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Advanced Engine Design Q&A

 

Everything you ever wanted to know about engine design and more!  One of our engineers recently completed a graduate course in advanced engine design in a university's automotive mechanical engineering program and a selection of various test questions and answers are presented here.  Frankly, this is pretty boring stuff but he felt he'd worked too hard answering these test questions not to have us offer it here as a sleep aid!  The course number and university aren't mentioned just in case this professor recycles old test questions - we can't afford any ill-will from this prof!  No guarantees on the correctness or validity of the answers of course - no debates on the finer points of single cylinder engine balance please!  This info is just FYI, actually, we're really not sure why we're posting it here.  Peruse at your own risk - or if you need help sleeping.

Question #1
 a. Discuss engine vibration and balance from the viewpoint of

(i) configuration:

Answer:

There is nearly an infinite number of possible 4-cycle SI engine configurations, as almost any geometric shape can and has been used.  Essentially engines can be designated as being in-line, opposed, V-type, radial, or W-type.  However, almost every current production automobile uses one of the following configurations: in-line 4 cyl, in-line 6 cyl, V8 90 deg and V6 60 deg so discussion will be limited to these.  It might be noted however that the V6 90 deg is also widely used in production including the GM 3.8L (Buick) and 4.3L (Chevy), both currently use countershafts for balance although earlier versions did not, and there is an inline 5 cyl upcoming (as well as an earlier one from Audi); unfortunately, very little data is available in textbooks for these. 

Primary and secondary inertia forces are discussed in detail in response to a latter question, as well as torsional forces.  In order to eliminate all moments, in-line and V engines can be designed around a symmetrical crankshaft with an even number of cylinders such that the cranks are arranged symmetrically about a plane perpendicular to the crankshaft axis.  Therefore, the unbalanced force at each crank is balanced by an equal and opposite force on another crank 180° from it so that moment arms around the plane of symmetry are equal.  All inline (and V12, V16) 4-stroke engines utilize a symmetrical crankshaft.

 [Taylor Vol 2 Ch 8 page 254-256]

In-line 4 cylinders are discussed in detail in response to part iii of this question regarding the use of a balance shaft, but basically they can be inherently balanced without the use of an additional countershaft to the extent that all primary vertical and horizontal inertia forces and moments are canceled out, as well as secondary horizontal forces and moments.  The secondary vertical shake of Fv = Q2cos2q does exist though, and is addressed by the use of 2 countershafts. 

Inline 6 cylinders 4 strokes with three crank positions are inherently extremely well balanced and vibration free and do not require any additional countershafts.  All primary and secondary inertia forces and moments can be completely eliminated.

 [Taylor Vol 2 Ch 8 page 260]

V6 engines with 60 degs between banks are also discussed in detail in response to Part B of this question.  Basically both primary and secondary inertia forces are balanced, but there is an inherent large primary couple that must be canceled with counterweights and a much smaller secondary couple that is simply tolerated.  The following moments exist, both vertically and horizontally: Primary moments Mv = ½ aÖ3 · Z sin q, Mh = 3/2 aÖ3 · Z cos q, and secondary moments Mv = ½ aÖ3 · lZ sin 2q, Mh = 3/2 aÖ3 · lZ cos 2q

V8 90 deg engines are unique in that they use an antisymmetrical crankshaft, where the crank of the last cylinder is 180° from cylinder #1, the next to last cylinder is 180° from cylinder #2, etc.  In engines with antisymmetrical crankshafts, the moment about their center due to secondary forces is zero, and if the primary moment can be offset by counterweights, all moments can be eliminated. With the 90° arrangement of the crank throws balancing the primary and secondary inertia forces in each bank, 90° V8 engines thus have excellent balance and freedom from vibration.

[Taylor Vol 2 Ch 8 page 257-259]

While detail discussion will not be presented on these configurations, it should be noted that V12 engines with 60° V angle for even firing, have excellent balance and freedom from vibration, while V2 and V4 engines all have balance problems.

[Ubong Winter 99 ME672 lecture Week 2 Part b]

A detailed listing of many possible engine configurations and their inertia balance can be found in Taylor, Vol 2, Chapter 8, Table 8-2 page 298-301.

a. Discuss engine vibration and balance from the viewpoint of

(ii) firing order:

Answer:

In a 4 stroke engine, each cylinder has a power stroke every 2 revs, thus the engine will rotate 720 degs between successive combustion events of a given cylinder, and the crankshaft rotation separating the firing of successive cylinders must be (720 / # of cyl) if evenly spaced firings are desired. 

Although it is desirable, it is not mandatory that engines must have even firing as there have been many production engines that were not.  An example would be the Buick 3.8L (231 cu in) 90 degree V6, which in the past was once an odd fire engine, although in the past 20 years it has since been produced as an even fire engine with split crank throws where cylinders share a common crank journal.

[Ubong Winter 99 ME672 lecture Week 2 Part b]

With respect to balance and vibration, torsional vibration is the chief consideration that firing order can affect.  Torsional crankshaft vibration is the rapid back and forth angular twisting along the crankshaft’s length.  Torsional vibration occurs when the natural torsional oscillations of the crankshaft (or a whole multiple of this natural frequency) happen to coincide with the power impulses of the power strokes that wind up the crankshaft – the speed at which this occurs is the critical speed.  In production engines, it is really the whole multiples of this natural frequency that synchronize with the power pulses to produce any torsional vibration because the natural frequency of the crankshaft vibration is always well above the designed maximum power pulse frequency (i.e. well above engine redline).  Operating an engine at these critical speeds for prolonged periods of time could result in fractured crankshafts. 

Generally, successive power impulses are applied to alternate ends of the crank, but in some instances successive firing of adjacent cylinders prevents relative unwinding, thereby suppressing certain critical torsional vibrations.  Besides a well chosen firing order, a further protection against torsional vibration is the use of a tuned rubber or viscous fluid torsional damper (also known as the harmonic balancer).

Besides being a factor in suppressing vibration, firing order is also very important in improving intake charge distribution and exhaust gas release, as tuning of the intake and exhaust manifold are dependent on the firing order.

[Heisler 2nd Ed.  Ch 11.4, Page 313 and Ch 12.3 page 335]  

a. Discuss engine vibration and balance from the viewpoint of

(iii) use of balance shafts (use a 4 cylinder engine for your analysis)

Answer:

Primary forces are inertia forces created by the acceleration (+ and -) of the piston assembly mass caused by the rotating crankpin’s projected motion along the line of stroke due to the reciprocating motion of the piston assembly.

Secondary forces are those inertia forces caused by the projected motion perpendicular to the line of stroke caused by the rotating motion of the connecting rod.  That is, the secondary force is due to the additional piston acceleration (both + and -) produced by the rotating crankpin increasing or decreasing the inclination of the connecting rod to the line of stroke.  During the first 90 deg of crank rotation, this secondary movement of the connecting rod is away from the line of stroke, thus adding to the piston movement while during the second 90 deg of crank rotation, this secondary movement of the connecting rod is toward the line of stroke, thus subtracting from the distance the piston moves.  Furthermore, secondary forces increase and decrease magnitude at twice the frequency of the primary force, but their maximum values are only about ¼ of the dominating primary force.

[Heisler 2nd Ed Ch 12.1, p328]
[Ubong Winter 99 ME672 lecture Week 4 Part a] 

Inertia Force F = Q1 Cos q + Q2 Cos 2q     q  = angle between crankpin and cylinder axis

Q1 Cos q represents the Primary inertia force (1st Harmonic), occurring at engine speed due to Cos q term
Q2 Cos 2q represents the Secondary inertia force (2nd Harmonic), at twice engine speed due to Cos 2q 

Given a 4 cylinder with the following crank configuration, the following moments can be defined:

X = distance between crank throws,
lever arm to cyl1 is +3X/2
lever arm to cyl2 is +X/2
lever arm to cyl3 is –X/2
lever arm to cyl4 is –3X/2 


 Sketching the crank in some general position and sum the forces as follows:


 As shown above, the sum of the primary forces is zero.  Completing the analysis,
 We can see that all primary forces and moments for this 4 cylinder engine sum to zero,

 
and secondary moments sum to zero as well.  However, a second order force unbalance remains at twice engine speed.

 [Ubong Winter 99 ME672 lecture Week 3 Part a] 


This diagram illustrates the Primary and Secondary forces again.  The full wide arrows represent the primary forces and the narrow half arrows represent the secondary forces.  Note that the primary forces are actually not exactly balanced but are very close, so in reality there still exists some vertical shake.  The primary couples are absorbed by the rigidity of the crankshaft material, while the four secondary forces definitely cause vehicle shake.

 
 This diagram demonstrates how the primary forces are canceled out, while the 4 secondary forces are additive at 90 deg intervals, contributing to horizontal shake.  On engines of less than 2 liters, this shake can be tolerated; however, on larger displacement engines this secondary imbalance must be corrected by using twin countershafts.


The weight and rotation direction of the countershaft is critical for proper operation.  These two countershafts must have weights Bl and Br equivalent in magnitude to the secondary inertia forces of all 4 cylinders.  Bl+Br = 4Fs, Fs = secondary force.  The countershafts revolve at twice crankshaft speed using a 2:1 chain and sprocket configuration and rotate counter to each other (Bl clockwise, Br counterclockwise) and are timed so that they counteract +Fs at 0 deg (TDC) and 180 deg ATDC, and –FS at mid-crank position (90 deg ATDC). 

 
 Secondary balance is accomplished by the weights facing in the opposite direction to these secondary forces when in the vertical plane. In the horizontal plane the weights will oppose each other, with one facing inward (135 deg ATDC) while the other faces outward (45 deg ATDC).

[Ubong Winter 99 ME672 lecture Week 4 Part a]

[Heisler 2nd Ed, 12.2 page 335]

(b) Given a choice to design a V-6 engine, what would be your preference of the angle 60° or 90° and why?

Answer:

60° would be the preferred choice. 

In a 4 stroke engine, each cylinder has a power stroke every 2 revs, thus the engine will rotate 720 degs between successive combustion events of a given cylinder, and the crankshaft rotation separating the firing of successive cylinders must be (720 / # of cyl) if evenly spaced firings are desired. 

720/6 = 120, so every 120 degrees the crankshaft must present a piston into firing position.  120 is an even multiple of 60, so a V6 engine with 60 degrees between banks naturally presents a piston into firing position.  A strong crankshaft is easily designed since cylinders can easily share a common crank throw and journal. 

In the case of a V6 with 90 degrees between banks, this is not the case.  In order for cylinders to share a common crank throw, the journals would have to be split and twisted to accommodate the 30 degree discrepancy in order to maintain evenly spaced firings.  This weakens the cranks somewhat as compared to a crank were the common journals do not need alteration.  The only alternative to splitting and twisting the crank journal in a 90 degree V6 is to forego the evenly spaced firings and produce an odd fire engine.  This was actually done in the case of an early 3.8L Buick V6 as previously mentioned.

The 60 deg V6 has six cranks evenly spaced at 60 deg intervals, with the pistons of each bank linked to every second crankpin so that the crankpin interval for each bank is 120 deg.  This crankshaft can be considered to be 2 inline three-cylinders merged into one, so in effect this engine is 2 in-line three-cylinders sharing a common crankshaft.  Both primary and secondary forces are balanced, but there is an inherent large primary couple that must be canceled with counterweights and a much smaller secondary couple that is simply tolerated. 

[Ubong Winter 99 ME672 lecture Week4 Part b]

[Heisler 2nd Ed, 12.1.4 page 332] 

However, there are some valid reasons why one might choose to build a 90° V6 rather than a 60°.  Packaging considerations for fitting the engine into a particular vehicle’s engine compartment may dictate your choice.  Given the same stroke, the 60° V6 would be more narrow than a 90° engine, but it would also be taller.  If there was a limitation on height in a particular application, this may eliminate the use of a 60° V6.  Additionally, the 90° V6 could use existing V8 tooling and thus may financially justify it’s use (as in the case of the 3.8L Buick V6).

(c) Discuss with diagrams the balancing of a single cylinder engine.
 

Answer:

 
 For a single crank-connecting-rod at a constant rotational speed, the following forces exist:

A force Fp acts along the cylinder axis on the piston assembly to accelerate the piston assembly and the reciprocating part of the connecting rod. 

A force Fcp acts on the crank pin and lower end of the connecting rod, directed radially inward toward the center of the crankshaft to produce the centripetal acceleration of the parts revolving with the crank pin. 

Fa will be the force acting on the engine frame due to Fp, and is equal and opposite Fp.

Fct will be the force acting on the engine frame due to Fcp, equal and opposite Fcp.  This force is usually balanced by the arrangement of cranks in a multicylinder engine, or by suitable counterweights on the crankshaft.  If Fct is so balanced, the only unbalanced force acting on the engine is Fa.

[Taylor Vol 2 Ch 8 page 247-248]

As previously discussed, Inertia Force F = Q1 Cos q + Q2 Cos 2q where q  = angle between crankpin and cylinder axis

Q1 Cos q represents the Primary inertia force (1st Harmonic), occurring at engine speed due to Cos q term, while Q2 Cos 2q represents the Secondary inertia force (2nd Harmonic), at twice engine speed due to Cos 2q.  A complete discussion of primary and secondary inertia forces was done in response to Question #1 part III concerning balance shafts, and will not be repeated here.

 


In a single crankshaft, the sum of the centrifugal forces of the counterweights must be equal and opposite to the centrifugal force NR.  Therefore, 2 Ncwt = NR    


 Balance of the centrifugal inertia force KR is achieved by the two counterweights at the continuation of the web, with their center of gravity at a distance r from the center of the crankshaft axis such that 2mcwR · r · w2 = mR · R · w2

 
 The secondary inertia force, PjII is never balanced.  The primary inertial force unbalance in PjI » 0.5 PjI results as a consequence of installing counterweights, transferring this value from the vertical plane to the horizontal.  The above diagram shows that the counterweight has 2 components, Rcw,v and Rcw,h – the vertical and horizontal components of Rcw.  Rcw,v reduces PjI but the horizontal component occurs in the engine.

The addition of counterweights will decrease vertical shake as the vertical forces are balanced ( Rcw,v = PjI ), but the horizontal primary inertia force is now unbalanced and introduced when the crank rotates to 120 degrees.

The mass of the counterweight is computed from the conditions 2 mcw,j r = 0.5 m R, or 2 mcw,j  = 0.5 mj R / r

[Ubong Winter 99 ME672 lecture Week3 Part b]

The cause of vertical shake in a simple single cylinder engine is the unbalanced primary inertia force (PIF).  For even number of cylinders, the configuration of the crankshaft and counterweights would solve these problems.  Unfortunately, as we have demonstrated the addition of counterweights transfers the vertical shake to horizontal shake.  Thus, twin balance or countershafts are needed here, as they are often used to eliminate or balance primary couples and secondary forces when opposing crank throw configurations and counterweights are not effective.

PIF balance is achieved by using two counterweights, with each weight being half the primary reciprocating inertia force, rotating opposite directions of each other at crankshaft speed.

Figure 6 below shows these PIF balancing when the crankshaft approaches TDC, where the PIF is greatest.  The twin countershafts are directed downward to offset this PIF.


The balance shafts cancel out Fp to eliminate vertical shake, and the mass of the balance shafts B1 and B2 cancel each other out when Fp=0  to give no horizontal shake either.

Rotating through a full 360 degrees gives the following:


 

[Ubong Winter 99 ME672 lecture Week4 Part a]

However, there are still secondary inertia forces (SIF) which contribute to vibration and shake as well.  SIF from the reciprocating mass (piston group mass + 1/3 mass of connection rod) has a magnitude equal to ¼ of the PIF.

These unbalanced SIF can be balanced by twin countershafts spinning in opposite directions at twice the crankshaft speed, and is graphically shown in the Figure 7 below.

 
 

 Again, rotating the crank through a full 360 degrees provides the following:


 

[Ubong Winter 99 ME672 lecture Week4 Part a]

Since Q1 and Q2 are sinusoidal, they may be balanced by sinusoidal forces having components always along the cylinder axis, such as with counter–rotating masses that always cancel two rotating components.   For a single cylinder engine, this means 2 weights rotating at engine speed, one on the crankshaft and one on an independent shaft, and 2 weights rotating on 2 separate shafts at twice engine speed to balance secondary forces.  This will create a rotational force equal to ½ the primary force, and reduces maximum vertical shake, leaving a horizontal shake equal to ½ Q1.

 [Ubong Winter 99 ME672 lecture Week3 Part a]

 A complete diagram of all 4 balance shafts is shown below in Figure 3.  All 4 are needed to completely address both PIF and SIF in a single cylinder engine.

 
 The countershafts A and A’ spin in opposite directions but at the same speed as the crankshaft to counteract the primary inertia forces.  Two smaller countershafts, also spinning in opposite directions but at twice the speed of the crank, counteract the secondary inertia forces.

 The sum of cf is balanced by counterweights attached on the prolongation of the crank web.  This system of counterweights also balances the first and second harmonics of the inertia forces.  To balance the primary inertia force the use of counterweights as mentioned above are used.  Two shafts with counterweights, shown in Figure 3 below as A and A’ are mounted on both sides of the crankshaft axis and rotated in opposite directions (one clockwise, the other counterclockwise) at the same speed as the crankshaft.

Question #2

    (Carry out all the calculations for this problem in the “inconsistent” English units, with lbf and lbm)   In the preliminary steps for the design of a four-stroke-cycle, spark-ignition engine, the objective has been set at a maximum brake power of 180 hp at 5000 rpm. For the purpose of the initial sizing, the following assumptions are to be made for the maximum power condition:

air/fuel ratio = 12.3 lbmair/lbmfuel

brake thermal efficiency = 27% (based on an LHV = 19,100 Btu/lbm)

volumetric efficiency = 75% (based on intake at 29.50 in.Hg (abs) and 75°F.)

Treat air as an ideal gas with gas constant = 53.3 lbf•ft/lbm•°R. Assume further that, for satisfactory combustion and smooth operation, the bore must not exceed 4 in. and the stroke 3.5 in. Moreover for ease of balance and packaging, the choice of the number of cylinders is restricted to 4 or 6 or 8.

Answer the following:

(a)    What engine displacement is required to satisfy the assured conditions?

 229.77 cubic inches

             (b)   What is the smallest acceptable number of cylinders?

 6 cylinders

  
  

 
 

 
 

Question #3

         The following values for the crank angles at inlet and exhaust valve opening and closing are fairly typical for four-stroke-cycle engines:

                        IVO = 10-20°BTC                              IVC = 40-55°ABC

                        EVO = 50-60°BTC                             EVC = 10-15°ATC.

These values correspond to significant overlap, compared with valves opening and closing at the beginning and ends of the intake and exhaust strokes. Explain how the overlap improves engine breathing at high speed. Identify other design issues that may be important in determining engine breathing.

Answer:

 Explain how overlap improves engine breathing at high speed.

 Overlap is the period of time where both inlet and exhaust valves are open at the same time in the TDC region, between the exhaust and intake strokes.  In this example this period is between 10-20°BTC and 10-15°ATC.  During this period the piston actually moves very little and almost remains stationary as the crank and connecting rod “rocks over”. 

As the piston moves up toward TDC, EVO occurs at 50-60°BTC and the hot exhaust gases begin moving out the exhaust valve.  At 10-20°BTC the intake valve opens, but the exhaust valve is already open and will remain open until EVC at 10-15°ATC.  By keeping the exhaust valve open until 10-15°ATC, and assuming that there is very little backpressure so that exhaust pressure (Pex) is less than inlet pressure (Pin), the inertia of the gas already flowing out the exhaust valve should cause the pressure inside the combustion chamber (Pc) to be less than Pin as well, causing the fresh intake charge to be sucked into the chamber, thus improving engine breathing.  If either the intake or exhaust valves were not open during this period, this fresh intake charge would not be sucked in and no improvement would occur.

A large overlap period encourages the removal of any remaining exhaust gases in the combustion chamber while providing an early start to induction.  Unfortunately, this phenomenon is mainly applicable only at high engine (piston) speeds.  At low engine speeds there are two detriments to having a large overlap period.  If there is a large exhaust lag, at low engine speeds there may be sufficient time for fresh charge to be drawn into the exhaust port before the exhaust valve closes, causing unacceptably high hydrocarbon and CO emissions due to the large amounts of unburnt or partially burnt gases escaping.  Additionally, at low engine speeds where the engine is heavily throttled and the intake valve is opened very early, and/or high exhaust backpressure exists, Pex may be greater than Pi.  Consequently, the inertia of the exhaust gases is not sufficient to cleanse the combustion chamber of all residual gases; rather, some residuals occupy the clearance volume and may be pulled back into the intake, diluting the fresh charge.  This upsets the AF ratio and promotes slower burning and incomplete combustion, again resulting in unacceptably high hydrocarbon and CO emissions for production automotive use.  This resulting incomplete combustion is especially evident at idle, and accounts for the rough idle instability of competitive high performance racing machines.  Therefore, the amount of overlap is always a compromise between increased high RPM power and decreased lower RPM performance.  Eliminating this compromise is one of the goals of a variable valve timing system where overlap may be calibrated as a function of engine speed.

In supercharged or turbocharged SI engines, a disadvantage of large overlaps is poor fuel economy as the extremely high intake pressures will literally blow raw fuel into the exhaust during the overlap period.  With CI engines, Pex/Pi is never much greater than 1, so some disadvantages of large overlap at low engine speeds experienced in SI engines (such as rough idle) are not a problem.  It should be noted that the largest overlaps occur in turbocharged diesel engines as the fresh air (without fuel) flowing through the cylinder during overlap cools the turbine temperature.

[Heisler 2nd Ed, 17.1.5 page 426-427]

[Ubong Winter 99 ME672 lecture Week 11Part A]

Identify other design issues that may be important in determining engine breathing.

 The ability of an engine to breathe directly affects volumetric efficiency.  Some of the factors affecting engine breathing are:

 Ø       Flow resistance (frictional losses) in the intake system.  Flow resistance of components in the intake system are additive.  These components include the intake ductwork, the air filter and air cleaner assembly, the mass airflow sensor (if equipped), the throttle body (including throttle blade), the intake manifold, intake ports, and valve openings.   In an older carbureted system the mass airflow sensor and throttle body would be replaced by the carburetor. 

 Increasing flow resistance negatively affects an engine’s breathing ability.  This flow resistance can be expressed in terms of pressure drop, where the total pressure drop is the sum of the pressure loss in each component of the intake system.  During the intake stroke, the cylinder pressure is less than atmospheric pressure by an amount dependent on the square of the speed.  Typically the cylinder pressure is 10-20% lower than atmospheric pressure when the piston speed is close to maximum.

 The shape of the valve ports are very important to flow.  Basic considerations are to have minimum protrusion of the guide boss into the port pocket and to have the largest possible radiuses in order to reduce flow resistance.

The amount of valve lift also affects flow resistance, with increasing lift comes increased valve curtain areas, increasing volumetric efficiency.  As a general rule, the lift/diameter ratio should not exceed 0.25 as that is a mechanical stress limit.  Increasing lift beyond certain limits puts too much stress on the valvetrain system, as the cam profile ramp rates become unrealistic and problems with valve/pushrod/rocker geometry and valvespring oscillations become evident.  Another means of increasing valve lift is to increase the rocker ratio on pushrod type engines.

Valve sizing obviously influences flow resistance as well.  Increasing valve size as much as physically possible will reduce frictional losses and increase an engine’s breathing ability.

[Ubong Winter 99 ME672 lecture Week 11 Part A and Part B]

 Ø       Heating of the inlet air temperature / Density of the inlet air.  Although this does not affect the volume of air flow, it does affect the density of the air in the same volume, and it is density, not volume, that directly affects volumetric efficiency.  If air density is reduced by heat anywhere between the inlet and the cylinder, the air mass induced will decrease, lowering volumetric efficiency.  Inlet air heating effects are greater at low engine speeds due to longer residence times of the intake charge, giving it more time to absorb heat.

 [Ubong Winter 99 ME672 lecture Week 11Part A]

Air density is of course also affected by altitude.  Air at high altitude has less density, hence less air mass will be induced, lowering volumetric efficiency.

Ø       Ram effect.  The air mass inducted into the cylinder is almost entirely determined by the inlet port pressure level during the short period immediately before the intake valve is closed.  This inlet port pressure varies greatly during each cylinder’s intake process due to piston velocity variations (even at a constant engine speed the piston velocity is always accelerating and decelerating due to it’s reciprocating motion – velocity is 0 at TDC and BDC), valve open area variation, and the resulting unsteady gas flows from these geometric variations.  At higher engine speeds, this pressure at the inlet port is increased due to inertia from the intake charge.  As engine speed is increased this pressure becomes progressively greater.  By keeping the intake valve open until 40-60 degs ABDC, this increased pressure will continue forcing fresh charge into the intake.   

An inevitable consequence of closing the intake valve late to take advantage of this ram effect at high engine speeds is that at low engine speeds, a reverse flow of fresh intake charge back into the intake port (backflow) may occur as cylinder pressure increases due to the piston moving back toward TDC on the compression stroke.  Thus at low engine speeds, volumetric efficiency is decreased by closing the intake valve late.  This suggests that at low engine speeds an earlier than normal inlet valve closing will reduce backflow losses and increase volumetric efficiency , but at the expense of reduced airflow at high speed. 

Ubong Winter 99 ME672 lecture Week 11Part A and Part B]

Ø       Exhaust Tuning.  Assuming the exhaust manifold combines each cylinder’s independent exhaust charge from the primary pipes into a single collector, the pulsating flow from each cylinder’s exhaust charge produces pressure waves which interact with the pipe junctions and terminations of the exhaust system.  Oppositely moving pressure waves will be produced from reflection of pressure waves at the boundaries of this piping system.  These boundaries, include junctions at the open or closed end of the pipe, any bends and gradual or sudden changes in area within the pipe.  

Additionally, in 4 stroke SI engines with long valve overlap periods, some in-cylinder colder air will be induced into the hot exhaust during this overlap period, and pressure waves propagating through these exhaust pipes will encounter gas at varying temperatures and reflections result.   

[Blair, 2.5 page 193 and 2.6 page 196] 

These interactions cause pressure waves to be reflected back towards the engine, and with multiple cylinders exhausting into a common collector, all these individual pressure waves will interact, either aiding or inhibiting the gas exchange process.  A tuned exhaust is the result of varying the pipe bends and dimensions so that the pressure in the exhaust port at the end of the exhaust process is reduced, thereby aiding the gas exchange process.  This tuning can be optimized only for a specific engine speed range.  

The length of the exhaust primary tube (runner) has a great effect on the speed of the exhaust gas.  When the exhaust valve opens, the exhaust gas exits as a high-speed pulse, leaving immediately behind an area of much lower pressure.  Given a primary tube of sufficient length, this low pressure remains until the next time the exhaust valve opens, thus providing an immediate area of lower backpressure to help scavenge the cylinder, allowing it to empty more completely.   

[Ubong Winter 99 ME672 lecture Week 11 Part B]  

Ø       Inlet tuning.  Inlet tuning follows the same principles as exhaust tuning.  In naturally aspirated engines, time varying inlet flow causes expansion waves to be propagated back into the intake manifold.  These expansion waves can be reflected at the open end of the manifold at the plenum causing positive waves to propagate towards the cylinder.  By tuning the intake, these positive waves will be timed so that they raise the pressure at the intake valve at the end of the intake process above nominal inlet pressure, thus increasing the inducted air mass.  

[Ubong Winter 99 ME672 lecture Week 11Part A]

Long inlet pipes may also be utilized to improve volumetric efficiencies at certain engine speeds due to the inertia and elasticity of the gases in the inlet pipe and cylinder.  Long pipes with smaller cross sectional area generally improve low speed performance at the expense of high speed power while shorter pipes of larger cross sectional area favor high speed at the expense of decreased low speed performance.  Long pipes with small diameter/bore ratios improve volumetric efficiency at low piston speeds due to a high level of kinetic energy built up in the pipe at the end of the induction process, but at high speeds flow restrictions overcome this kinetic energy advantage and volumetric efficiency falls.  Long pipes with large diameter/bore ratios move this power band up to intermediate piston speeds, but also suffer at high speeds due to the slow acceleration of the air mass.  

Ram tuning of the induction system: This phenomenon is based on the resonance of the columns of air in the pipes by selecting lengths of induction pipes such that they resonate at a fundamental frequency close to that at which the inlet valves open.  This aids the pressure wave to drive the charge into the cylinder, much like supercharging without actually using an external blower.

 [Ubong Winter 99 ME672 lecture Week 11Part B]  

With artificially (externally) supercharged or turbocharged applications, inlet tuning does not offer as much benefit as compared with naturally aspirated engines since the inlet pressure is artificially raised substantially above nominal inlet pressure already.  

Ø       Stroke/bore ratio:  A long stroke coupled with low engine speed has greater volumetric efficiency than a short stroke at high engine speed.  

Ø       Compression ratio:  A high compression ratio will produce less residuals, and thus have a higher volumetric efficiency. 

[Ubong Winter 99 ME672 lecture Week 11Part B]  

Ø      External superchargers or turbochargers.  External devices may be placed on the inlet to increase inlet pressures to drive the intake charge into the combustion chamber, thus improving power output.  

Superchargers fall into different categories depending on their design, whether they be roots type blowers or centrifugal, but they all share the basic characteristic of being parasitic, as they are usually driven in some manner by the engine crankshaft.  Hence they require engine power to operate; however the power gained by use of forced induction will overcome by far the parasitic loss from operating the blower.  Advantages of a supercharger over a turbocharger is their instantaneous response, as the compressor is driven directly off the crankshaft, and for a roots type blower they are easily packaged within the “v” of the engine between the banks.  

Turbochargers also increase inlet pressures but do not directly require crankshaft power to operate.  They are spun by the exhaust energy, so there are no parasitic losses from their use.  However, they suffer from a time lag as boost cannot be produced instantaneously on demand as they require sufficient exhaust energy to spin their compressor.  Nonetheless, proper calibration with timing and fuel delivery on the part of the engine management system can help minimize this turbo lag, as things such as reducing spark advance can force the exhaust gas temperature to rise, providing more exhaust energy to initially spin the compressors up to speed.  Turbochargers are not as prevalent today in production SI automotive engines due to tighter government mandated emission controls, especially in the area of cold starts.  Turbos suffer in the area of emissions as they inhibit the catalytic converter from reaching it’s necessary light off temperature where it’s catalyst become effective.  The turbo is such a large heat sink that there is considerable lag between the time the engine is started until the catalyst becomes operational.  Fortunately, advances such as electrically heated catalysts are being developed which may once again allow turbochargers to be used in production automotive environments on SI engines.  Turbochargers are widely available in many commercial CI engines today.  

Both turbo and superchargers increase the inlet air temperature as they compress the air, so in each case an intercooler (or aftercooler) to remove this heat from the intake air would increase the intake charge density even more.

Ø      Question #4

            The length of the intake system (working back from the valve, through the port, the intake runner, the plenum where the runners join, and so on back to the air cleaner)  would be filled with enough fresh charge to fill each cylinder. Look under the hood of your car to get reasonable estimates of cross-section areas of the various portions of the intake system. The capacity  of the 4 cylinder engine is 3.0 liters.

 Answer:

 I don’t quite understand the question here, I believe the objective is to compare the actual measured intake volume to see how closely it compares to the calculated volume of each individual cylinder event on our engine. 


 

This is an underhood shot of my car.  The engine is a 3.8L 90 degree even-fire Buick V6 with splayed crank journals and a balance shaft.  The displacement is 3800cc or 231ci, so each cylinder event would require 38.5 ci or 633.33 cc of air.  There is a roots type supercharger, an Eaton M90 model, on it.  This supercharger sits on an intake manifold shown below, which distributes the intake charge to each individual cylinder. 


 

Each port is approximately 3 cm wide by 5 cm tall, with an area of 15 cm^2.  The length of the intake runner is approximately 7 cm before it reaches the valve, so each intake runner volume is approximately 105 cc.  The plenum volume of this intake manifold appears to be about 25cm long  x 15 cm wide x 7 cm high, giving a total plenum volume of 2625 cc.  Divided by 6 this is 437.5 cc.  Thus the total intake volume in the intake manifold per cylinder is 437.5+105 = 542.5 cc. 

It is difficult to estimate the volume inside the roots blower as the lobes take up most of that space with very little free volume left.  I would estimate that the volume in the blower intake, with a 7 cm opening for a length of about 7 cm, is (3.5)^2 * Pi * 7 = 270 cc.  Again, divided by 6 cylinders this is approximately 45 cc per cylinder. 

The throttle body, which includes an integrated Hitachi mass airflow meter, is approximately 12 cm long, and has a 7 cm opening.  Thus it’s volume is 3.5^2*Pi*12 = 462 cc.  Again divided by 6 cylinders, this is 77 cc per cylinder.  However, the throttle blade sits approximately 5 cm into the throttle body, so we can effectively reduce its length to 5 cm since anything after the throttle blade probably should not be included for this exercise.  Thus, this volume becomes 3.5^2*pi*5 = 192.4 cc, and divided by 6 this is 32 cc. 

Adding up this volume, 542.5+45+32 gives 619.5 cc, which is remarkably close to the 633.33 cc each cylinder event would require.  

The cross sectional areas of any components after the throttle blade (between the throttle blade and the atmosphere) is probably irrelevant to this exercise.  The presence of a blower on this engine may also diminish the importance of having sufficient plenum volume as there is now an excess of available air with the forced induction.

Question #5

 Explain how each of the following design and operating changes affect the mass burning rate in a spark-ignition engine. Assume that the bore, stroke, compression ratio, speed of the engine, as well as the temperature and pressure of the incoming mixture are not changed. The changes in question are:

·        moving the spark plug closer to the center of the combustion chamber;

·        changing valve and port design so as to increase swirl;

·        changing the head design so as to make the combustion chamber more nearly hemispherical, with the spark plug near the center;

·        using two spark plugs instead of one;

·        increasing EGR.

Answers:

 Ø       Moving the spark plug closer to the center of the combustion chamber

 Moving the spark plug closer to the center of the combustion chamber should approximately double the flame speed because the center plug location gives approximately twice the flame area of a side plug geometry at a given flame radius.  The larger the flame front surface area, the more fresh charge is able to contact this surface and ignite. 

[Heywood 9.3.1 page 394] 

Combustion time is also affected due to the different distances which the flame must travel.  By placing the spark plug near the center of the chamber as opposed to near the edge, the flame travel distance is decreased, therefore decreasing combustion time. This decrease in combustion time due to both the increase in flame speed and the decrease in travel distance results in a slight reduction in the tendency to detonate as well, since there is less time for cylinder pressure and temperature to increase to the autoignition point of the end gas. 

[Taylor Vol 2, Ch 1 page 26]

[Ubong Winter 99 ME672 lecture Week 9 Part B] 

This is, of course, assuming a homogeneous mixture of a consistent, uniform air-fuel ratio in every part of the combustion chamber.  If this were not true, locating the spark plug in an area where the air-fuel charge happened to be optimized for ignition would aid flame kernel development and probably produce the highest initial flame speed. 

Ø       Changing valve and port design so as to increase swirl 

Generally increasing in-cylinder gas velocities with intake generated swirl increases the burning rate.  Swirl produces higher turbulence inside the combustion chamber and increased the rate of flame development and propagation. 

[Heywood 9.3.1 page 394, 8.3.1 page 343 and 15.4.1 page 846] 

However, this is also dependent upon the placement of the spark plug.   A spark plug located at the side of the combustion chamber where the flame propagation is in the same direction of the swirl would benefit from this swirl in reduced combustion time, while a spark plug placed at the opposite end of the combustion chamber where the flame propagation is against the flow of the swirl would suffer from increased combustion time.  

[Taylor Vol 2, Ch 1 Page 26]

[Ubong Winter 99 ME672 lecture Week 9 Part B] 

Swirl would also have positive effects on the flame initiation and propagation of a non-homogenous charge mixture, because the combination of swirl and a long duration spark event would increase the probability that an optimized air-fuel charge would come in contact with the spark and ignite. 

To a lesser extent, the burn rate may be slightly affected by turbulence made up of vortices, Flame speed will depend on whether ignition begins in the vortex center or boundary, resulting in cyclic variation.  Swirl will force the spark event to pass through several vortices and vortex boundaries, resulting in more stable combustion. 

[Taylor Vol 2, Ch 1 Page 30-31]

[Heywood 15.4.1 page 846] 

Ø       Changing the head design so as to make the combustion chamber more nearly hemispherical, with the spark plug near the center

 Hemispherical chambers and other open chambers have nearly the maximum flame front surface area and therefore have a faster burn rate for the same reasons as described above.  For a given engulfed volume, the open hemispherical chamber gives flame surface areas of 30% more than equivalent disc configurations.  As mentioned above, locating the spark plug near the center provides for maximum surface area and a resulting faster burn rate as well.

 [Heywood 15.4.5 page 857] 

Unfortunately, the spark plug cannot be place exactly between the two valves due to room constraints and must be slightly offset.  This forces a long flame path, so there is a tendency to detonate, especially when compression ratios are greater than 8.5:1. 

[Ubong Winter 99 ME672 lecture Week 9 Part B]

Ø       Using two spark plugs instead of one

 Using two spark plugs instead of one should not have any appreciable effect on the actual flame propagation velocity (flame speed) given the same fixed spark timing for both cases.  

[Taylor Vol 2, Ch 1 page 26] 

One would believe that having multiple spark plugs would increase flame speed because there would be multiple flame fronts generated and thus more flame area compared to a single center plug geometry at a given flame radius.  The more flame front surface area, the more fresh charge is able to contact this surface and ignite.  Unfortunately, it appears that having two spark plugs at opposite sides of the chamber does not significantly differ in enflamed volume from the single center plug because the flame front areas are comparable once intersected by the cylinder wall. 

[Heywood 9.3.1 page 394] 

However, with spark timing optimized for the situation, combustion times decrease with multiple plugs.  Large differences in combustion time would occur simply due to the different distances which the flame must travel, as distances would be shorter with multiple plugs. This results in a slight reduction in the tendency to detonate as well, since there is less time for cylinder pressure and temperature to increase to the autoignition point of the end gas. 

[Taylor Vol 2, Ch 2 page 75]

[Ubong Winter 99 ME672 lecture Week 9 Part B] 

This is, of course, assuming a homogeneous mixture of a consistent, uniform air-fuel ratio in every part of the combustion chamber.  If this were not true, having multiple spark plugs would increase the probability that at least one plug would be located in an area where the air-fuel charge happened to be optimized for ignition, aiding flame development and probably  producing the highest initial flame speed. 

Ø       Increasing EGR

 Increasing EGR will dilute the intake charge and this will reduce the flame speed.  

[Taylor Vol 2, Ch 1 Page 20]

[Heywood Ch 9.3.4 page 395] 

To compensate for this reduction in flame speed and combustion temperature (cooled EGR reduces combustion temps), spark timing is advanced when EGR is artificially induced (such as through an EGR valve) in order to maintain optimum spark advance just under the knock (detonation) limit to try to keep peak cylinder pressure as close to the optimum crank angle (15 to 20 degs ATDC) for best power.  Dilution of the intake charge also reduces the tendency to detonate, further necessitating a change in spark advance.  (This is assuming that the engine is knock limited due to insufficient octane of the fuel, so that it is impossible to achieve MBT spark timing as the octane requirement for MBT spark is too high.  This is typical of most production passenger car automotive engines.)  This spark advance is proportional to the amount of EGR added, both factors are calibrated values scheduled and controlled by the engine management system.  EGR is calibrated so that desired emission goals are met as long as misfire and unstable combustion are not occurring, and spark advance relative to EGR is calibrated corresponding to the amount of EGR added to be just below the knock limit.
 
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