Xin Q. Diesel Engine System Design

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780 Diesel engine system design

observed at low engine speeds/loads. Diesel knocking can be reduced by

combustion calibration and/or by reducing the structural transfer function

for the combustion noise.

Advanced diesel engines use multiple combustion modes over the operating

range in order to meet stringent emissions standards. In addition, diesel

aftertreatment systems may require step (sudden) changes in parameters such

as air–fuel ratio, EGR rate, and fuel injection timing. Sudden changes in the

combustion modes or the operating parameters may cause sudden changes in

the noise level and the sound quality perceived by the customer. This poses

new NVH challenges. A smooth gradual transition between different operating

modes must be achieved by engine calibration and control algorithms in

order to eliminate the NVH issues to make the transitions undetectable by

the driver.

Combustion noise can be reduced by reducing the cylinder pressure level

(in dB) or increasing the structural attenuation (especially in the medium- and

high-frequency range). Note that it is the shape, rather than the maximum

level, of the cylinder pressure trace that affects the frequency spectrum of

the pressure. Basically, when the ignition delay is reduced, the rise rate of

the cylinder pressure and the high-frequency contents of the pressure are

reduced, thus the combustion noise can be reduced. In general, the following

factors can reduce the combustion noise: the combustion chamber design,

which can reduce the ignition delay and the rise rate of the cylinder pressure;

indirect injection combustion chamber; higher compression pressure and

temperature at the TDC; higher compression ratio; increased intake manifold

boost pressure; higher cylinder wall temperature; retarded fuel injection

timing; higher EGR rate; lower engine speed; lower load or lower fueling

rate; increasing structural attenuation of the cylinder bore (e.g., increasing

the structural stiffness by using larger stroke-to-bore ratio or more cylinders,

using a stiffer liner and block); and suppressing the vibration of the top deck

by design changes and raising the powertrain mode frequency (e.g., adding

ribs in the cylinder head top deck as reinforcement).

The timing and quantity of pilot fuel injection and the timing and pressure

of main injection affect the combustion noise. Modern fuel injection systems

offer exibility for NVH tuning, but this often has to be compromised

by emissions and performance requirements during engine combustion

development and calibration. If less opportunity could be offered from the

combustion calibration for noise reduction, the transfer characteristics of the

engine structure and the noise transfer paths in the vehicle would become

more important for the combustion noise control in both sound pressure

level and sound quality. This requires a system approach to optimize.

Reviews on engine combustion noise are provided by Challen (1975),

Anderton (1979), and Wolschendorf et al. (1991). The methodology of

separating the combustion noise and the mechanical noise was proposed

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781Noise, vibration, and harshness (NVH) in engine system design

by Brandl et al. (2007). Separation of the combustion noise and the piston

slap noise was researched by Badaoui et al. (2005). The transfer function

of engine structural attenuation was researched experimentally by Shu et

al. (2005a). Combustion noise optimization by structural attenuation and

an application of using the combustion noise meter was presented by Wang

et al. (2007). Torregrosa et al. (2007) developed an important innovative

approach for combustion noise assessment by using cylinder pressure

decomposition, and they showed that the result was more accurate than the

classical ‘block attenuation curve’ approach and the method may be used as a

promising alternative to calculate the combustion noise. Regarding the effect

of fuels, the impact of diesel cetane number on engine noise was measured

by Machado and De Melo (2005). The effects of fuel system design, fuel

injection strategies and emissions calibration optimization on engine noise

were studied extensively by Russell et al. (1990), Kohketsu et al. (1994),

Tabuchi et al. (1995), Badami et al. (2002), Mallamo et al. (2002), Roy and

Tsunemoto (2002), and Mendez and Thirouard (2008). Diesel cold start noise

was analyzed by Alt et al. (2005). Detailed in-cylinder CFD modeling of

combustion noise origins and sensitivities was conducted by Blunsdon et al.

(1995) and Luckhchoura et al. (2008). Diesel combustion noise variations

(cycle-to-cycle and cylinder-to-cylinder uctuations) were studied by Gazon

and Blaisot (2006). An overall combustion noise optimization process in

powertrain product development was described by Alt et al. (2001).

11.5 Piston slap noise and piston-assembly

dynamics

11.5.1 Piston slap noise

There are three possible types of piston noise, namely piston rattling noise

(i.e., the top land contacts the cylinder bore), piston pin ticking noise (i.e.,

pin-bearing impact), and piston slap noise (i.e., the skirt contacts the bore).

The rst two types of noise can be avoided or eliminated by proper design.

Piston slap cannot be eliminated because it is caused by the piston secondary

motions within the skirt-to-bore clearance that is originated from the crank–

slider mechanism by nature. Piston slap is usually the largest contributor

to mechanical noise, especially in the diesel engine. For a clearance-free

piston, the side thrust is a low-frequency forcing function which is related

to engine speed (see Chapter 10). With the clearance in the real engine, the

time history of the side force acting on the liner is changed by the additional

sharp impact forces when the piston moves within the clearance. These

impulsive impact forces are the high-frequency forcing functions driving

the cylinder liner and the engine block to vibrate and radiate an impulsive

type of noise. The piston slap noise is also transferred from the piston to the

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782 Diesel engine system design

connecting rod and the crankshaft and nally to the engine block. Moreover,

piston slap causes liner cavitation erosion in heavy-duty diesel engines with

the induced excessive liner vibration (Yonezawa and Kanda, 1985). Some

design parameters that can reduce piston skirt friction unfortunately affect

piston slap adversely.

In engine system design, a good understanding of the following topics is

required: (1) the characteristics of piston slap; (2) the modeling approaches;

and (3) an optimized overall planning for the piston assembly to balance

the trade-offs between fuel economy and noise. Although the piston skirt

mass, the skirt exibility, and the cylinder pressure load are quite different

between the diesel engine and the gasoline engine, they share many similar

characteristics. Some references mentioned in this section are from the

gasoline engine. Introduction to piston slap excitation, noise, and its related

design features are provided by Ross and Ungar (1965), Munro and Parker

(1975), Whitacre (1990), Slack and Lyon (1982), De Luca and Gerges (1996),

Chien (1995), Künzel et al. (2001), and Fabi et al. (2007).

Piston slap between the skirt and the cylinder bore is caused by the secondary

motions (transverse or lateral, and tilting) driven by the alternating piston

side thrust within the skirt-to-bore clearance. Not only moving transversely,

the piston also tilts around the piston pin and this usually results in the upper

(or lower) portion of the skirt slapping against the bore. There are several

piston slap events within one engine cycle due to those side thrust reversals

(Fig. 11.2). The most signicant one is usually the slap right after the ring

TDC (0° crank angle). For this impact, the piston moves across the ring

Piston skirt

corner 3

Piston skirt

corner 4

Piston

in bore

Slap

Scraping

the bore

0 90 180 270 360 450 540 630 720

Crank angle (degree)

Clearance between piston skirt top corners and

cylinder bore at thrust side (micron)

200

150

100

11.2 Simulation of cold piston slap motion without effective

lubrication.

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783Noise, vibration, and harshness (NVH) in engine system design

TDC to transition from a sliding motion against the anti-thrust side of the

piston in the late compression stroke to a slapping event on the thrust side

just after the TDC. The gas load acting on the piston may create a moment

around the piston pin to turn the piston, thus affecting the slap noise. Efforts

should be directed to reduce the most severe piston slap near the ring TDC

for piston slap noise control.

Piston slap noise is affected by all the factors involved in this mechanism,

i.e.,

∑ Piston side thrust force: lower reciprocating mass, lower engine speed,

lower cylinder pressure, and larger ratio of the connecting rod length

to the crank radius can reduce the side thrust, thus reducing the piston

slap noise (e.g., Oetting et al., 1984).

∑ The moment about the piston pin: lower moment of inertia, proper piston

pin offset, the crankshaft offset, the cylinder pressure acting point, the

moments from the cylinder pressure force, the gravity force, the lubricant

normal forces, the piston ring lateral friction force, and the piston pin

friction force may reduce or alter the moment about the piston pin to

reduce the piston slap noise. The friction force between the ring and its

groove bottom is in the boundary lubrication regime (with a coefcient

of friction equal to 0.1–0.2) and has a large inuence on piston slap

timing. When the ring oats, the resistance to piston slap from the ring

disappears and this usually aggravates piston slap. Munro and Parker

(1975) reported that a ring lateral friction force with a coefcient of

ring–groove friction equal to 0.1 can halve the impact velocity and

reduce the kinetic energy by a factor of 4.

∑ The allowable travel distance of the piston before hitting the wall: smaller

skirt-to-bore clearance can reduce the piston slap noise. For example,

lower liner temperature decreases the clearance by liner contraction.

Higher piston temperature may reduce the clearance. Using smaller

skirt-to-bore clearance can reduce the severity of piston slap but at the

expense of increased viscous shear friction. There is a trade-off between

low engine noise and high mechanical efciency when the piston clearance

is designed.

∑ The damping force to resist the piston secondary motions: adequate oil

supply on the skirt can reduce piston slap signicantly. Lower tension

in the piston rings (especially the oil-control ring) can increase the oil

lm thickness on the cylinder liner. It would allow a larger damping

effect of the oil lm to cushion the piston slap thus reducing the impact

velocity and the noise. The lubricating oil lm thickness and lubricant

force, which are also affected by the skirt length, the lubricant viscosity,

and the surface waviness or roughness, may reduce the piston impact

velocity and thus the slap noise. Ryan et al. (1994) showed that there

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784 Diesel engine system design

appeared to be an optimum oil viscosity to minimize the piston slap noise;

either a higher or lower viscosity increased the slap intensity. Using a

longer skirt length as a better guiding and damping surface may reduce

the piston slap noise. Increasing the contact area of the piston slap by

bore or skirt design changes (e.g., the piston ovality) to enhance oil lm

damping may reduce the piston slap noise.

∑ Piston impact velocity: all the above factors eventually affect the piston

impact lateral velocity, which partly characterizes the severity of the

slap.

∑ Piston mass: the piston mass contributes to the impact impulse or kinetic

energy. Larger piston mass and higher impact velocity make the piston

slap noise louder.

∑ The contact area during the slap: the contact area affects the transient

elastic collision process and the impact force. If the piston slap occurs

over a larger contact area, the impact energy can be better absorbed to

reduce the slap noise. Both vertical shape (skirt prole) and circumferential

shape (ovality) affect the contact area and hence the piston slap noise.

∑ The stiffness and the damping of the parts in contact: the stiffness and

the damping affect the impact force during the elastic impact process or

the coefcient of restitution. If the softer portion of the piston skirt (e.g.,

the lower portion of the skirt) slaps the bore, the noise will be lower

due to larger deformation. The elastic distribution of the skirt stiffness

should be uniform. It is important to increase the top land clearance

to avoid the contact between the very stiff top land and the bore. The

top land is essentially a solid disc of metal of high stiffness. Its contact

with the cylinder wall produces a sharp rattling noise. For the reasons

of noise and scufng, the top land should not touch the cylinder wall.

∑ The structural sound attenuation characteristics of the cylinder liner/

block.

Among the design factors, offsetting the piston pin relative to the lateral

location of the center of gravity of the piston is the most commonly used

technique to control the piston slap noise. As explained in the piston slap

mechanism above, it is the moment about the piston pin that controls the

piston tilting. The moment is affected by both the lateral pin offset and

the vertical pin position relative to the center of gravity of the piston. The

cylinder gas force largely inuences the titling moment. The lateral forces

(e.g., the lubricant force) play an equally important role to control the titling

moment. Therefore, the effectiveness of the lateral pin offset is dependent on

the vertical pin position. With the pin offset toward the thrust side, the gas

force will rotate the piston about the piston pin toward the anti-thrust side.

This rotation assures the bottom of the skirt reverses to contact the thrust

side before the top of the skirt crosses over, thus reducing the force that the

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785Noise, vibration, and harshness (NVH) in engine system design

top-side reversal would otherwise generate. The bottom portion of the skirt

is usually less rigid than the top portion, thus the piston slap can become

less noisy. On the other hand, offsetting the pin to the anti-thrust side causes

large slap noise because the reverse moment on the piston causes the upper

stiff portion of the skirt to contact the bore near the ring TDC. However,

offsetting to the anti-thrust side may yield a small (often marginal) decrease

in skirt friction. It should be noted that a large piston pin offset may cause

excessive piston tilting around the TDC and cause increased blow-by, oil

consumption, and friction. Sometimes there are trade-offs between the piston

slap noise and the piston tilting. An optimized skirt prole design may relieve

this trade-off by altering the lubricant moment acting around the piston

pin. As observed from the above-mentioned factors, piston slap control is a

complex task, but there are many opportunities to optimize it.

The importance of piston slap noise depends on engine applications. For

example, piston slap predominates in marine diesels which have relatively

large piston clearances, while it is less prominent in small gasoline engines.

The piston slap noise is especially prominent when the engine is cold and

the piston clearance is large without effective lubrication (e.g., at cold start).

The noise increases with engine speed and peak cylinder pressure. The

piston slap noise is most apparent at cold start and idle conditions, as well

as at low-speed high-load where other noises are relatively less apparent.

Künzel et al. (2001) discovered that the piston slap noise was most prominent

(audible) at low engine speeds (e.g., 1000–2000 rpm) from low loads to high

loads for passenger car diesel engines. Another important scenario is that

the piston slap noise occurs prominently after a cold start, when the piston-

to-bore clearance is at a maximum but the metal is cold without effective

lubrication. For example, Richmond and Parker (1987) found that at mid-

speed and low-load (e.g., 1600 rpm, one-third load, accelerating to 30 mph

after cold start) the piston slap noise may become the most intrusive. The

primary design measure to minimize the piston slap noise is to optimize the

piston secondary motions at all operating conditions so that with a change in

the skirt–bore contact pattern only a minimum amount of impact energy is

transmitted into the engine structure. The two most commonly used techniques

to control the piston slap noise are reducing the skirt-to-bore clearance and

offsetting the piston pin. Piston skirt prole also plays an important role in

noise control.

Cylinder liner or block vibration (or acceleration) has been proven to

be a good indicator of piston slap noise. It is found that the liner vibration

correlates with the piston impact kinetic energy very well. Kamiya et al.

(2007) used small thin-lm pressure sensors to directly measure the oil lm

pressure at the piston slap locations to try to understand the excitation force

at the slap location. They discovered that a clear correlation exists between

the oil lm pressure near the top of the skirt (located at the anti-thrust side)

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786 Diesel engine system design

and the cylinder liner acceleration measured near the liner top. This veries

that when the piston slap occurs, a large squeeze lm reaction happens in

hydrodynamic lubrication to produce the oil pressure. It indicates that the

piston slap velocity on the lubricated surface can also be used as an indicator

for the slap noise.

Piston slap noise and liner/block vibration measurements have been

conducted extensively over the past 30 years (DeJong and Parsons, 1982;

Furuhama and Hirukawa, 1983; Kaiser et al., 1988; Richmond and Parker,

1987; Vora and Ghosh, 1991; Kamp and Spermann, 1995; Ryan et al.,

1994; Nakada et al., 1997; Teraguchi et al., 2001). The measurement results

facilitate good understanding of the parametric dependency of piston slap

and provide support for analytical model development.

11.5.2 Piston-assembly dynamics modeling for piston slap

Piston slap is a very complex dynamic process involving the following key

topics:

∑ multi-body dynamics of the entire piston assembly and the crankshaft

∑ multi-phase dynamics including the transitions between non-scraping

and scraping motions

∑ rigid-body impact and rebound dynamics

∑ boundary lubrication friction for surface asperity load carrying

∑ lubrication and squeeze lm effect

∑ elastic impact and rebound

∑ deformation

∑ cylinder liner vibration response.

Numerical modeling can provide valuable tools to reveal the physical

mechanisms that are difcult to measure, and to predict the effects of design

changes. The ultimate goal of piston slap modeling is to predict the slap

noise. However, this can be extremely complex and time-consuming in

computation. Even in the experimental work, direct measurement of piston

slap noise is difcult because it is not easy to separate out the piston slap

noise from other noises. Instead, alternative and easier available parameters

such as the liner vibration (acceleration) signal are used to judge the severity

of the piston slap. With proper calibration the liner acceleration data can

be converted through a transfer function to an estimated piston slap noise.

A similar approach can be taken in simulation by evaluating the kinetic

energy loss or impact impulse. Piston slap can be accessed by using the

piston impact and rebound velocities as key parameters. Experimental work

showed that the piston slap noise is almost linearly proportional to the total

kinetic energy due to piston impact excitation. A time history and frequency

spectrum of the impact force are also important.

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Piston slap is essentially driven by the side thrust. After a series of impact

and rebound events within a short period of time, when the impact velocity

is damped to negligibly small, the piston enters a scraping phase of motion.

During the scraping phase the skirt-to-bore contact portion is supported by the

surface asperity contact under the boundary friction, and the rest of the skirt

may undergo hydrodynamic lubrication. The lubrication dynamics formulation

of the piston assembly is introduced in Chapter 10. This section will address

the key issues in the detailed level-2 and level-3 NVH modeling of piston

slap and the link between piston slap and piston friction. Both the level-2 and

level-3 models have instantaneous crank-angle resolution. When the model

is complex enough to include structural attenuation or noise calculation with

FEA and the BEM, the model is at level-3. The simplied level-1 modeling

of piston slap will be introduced in the system-level summary in Section

11.12.

The research on piston slap modeling started in the 1970s. The numerical

modeling initially focused on using the multi-body multi-phase dynamics

to calculate the piston slap velocity for a rigid-body cylindrical piston, and

to calculate the piston rebound velocity based on an assumed coefcient

of restitution. The piston-to-bore friction was calculated by using a simple

model of Coulomb friction without the hydrodynamic lubrication model.

The model predicts the trends reasonably well compared to the experimental

measurement data of the piston secondary motions, and is able to reveal

many fundamental mechanisms of the piston slap. The model is especially

suitable for the cold engine conditions where the skirt lubrication is not

effectively established. Typical work in this category includes Wilson and

Fawcett (1974), Haddad and Howard (1980), Haddad and Tjan (1995), and

Haddad (1995). Detailed derivations of the multi-body dynamics equations of

piston assembly including impact dynamics are given by Wilson and Fawcett

(1974), Haddad (1995), and Xin (1999). When the piston dynamics model

is simplied from multi-body to single-body (i.e., only the piston skirt),

errors are introduced in the calculations of the piston side thrust and the

tilting moment. Some researchers use this method for simplicity. However,

the piston slap timing and the impact velocity cannot be accurately modeled

with the simplied method. Piston slap and scraping motion measurements

were conducted by Rohrle (1975) and Sander et al. (1979).

The energy transferred from the piston slap to excite the block vibration

and radiate the noise comes from the piston impact against the bore. The

impact is not a one-time event. Instead, it is a consecutive series of impact

events with gradually decreasing impact velocity due to the dissipation of

impact energy into the structural material. The impact and rebound events are

modeled by using the concept of coefcient of restitution, which reects the

attenuation in the rebound velocity relative to the impact velocity. When the

impact velocity becomes negligibly small and there is still a normal force to

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788 Diesel engine system design

push the skirt against the bore, the piston motion can be regarded as a scraping

motion sliding on the bore with the boundary lubrication. It should be noted

that the rebound motion is important in modeling because it is possible that

after the impact events the piston will rebound to leave the bore without

recurring impacts and continue with another non-scraping motion without

encountering the boundary lubrication friction. In the scraping phase of the

motion, the load carried by asperity contacts is calculated based on the force

and the moment balances of the piston. The friction can be calculated by

either assuming a friction coefcient or using more sophisticated asperity

contact models.

The piston impact is usually assumed to occur over an innitesimally

small time interval with negligible changes in the mechanism’s kinematic

positions between right before and just after the impact. When the impact

occurs, the instantaneously released kinetic energy is calculated using the

theory of conservation of momentum from both the lateral and rotational

piston velocities. The kinetic energy dissipated at the impact depends on the

coefcient of restitution used. The topics of the coefcient of restitution,

the impact and the rebound on lubricated and non-lubricated surfaces, and

the impact with friction are important subjects in the areas of mechanical

impact dynamics and tribology research. The impact dynamics is the bridge

between the non-scraping phase of motion and the scraping phase of motion.

The assumed coefcient of restitution provides the initial condition of the

piston lateral velocity after the rebound in order to carry on the numerical

solution of the ordinary differential equations of the dynamic motion. Note

that the focus of piston slap analysis is on the series of impact events, while

the focus of the boundary lubrication friction analysis is on the scraping

motion after the impact events.

The existence of effective lubrication on the piston skirt depends on

particular engine operating conditions. Furuhama and Hirukawa (1983)

found that in their experiment under the low-temperature cold idle condition,

almost all the spaces in the large piston-to-bore clearance were lled with

gas, and severe oil starvation happened with little squeeze lm action from

the oil. Under this circumstance, a model without lubrication is justied.

Although many useful simulation results of the piston impact and the

scraping motions can be obtained by ignoring the skirt lubrication model,

experimental results showed that adequate oil supply and oil attainment in the

skirt area can reduce the piston slap noise. A lubrication model may affect

the predicted piston lateral, tilting, and scraping motions and the impact

velocity in the following aspects (either more accurate or less):

∑ A skirt lubrication model may signicantly change the moment balance

around the piston pin due to the oil lm distribution on the skirt. This

is particularly true when a non-cylindrical barrel-shape piston skirt is

modeled where strong lubricant forces are developed simultaneously at

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789Noise, vibration, and harshness (NVH) in engine system design

both the thrust and anti-thrust sides and the lubricant moments play a

key role.

∑ A skirt lubrication model provides an oil lm damping effect mainly

through the dynamic squeeze lm effect to cushion/reduce the piston

impact velocity.

∑ The existence of an oil lm affects the equivalent stiffness and damping of

the contact area during the piston slap process, thus affecting the kinetic

energy dissipation, the coefcient of restitution (the piston rebound),

and the piston slap noise.

∑ A skirt lubrication model affects the force and moment balances of

the skirt in the scraping motion so that the load carried by the asperity

contacts is affected in the calculation. When the piston scrapes on the

bore, the hydrodynamic lubricating oil lm pressure may support a large

portion of the load, and the rest of the load has to be carried by the

surface asperity contacts. This may affect the accuracy of the boundary

lubrication friction during the scraping.

The piston skirt lubrication model can be as simple as a damper or as

complex as a full Reynolds equation. Theoretically, according to the Reynolds

equation of the squeeze lm lubrication, the damping factor of a squeeze

lm damper is inversely proportional to the oil lm thickness to the power

of 3. Haddad and Tjan (1995) provided a simplied oil lm thickness model

and a semi-empirical equation of the oil pressure as a function of skirt

clearance. Gerges et al. (2005) used a simplied dynamic model consisting

of a spring (representing the air bubbles in the oil lm generated during the

impact event) and a damping dashpot (representing the oil lm) connected

in series to simulate the behavior of the oil lm mixed with the air bubbles,

and they found that the oil aeration could greatly reduce the load carrying

capacity of the oil lm. Kobayashi et al. (2007) used an elastohydrodynamic

lubrication model including the aeration effect of the air bubbles inside the

oil lm. Nakada et al. (1997) used a stiffness element and a damper to model

the skirt, a viscous damper to model the oil lm, and a stiffness element

and a damper to model the liner. These elements were connected in series.

In order to speed up the calculation, they simplied the Reynolds equation

from two-dimensional to one-dimensional by dropping the tilting-motion

term and only considered the lateral motion. They obtained a closed-form

analytical solution of the oil pressure and the lubricant force. Even such a

highly simplied lubrication model could predict piston slap motion quite

accurately compared to their measurement data. Wong et al. (1994) introduced

the two-dimensional Reynolds equation to simulate the lubricated piston slap

with elastohydrodynamic and deformation effects for a barrel-shape skirt.

Deformation modeling for piston slap covers two aspects: the component

deformation before the impact, and the elastic deformation during the short

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