necessary to suppose an oil film height. The characteristics of the engine used are close to a diesel engine, therefore the assumption of an oil film of 20 μm will be retained. Figure 8 shows the amount of oil swept by the blowby gases at 1500 rpm for two endgap positions with an assumed oil film height of 20 μm. When the endgaps of the first two rings are opposite, i.e., at 180◦ (cf Figure 1a), the amount of oil swept by the blowby gases is maximum and can reach 20 μg.cycle<sup>−</sup>1.

**Figure 7.** Blowby flowrate as a function of engine speed and load.

**Figure 8.** Quantity of oil swept by the blowby gases as a function of the endgap position at 1500 rpm.

Whereas if the endgaps are close, i.e., only 30◦ apart (cf., Figure 1b), the maximum amount of oil swept by the blowby gases can be at best 6 μg.cycle−<sup>1</sup> for the same blowby flowrate. The same calculation can be made for each rotation speed, assuming an oil film height. It is therefore very clear that the oil concentration of the blowby gases is directly related to the position of the endgap and not only to its flowrate. The more oil is swept by the blowby gases, the less oil is returned to the cylinder via the backflow gases. This variable phenomenon linked to the rotation of the rings is difficult to predict and is the reason for the variations in particle emissions linked to the oxidation of the oil on strictly identical tests [49].

### *3.3. Influence of Blowby on Particulate Emissions at Idle*

At no load, i.e., when decelerating or idling, the pressure inside the cylinder is very low. The maximum pressure varies between 4 bars and 10 bars. Therefore, the blowby flowrate is very low and the oil sweeping phenomenon described by Thirouard, et al., [26] (Equation (1)) is minimal. In Figure 9, the pressures between rings and the ring positions evolving at idle speed over a complete cycle are shown from the simulation. The upper graph represents the evolution of the pressures above the compression ring and below. The evolution of the pressures between the oil seal and oil scraper rings visible in the middle graph is approximately equal to the pressure in the oil pan, i.e., equal to 0.98 bar. Compared to the graph in Figure 5 at 2000 rpm and full load, the difference is very significant. The maximum pressure reaches 8 bars for Land 1 and 3.5 bars for Land 2. The lower graph shows the evolution of the axial position of the first two rings in relation to their respective lower grooves. During the intake stroke and most of the compression stroke, the compression ring (Ring 1) stays on the upper axial part of its groove. At the end of compression and expansion strokes, the ring stays on the lower part of its groove. Then, at the end of expansion stroke and at the beginning of the expulsion of the exhaust gases, the ring oscillates axially to come into contact with the upper part of the groove and finally returns into contact with the lower part of the groove at around 540◦ CA and until the end of the cycle. The sealing ring (Ring 2) evolves with the same trend as the compression ring with an additional oscillation between 680◦ CA and 720◦ CA.

**Figure 9.** Evolution of pressures and axial positions of the piston rings at idle.

From above, it appears that the blowby flowrate seen in Figure 10 is very low and reaches its maximum of 0.28 g.s−<sup>1</sup> at 320◦ CA and then stabilises at around 0.18 g.s−<sup>1</sup> during the remainder of the compression stroke and the beginning of the piston expansion. During the expulsion of the exhaust gases, the blowby flowrate is close to 0. Then, due to the low pressure inside the cylinder during the intake stroke, the flowrate is negative and stabilises at around −0.15 g.s−1. During this period, gases from the oil pan are redirected to the cylinder.

**Figure 10.** Evolution of blowby flowrate at idle.

In the end, the amount of blowby gas that passes through the piston to the crankcase is zero since this accumulation is negative and amounts to −3.4 mg.cycle−1. This means that the phenomenon is reversed and that the gases from the oil pan could be directed towards the cylinder.

The oil sweeping effect is non-existent and the direction of the blowby flowrate tends to retain and push the oil stored in Lands 1 and 2 back into the cylinder. In addition, the compression and sealing rings oscillate two to three times per cycle within their respective grooves, and this creates a phenomenon of pumping and expulsion of the oil from the scraper ring to the crown of the piston via the sealing and compression rings. This phenomenon, described in particular by Thirouard [48] and Yilmaz, et al., [50] is significant at idle or low load.

Previous simulation results have shown that idling is conducive to an accumulation of oil at the piston crown. During rapid acceleration, the sudden increase in pressure and temperature in the cylinder will cause some of this oil to burn off, while some will be returned to the crankcase via the blowby phenomenon. The duration of the idling time between accelerations could have an impact on the amount of oil stored in the piston crown and consequently on oil consumption and particulate emissions. It is therefore important to study the impact of the duration of the idle time between two accelerations in order to check whether this accumulation phenomenon revealed by our simulation model is real. This is the subject of the following section.
