Xin Q. Diesel Engine System Design

Simulation of cylinder deactivation at 1000 rpm

Simulation of cylinder deactivation at 1600 rpm

Simulation of cylinder deactivation at 1400 rpm

Simulation of cylinder deactivation at 2000 rpm

S1 S3 S7 S8

0 5 10 15

BMEP (bar)

0 5 10 15 20 25 30

BMEP (bar)

0 5 10 15 20 25 30

BMEP (bar)

0 5 10 15 20 25 30

BMEP (bar)

Overall engine A/F ratio

115

105

95

85

75

65

55

45

35

25

15

95

85

75

65

55

45

35

25

15

95

85

75

65

55

45

35

25

15

105

95

85

75

65

55

45

35

25

15

9.28 Air–fuel ratio of cylinder deactivation at different engine speeds and loads.

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Simulation of cylinder deactivation at 2400 rpm

Simulation of cylinder deactivation at 3000 rpm

Simulation of cylinder deactivation at 2600 rpm

S1 S3 S7 S8 S1 S3 S7 S8

S1 S3 S7 S8

0 5 10 15 20 25

BMEP (bar)

0 5 10 15 20 25

BMEP (bar)

0 5 10 15 20 25

BMEP (bar)

Overall engine A/F ratioOverall engine A/F ratio

Overall engine A/F ratio

85

75

65

55

45

35

25

15

65

60

55

50

45

40

35

30

25

20

75

65

55

45

35

25

15

9.28 Continued

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9.29 Cylinder deactivation performance at 2600 rpm and limiting constraints at 3000 rpm.

Simulation of cylinder deactivation at part load and 2600 rpm

S1 S3 S7 S8

0 5 10 15 20 25

BMEP (bar)

0 5 10 15 20 25

BMEP (bar)

Exhaust manifold gas

temperature (°F)

LP turbine outlet temperature

(°F)

1400

1300

1200

1100

1000

900

800

700

600

500

400

1100

1000

900

800

700

600

500

400

300

Simulation of cylinder deactivation at part load and 2600 rpm

S1 S3 S7 S8

0 5 10 15 20 25

BMEP (bar)

0 5 10 15 20 25

BMEP (bar)

Peak cylinder gas temperature of

cylinder 1 (°F)

Engine delta P (back minus

boost, mbar)

3500

3000

2500

2000

1500

1000

800

600

400

200

0

–200

–400

Diesel-Xin-09.indd 632 5/5/11 11:59:17 AM

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9.29 Continued

Simulation of cylinder deactivation at part load and 3000 rpm

S1 S3 S7 S8

0 5 10 15 20 25

BMEP (bar)

0 5 10 15 20 25

BMEP (bar)

Peak cylinder gas pressure (psia,

maximum of all cylinders)

HP compressor of outlet

temperature (°F)

4000

3500

3000

2500

2000

1500

1000

500

0

450

400

350

300

250

200

150

100

50

0

Simulation of cylinder deactivation at part load and 3000 rpm

S1 S3 S7 S8

0 5 10 15 20 25

BMEP (bar)

0 5 10 15 20 25

BMEP (bar)

Intake manifold boost pressure

(bar, absolute)

Exhaust manifold back pressure

(bar, absolute)

4

3.5

3

2.5

2

1.5

1

0.5

0

4

3.5

3

2.5

2

1.5

1

0.5

0

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Table 9.2 Simulation cases of cylinder deactivation

Scenario

number

Mode Turbine area HP-stage turbine

wastegate

opening

Valvetrain operation Strategy of engine

valve shut-off timing for

deactivated cylinders

S1 Part load Baseline turbine Fully open No cylinder deactivation

S2 High load Baseline turbine Open as needed No cylinder deactivation

S3 Part load Baseline turbine Fully closed No cylinder deactivation

S4 High load Baseline turbine Closed No cylinder deactivation

S5 Part load Baseline turbine Partially closed No cylinder deactivation

S6 Part load Baseline turbine Fully closed Cylinder deactivation by shutting off

fuel only

S7 Part load Baseline turbine Fully closed Cylinder deactivation by shutting off

fuel, intake valve and exhaust valve

(shut-off strategy 1)

Shut-off strategy 1: Shut

off intake valve before each

cylinder’s IVO

S8 Part load Baseline turbine Fully closed Cylinder deactivation by shutting off

fuel, intake valve and exhaust valve

(shut-off strategy 2)

Shut-off strategy 2: Shut

off intake valve after each

cylinder’s IVC

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S9 Part load Baseline turbine Fully closed Cylinder deactivation by shutting off

fuel and only intake valve (shut-off

strategy 1)

Shut-off strategy 1: Shut

off intake valve before each

cylinder’s IVO

S10 Part load Baseline turbine Fully closed Cylinder deactivation by shutting off

fuel and only intake valve (shut-off

strategy 2)

Shut-off strategy 2: Shut

off intake valve after each

cylinder’s IVC

S11 Part load 10% smaller turbine area

in both HP and LP stages

Fully open No cylinder deactivation

S12 High load 10% smaller turbine area

in both HP and LP stages

Open as needed No cylinder deactivation

S13 Part load 10% smaller turbine area

in both HP and LP stages

Fully closed Cylinder deactivation by shutting off

fuel, intake valve and exhaust valve

(shut-off strategy 1)

Shut-off strategy 1: Shut

off intake valve before each

cylinder’s IVO

S14 Part load 10% smaller turbine area

in both HP and LP stages

Fully closed Cylinder deactivation by shutting off

fuel, intake valve and exhaust valve

(shut-off strategy 2)

Shut-off strategy 2: Shut

off intake valve after each

cylinder’s IVC

Note: S9 and S10 have the same steady-state performance.

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636 Diesel engine system design

Figure 9.23 illustrates the speed–load modes used in the GT-POWER

simulation. Note that usually the load range in cylinder deactivation research

is 0–5 bar BMEP, but this simulation extends the load range up to 10–13 bar

BMEP at high speeds greater than 2500 rpm. Figure 9.24 shows the cylinder

pressure traces of different valve switching strategies in deactivation.

Figure 9.25 compares the engine performance given by different deactivation

strategies at 2400 rpm. It can be seen that, compared to the V8 operation

(S1), shutting off only fueling (S6) and shutting off only fueling and intake

valve (S9) do not provide any benet in BSFC reduction. In fact, the BSFC

of S6 and S9 is even higher than that of S1. In S7 and S8, the exhaust valve

is shut off after the intake valve is shut off. Some BSFC benet occurs at

very low load (0–2 bar BMEP) when both the intake and exhaust valves are

closed and a certain amount of air is trapped in the deactivated cylinders

(S8). The largest benet of BSFC reduction occurs with scenario S7 in a wide

range of BMEP (0–10 bar) when very little air is trapped in the deactivated

cylinders.

Figure 9.26 illustrates the effect of 10% smaller turbine area on the

air–fuel ratio. The smaller turbine can increase the air–fuel ratio in the

cylinder deactivation operation in order to extend its operating range to higher

BMEP level. However, the turbine area reduction in a x-geometry turbine

that is sized for achieving higher air–fuel ratio during cylinder deactivation

results in a pumping loss penalty in the non-deactivation operation due to the

excessively high air–fuel ratio. VGT can help alleviate this trade-off between

the cylinder deactivation operation at low loads and the non-deactivation

operation at full load.

Figure 9.27 summarizes the BSFC changes at different speeds and loads

for different turbine wastegate openings (S1 and S3) and deactivation valve

switching strategies (S7 and S8). The lower and upper bounds of the BSFC

reduction benet range bounded by S7 and S8 indicate the two extreme

limits of different valve switching strategies. It is observed that the BSFC

benet is a strong function of both engine load and speed for the diesel

engine. Very large benet of BSFC reduction can be obtained by cylinder

deactivation. Note that the BSFC benet computed here is with respect to a

baseline engine that already has very low BSFC with the turbine wastegate

set fully open at part load by using electronic controls (i.e., S1 rather than

S3). At high engine speeds, the BSFC benet can extend to a very high

BMEP level with acceptable air–fuel ratio (e.g., 10 bar BMEP at 2600

rpm) because the turbocharger can deliver more air than at lower speeds to

maintain a minimum required air–fuel ratio.

Figure 9.28 summarizes the air–fuel ratios corresponding to the cases in

Fig. 9.27. Note that due to reduced air–fuel ratio the cylinder deactivation

operation generally will produce more soot than the non-deactivation operation,

especially when the air–fuel ratio is close to the smoke limit. Other measures

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637Advanced diesel valvetrain system design

(e.g., retarding EVO and advancing IVC with VVA in the ring cylinders)

can be used to reduce the soot at relatively high BMEP level and at low to

medium speeds.

Figure 9.29 illustrates other key performance parameters at 2600 rpm

and the limiting design constraints at 3000 rpm in cylinder deactivation.

It is important to note the effect of cylinder deactivation on the following

two parameters: (1) engine delta P, which is the driving force for EGR ow

in EGR engines and can also be affected by the turbine area selection; and

(2) the low-pressure (LP) stage turbine outlet gas temperature, which can

help the operation and regeneration of aftertreatment devices (e.g., DPF).

Finally, note that the BMEP limit of cylinder deactivation can be restricted

by the peak cylinder pressure in the ring cylinders and the compressor

outlet air temperature at high engine speeds (e.g., 3000 rpm), in addition to

the minimum required air–fuel ratio.

The performance of diesel engine cylinder deactivation is summarized as

follows.

1. The optimum cylinder deactivation method is to shut off fueling and all

the intake and exhaust valves in the deactivated cylinders.

2. The deactivation valve switching timing and gas mass retained in

the deactivated cylinders greatly affect the BSFC benefit of the

deactivation.

3. The optimum number of deactivated cylinders largely depends on the

balance between the air–fuel ratio for combustion/emissions and the

pumping loss for the lowest BSFC. The optimum number is also affected,

although secondarily, by the design features of cylinder heat transfer loss

and piston ring friction of the engine. Deactivating half of the cylinders

may not always be the optimum for achieving the lowest BSFC.

4. Cylinder deactivation provides very large BSFC improvement under

low-load conditions. The benet of BSFC reduction strongly depends

on both BMEP and engine speed for turbocharged diesel engines.

5. BSFC improvement is difcult to achieve with cylinder deactivation at

medium to high loads due to the limit of air–fuel ratio at low to medium

speeds and due to other design constraints at high speeds such as peak

cylinder pressure and compressor outlet air temperature.

6. The extent of BSFC improvement by cylinder deactivation is dependent

upon the following factors: the number of cylinders deactivated, valve

shut-off strategy, valve switching timing, turbine area (variable or xed,

and the area size), turbine wastegate control capability and exibility at

low loads, cylinder heat transfer, and the pressure-dependent portion of

the piston ring friction.

7. As the cylinder heat transfer and piston ring friction loss increase in the

baseline engine design, more benet of BSFC reduction can be achieved

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638 Diesel engine system design

by using cylinder deactivation because a larger portion of the energy

losses can be avoided in the deactivated cylinders.

9.9.4 Design challenges of cylinder deactivation

Engine vibration

NVH and drivability are important concerns for cylinder deactivation. Cylinder

deactivation reduces the frequency and increases the amplitude of engine

vibration at the crankshaft, which may not be acceptable at all engine speeds

in terms of NVH and comfort. The success of cylinder deactivation and

the benet in fuel economy improvement can only be achieved if the NVH

issues are resolved in mass production vehicles. As pointed out by Leone

and Pozar (2001), the benet of BSFC reduction in cylinder deactivation is

constrained or limited by NVH. For example, NVH and drivability concerns

are particularly severe in the rst and second gears of the transmission. If

cylinder deactivation is constrained only to the gears higher than the rst

and second, 2–4% of BSFC reduction benet will not be gained. If cylinder

deactivation is not used during warm-up, another 1–2% of BSFC reduction

benet will not be gained (Leone and Pozar, 2001). The major challenge of

cylinder deactivation is to maximize the fuel economy benet while meeting

the requirements of NVH and drivability in both steady-state operation and

transient transitioning of turning the cylinders on and off (i.e., smooth ramp-

out and ramp-in behavior).

The gas ow through the ring cylinders and the gas pressure pulses in

the manifolds affect the intake and exhaust noises. The cylinder pressure

in the deactivated cylinders plays an important role in engine vibration. As

long as the cylinder pressure is not uniform between the cylinders, engine

vibration and speed variability will increase. The cylinder pressure is affected

by the valve switching strategy in deactivation and is dependent on the air

or gas mass retained in the deactivated cylinders. The regular compression

without shutting off the valves in the non-fueling cylinders balances better

the engine vibration caused by the ring cylinders. With the extreme case of

vacuum compression with both the intake and exhaust valves shut off and

without air mass retained, the engine speed will exhibit higher variability

due to more violent vibration and torque pulsation at the crankshaft. The

higher dynamic torque and the lower frequency associated with the halved

basic engine order in cylinder deactivation cause the vibration amplitude

of the powertrain and drivetrain to increase (Lee and Rahbar, 2005). In

particular, cylinder deactivation used at low idle may create the problem

of excessive NVH due to the very low frequency and high amplitude of

engine vibration at low speed. Watanabe and Fukutani (1982) found that

the amplitude increase of engine vibration in diesel cylinder deactivation at

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639Advanced diesel valvetrain system design

low idle was much larger than that in gasoline cylinder deactivation. The

much higher peak cylinder pressure in the diesel engine was believed to be

the reason.

Potential design improvement in vibration control for cylinder deactivation

includes optimizing the engine mounts, using additional torsional dampers

or isolators in the driveline, increasing the inertia of the rotating engine

components, and increasing torque converter slip (Hatano et al., 1993;

Falkowski et al., 2004), as well as active vibration control measures such

as active tuned absorbers or active engine mounts (Lee and Rahbar, 2005).

However, all these measures have some negative effects. In general, there is no

easy solution for the vibration problem caused by cylinder deactivation. New

technologies are needed to solve the NVH and fuel injection compensation

problems encountered in cylinder deactivation.

Exhaust noise

The engine operating with different number of active cylinders in cylinder

deactivation exhibits different exhaust sound characteristics due to the change

in exhaust pressure waves. The design objective for cylinder deactivation is to

avoid perceivable subjective changes in the tailpipe acoustics. The traditional

active noise control device for this problem in the gasoline engine is usually

to use a ap valve located in the exhaust pipe. The more advanced passive

noise control design is to use passive resonators to attenuate the noise during

the cylinder deactivation operation. Bemman et al. (2005) successfully used

Helmholtz resonators to attenuate the low frequencies inherent to the V4

cylinder deactivation mode for a V8 gasoline engine. Their system is more

cost-effective than the traditional approach of using active valves in the

exhaust system to tune the exhaust noise.

9.9.5 Cylinder deactivation matched with other

technologies

Cylinder deactivation is essentially a technology of variable engine

displacement achieved by VVA to try to optimize fuel consumption at low

loads and obtain high power at high loads. A xed cylinder displacement

downsizing via turbocharging does not have the NVH issues brought by

cylinder deactivation, and may also produce lower parasitic losses. However,

the displacement reduction in xed cylinder displacement downsizing is

severely limited by the achievable power density at full load.

The competing engine technologies for cylinder deactivation in pumping

loss reduction or fuel economy improvement include VVA, VGT, HCCI/CAI,

continuously variable transmission, hybrid electric, and hybrid hydraulic.

Both HCCI/CAI and cylinder deactivation have the potential to decrease

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Xin Q. Diesel Engine System Design

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