Mechanical Engineering

DrawWorks Brake System Training Course (Part I)

amrxp2005@gmail.com (Engineer) — Sun, 24 May 2009 04:19:25 PDT

DW Brake System
(ZDPE/JC-70DB)
Training Course

Agenda
1- Course Objectives
2- Introduction
3- Hydraulic Pumps and Pressure Regulation
4- Pneumatics
5- Types of control valve
a. Poppet valves
b. Spool valves
c. Pilot-operated valves
d. Check valves
e. Shuttle a valves
f. Proportional valves
6 - The DW Brake System construction
7 - The Brake System construction
8 - The Brake System Operation
9 - The Brake System Maintenance and Troubleshooting

1. Course Objectives
To understand:
• The operation of the valves, pumps and hydraulic components of the DW Brake System.
• The Hydraulic concepts behind the DW Brake System.
• The DW Brake System Construction
• The DW Brake System Operation
• The DW Brake System Maintenance and Troubleshooting
• Complete one written test & achieve an overall pass mark of 80%

2. Introduction
Most industrial processes require objects or substances to be moved from one location to another or a force to be applied to hold, shape or compress a product. Such activities are performed by Prime Movers; the workhorses of manufacturing industries. In many locations all prime movers are electrical. Rotary motions can be provided by simple motors, and linear motion can be obtained from rotary motion by devices such as screw jacks or rack and pinions. Where a pure force or a short linear stroke is required a solenoid may be used (although there are limits to the force that can be obtained by this means).

Electrical devices are not; however, the only means of providing prime movers. Enclosed fluids (both liquids and gases) can also be used to convey energy from one location to another and, consequently, to produce rotary or linear motion or apply a force. Fluid-based systems using liquids as transmission media are called hydraulic systems (from the Greek words hydra for water and aulos for a pipe; descriptions which imply fluids are water although oils are more commonly used). Gas-based systems are called Pneumatic systems (from the Greek pneumn for wind or breath). The most common gas is simply compressed air. although nitrogen is occasionally used.

2.1. Hydraulic Fundamental Principles (Pascal`s law)

Figure (1) Forces and pressure in closed tanks

· Pressure in an enclosed fluid can be considered uniform throughout a practical system.
· This equality of pressure is known as Pascal's law, and is illustrated in Figure (1) where a force of 5 kgf is applied to a piston of area 2 cm2

. This produces a pressure of 2.5 kgf cm-2 at every point within the fluid, which acts with equal force per unit area on the walls of the system.

Figure (2) Mechanical advantage

· This expression shows an enclosed fluid may be used to magnify a force.
· In Figure (2) a load of 2000 kg is sitting on a piston of area 500 cm2 (about 12 cm radius). The smaller piston has an area of 2 cm2.
· An applied force f given by;

will cause the 2000 kg load to rise.
· This is called to be a mechanical advantage of 250. Energy must, however, be conserved.

· To illustrate this, suppose the left hand piston moves down by 100 cm (one meter).
· Because we have assumed the fluid is incompressible, a volume of liquid 200 cm2 is transferred from the left hand cylinder to the fight hand cylinder, causing the load to rise by just 0.4 cm.
· So, although we have a force magnification of 250, we have a movement reduction of the same factor.
· Because work is given by the product of force and the distance moved, the force is magnified and the distance moved reduced by the same factor, giving conservation of energy.

3. Hydraulic Pumps and Pressure Regulation
· A hydraulic pump (Fig. 3) takes oil from a tank and delivers it to the rest of the hydraulic circuit. In doing so it raises oil pressure to the required level. The operation of such a pump is illustrated in Figure 3.a.
· On hydraulic circuit diagrams a pump is represented by the symbol of Figure 3.b, with the arrowhead showing the direction of flow.
· Hydraulic pumps are generally driven at constant speed by a three phase AC induction motor rotating at 1500 rpm in the UK (with a 50 Hz supply) and at 1200 or 1800 rpm in the USA (with a 60 Hz supply).
· Often pump and motor are supplied as one combined unit. As an AC motor requires some form of starter, the complete arrangement illustrated in Figure 3. c is needed.
3.1. Pump Types
There are two types of pump illustrated in Figure 4.

1- Hydrodynamic Pump (Figure 4.a, )
Fluid is drawn into the axis of the pump, and flung out to the periphery by centrifugal force. Flow of fluid into the load maintains pressure at the pump exit. Should the pump stop, however, there is a direct route from outlet back to inlet and the pressure rapidly decays away. Fluid leakage will also occur past

Figure (3) The hydraulic pump

Figure (4) Types of Hydraulic Pumps

the vanes, so pump delivery will vary according to outlet pressure.
Hydrodynamic pumps (Fig. 4.a), are primarily used to shift fluid from one location to another at relatively low pressures.

2- Positive Displacement (hydrostatic Pump (Figure 4.b)
As the piston is driven down, the inlet valve opens and a volume of fluid (determined by the cross section area of the piston and the length of stroke) is drawn into the cylinder.
Next, the piston is driven up with the inlet valve closed and the outlet valve open, driving the same volume of fluid to the pump outlet.
Should the pump stop, one of the two valves will always be closed, so there is no route for fluid to leak back. Exit pressure is therefore maintained (assuming there are no downstream return routes).
More important, though, is the fact that the pump delivers a fixed volume of fluid from inlet to outlet each cycle regardless of pressure at the outlet port. Unlike the hydrodynamic pump described earlier, a piston pump has no inherent maximum pressure determined by pump leakage: if it drives into a dead end load with no return route (as can easily occur in an inactive hydraulic system with all valves closed) the pressure rises continuously with each pump stroke until either piping or the pump itself fails.

3.2. Pump Power

Fig. (5)Derivation of pump power

The motor power required to drive a pump is determined by the pump capacity and working pressure.

In Figure 5, a pump forces fluid along a pipe of area A against a pressure P, moving fluid a distance d in time T. The force is PA, which, when substituted into above Eq.

3.3-Filters

Figure (6) Filter Position
Dirt in a hydraulic system causes sticking valves, failure of seals and premature wear. Even particles of dirt as small as 20 microns can cause damage.
Filters are used to prevent dirt entering the vulnerable parts of the system, and are generally specified in microns or meshes per linear inch (sieve number).
See the three filter positions shown in Fig. 6

4. Pneumatics
4.1. Stages of air treatment
Air in a pneumatic system must be clean and dry to reduce wear and extend maintenance periods. Atmospheric air contains many harmful impurities (smoke, dust, water vapour) and needs treatment before it can be used.
In general, this treatment falls into three distinct stages, shown in Figure (7).
First, inlet filtering removes particles which can damage the air compressor.
Next, there is the need to dry the air to reduce humidity . This is normally performed between the compressor and the receiver and is termed primary air treatment.
Finally; the treatment is performed local to the duties to be performed, and consists of further steps to remove moisture and dirt and the introduction of a fine oil mist to aid lubrication.

Fig. 7 Three stages of air treatment

4.2. Air Dryers

Figure (8) Air filter and water trap
Air flow through the unit undergoes a sudden reversal of direction and a deflector cone swirls the air (Figure 8-b). Both of these cause heavier water particles to be flung out to the walls of the separator and to collect in the trap bottom from where they can be drained.
Water traps are usually represented on circuit diagrams by the symbol of Figure 8-c.

Figure (9) Refrigerated Dryer
Dew point can be lowered further with a refrigerated dryer, the layout of which is illustrated in Figure 9. This chills the air to just above 0~ condensing almost all the water out and collecting the condensate in the separator. Efficiency of the unit is improved with a second heat exchanger in which cold dry air leaving the dryer pre-chills incoming air. Air leaving the dryer has a dew point similar to the temperature in the main heat exchanger.

5. Types of Control Valves
Generally; the load is connected to ports labeled A, B and the pressure supply (from pump or compressor) to port P. In the hydraulic valve, fluid is returned to the tank from port T. In the pneumatic valve return air is vented from port R. See Figure 10.

Figure (10) Valves in a pneumatic and hydraulic system
Figure 11 shows internal operation of valves. To extend the ram, ports P and B are connected to deliver fluid and ports A and T connected to return fluid. To retract the ram, ports P and A are connected to deliver fluid and ports B and T to return fluid.

Figure (11) Internal valve operation
Another consideration is the number of control positions. Figure 12 shows two possible control schemes. In Figure 12-a, the ram is controlled by a lever with two positions; extend or retract. This valve has two control positions (and the ram simply drives to one end or other of its stroke).
The valve in Figure 12-b has three positions; extend, off, retract. Not surprisingly the valve in Figure 12-a is called a two position valve, while that in Figure 12-b is a three position valve.

Figure (12) Valve control positions
A complete valve description needs;
1- Number of Ports
2- Number of positions and
3- Action
Figure 13 shows one possible action for the 4/3 valve (Port/Position).This unload the pump back to the tank (without need of a separate loading valve), while leaving the ram locked in position.

Figure (13) One possible valve action for a 4/3 valve
Other possible arrangements may block all four ports in the off position (to maintain pressure), or connect ports A, B and T (to leave the ram free in the off position).
5.1 Valve Symbols
Designations given to ports are normally as shown:

In Figure 14-a, for example fluid is delivered from port P to port A and returned from port B to port T when the valve is in its normal state a. In state b, flow is reversed.

Shut off positions are represented by T, as shown by the central position of the valve in Figure 14-b.

The internal flow paths can be represented as shown in Figure 14-c. This latter valve, incidentally, vents the load in the off position.

In pneumatic systems, lines commonly vent to atmosphere directly at the valve, as shown by port R in Figure 14-d.

Figure (14) Valve symbols
Figure 15-a shows symbols for the various ways in which valves can be operated. Figure 15-b thus represents a 4/2 valve operated by a pushbutton. With the pushbutton depressed the ram extends. With the pushbutton released, the spring pushes the valve to position a and the ram retracts. Actuation symbols can be combined. Figure 15-c represents a solenoid-operated 4/3 valve, with spring return to centre

5.2 Poppet valves

5.2 Spool valves

5.3 Pilot-operated Valves

With large capacity pneumatic valves (particularly poppet valves) and most hydraulic valves, the operating force required to move the valve can be large. If the required force is too large for a solenoid or manual operation, a two-stage process called pilot operation is used.

5.4 Check Valves
Check valves only allow flow in one direction. The simplest construction is the ball and seat arrangement of the valve in Figure, commonly used in pneumatic systems. Free flow direction is normally marked with an arrow on the valve casing.

5.5 Shuttle Valves
A shuttle valve, also known as a double check valve, allows pressure in a line to be obtained from alternative sources.
It is primarily a pneumatic device and is rarely found in hydraulic circuits.
Construction is very simple and consists of a ball inside a cylinder, as shown in the Figure. If pressure is applied to port X, the ball is blown to the fight blocking port Y and linking ports X and A.
Similarly, pressure to port Y alone connects ports Y and A and blocks port X. The symbol of a shuttle valve is given in Figure.

5.5 Proportional Valves
The solenoid valves described so far act, to some extent, like an electrical switch, i.e. they can be On or Off. In many applications it is required to remotely control speed, pressure or force via an electrical signal. This function is provided by proportional valves.
A typical two position solenoid is only required to move the spool between 0 and 100% stroke against the restoring force of a spring. To ensure predictable movement between the end positions the solenoid must also increase its force as the spool moves to ensure the solenoid force is larger than the increasing opposing spring force at all positions.

Alignment Handout

amrxp2005@gmail.com (Engineer) — Mon, 13 Apr 2009 05:42:39 PDT

Factors Influencing Alignment procedure
1- Eccentricity (runout)
Check
· This might be done by a dial gauge

2- Baseplate of machines (soft-foot)
· Machines feet must be mounted perfectly horizontal with the baseplate.
· The contact between the baseplate and the feet can be checked with a set of shims or with feeler gages.
· During a new installation,
o it is essential to use accurate straight edges and levels to make sure that all feet of the machine are on the same plane.
· The accepted tolerance level for these planes is usually 0.1 mm.
· Simple tests for soft-foot are by setting up the dial gage (at fixed place) and place a shim under one front foot and the reading noted.
· It is then removed and placed under the next front foot. The reading should be the same.
· The same procedure must be repeated for the rear feet.

3- Axial position of machines
· The axial position of shaft ends is referred to as the distance between shaft ends (DBSE).
· Normally, most couplings allow a large tolerance in the axial position.
· However, for couplings like disk couplings, an error in the axial position result in;
o places the discs under stress and
o decreases their life.
o may generate axial thrusts, which ultimately add extra load to the machine’s thrust bearings.
· It is therefore necessary to take this aspect into consideration, especially when machines operate at high temperatures.

4- SAG
· For spacer couplings, a sag (deflection) check should be done on the indicator bracket to be used for the alignment.
· The DBSE in these couplings may be long, and when alignment brackets are clamped to one hub and extended to the other hub, there is a tendency for them to sag.
· This sag can alter the dial gage readings, leading to misinterpretation and errors.
· For bracket lengths larger than 25–30 cm, it is essential to provide additional stiffness to minimize sag.
· It is therefore necessary to perform a sag check of the bracket.
· A sag check is essential only for aligning horizontal machines, because the sag is caused by gravity due to the weight of the bracket.

Alignment techniques
· There are many methods to align a machine. The appropriate method is selected based on;
Ø the type of machine,
Ø rotational speed,
Ø the machine’s importance & production,
Ø the maintenance policy and
Ø alignment tolerances.

· Machines “I”
{Which are not fragile (breakable) in their construction}.
Ø rotating at less than 1500 rpm,
Ø lower horsepower range,
Use merely a straight edge to align machines.
Considering all aspects, it is acceptable to align them to the range of 0.3–0.8 mm.

Machines “II”
{Majority of machines}
Fragile (breakable) in their construction (mechanical seals and expansion bellows)
Machines operating at;
Ø speeds of 3000 rpm and higher,
Ø in the medium power range of 20 kW–1 MW
should be aligned within 0.1 mm.
· This requirement necessitates the use of comparators like dial gages, and methods with minimum residual errors.

Alignment conventions using a dial indicator
· The dial gage is the most common comparator used during alignment.
· The dial gage functions based on the rack-and-pinion principle. The conventions that are followed are shown in Figure 6.9.
· When the spring is compressed, the dial pointer is pressed inward and the clock needle moves clockwise, indicating a positive reading.
· When the pointer moves outwards, the clock needle moves counter clockwise, indicating a negative reading.
· It is recommended to jog the pointer from the top to ensure that it is not stuck.

The dial gage functions based on the rack-and-pinion principle. The conventions that are followed are shown in Figure 6.9.

Figure 6.9 Dial indicator

Another convention for alignment readings in the horizontal plane is shown in Figure 6.10.

Figure 6.10 Alignment readings in the horizontal plane

Thus, the convention maintains left and right when standing behind the driver, facing the driver.
Left and right readings on the dial gage are recorded accordingly.

Shaft setup for alignment
· The connection to the shaft must be simple and rigid.
· The clamp shown in Figure 6.11 is a good example. Magnetic clamps must be avoided, because their attachments are not reliable.

Figure 6.11 Shaft setup for alignment

· There are many types of alignment brackets available in the market, and a typical one is shown in Figure 6.12.

· The guiding principle for the selection of brackets is that they should be rigid with minimal sag (see the rod diameters).

Figure 6.12 Alignment brackets

Types of misalignment
Misalignment in machines is due to;
· angularity and
· offset,
· but in almost all cases the misalignment of machines is a combination of both.

i- Angularity
· is the difference between the values on the comparator (dial gage) for a half revolution”180o” (because for one complete revolution we return to the original position).
· For a given angular misalignment, angularity depends on the diameter described by the dial gage.
· It can be seen that when d1 increases to d2, p1 increases with the same ratio to p2. This value must be fixed when a certain tolerance is given (Figure 6.13).

Figure 6.13 Angularity (parallelism)

Angle of misalignment:

Where p1, p2 = dial gage reading when rotated by 180°; d1, d2 = diameters described by the dial gage.

ii- Offset (concentricity),
· The offset is the radius of rotation for the dial gage, as indicated in Figure 6.14.

Concentricity =1/2 dial gage reading

· The dial gage readings would indicate the diameter, and hence should be reduced by half to obtain the true offset reading.

Figure 6.14 Radial misalignment (concentricity)

However, as mentioned before, in practice misalignment of machines is due to a combination of both factors, as depicted in Figure 6.15.

Figure 6.15 Misalignment of shafts with angularity and offset

Two dial method of alignment
The necessary steps to align a machine are:
1. The first step is to loosen the coupling bolts so there is no restriction during the measurement of angularity of the existing misalignment.
2. A feeler gage is then run through the coupling hubs to ensure that the hubs are not touching.
The necessary steps to align a machine are:
i- The radial test (R) to measure the OFFSET ;
· The dial gage is attached as shown in Figure 6.16.
· The test done in the vertical and horizontal planes.
· To obtain the offsets in both planes, four readings will be required.
1. Top, bottom, left and right
2. Clock positions – 12 o’clock, 3 o’clock, 6 o’clock and 9 o’clock positions.
Clamping
· The dial gage here generally placed on the top (12 o’clock) position, and the zero on the scale is turned to coincide with the needle.
· The pointer must be jogged to ensure that it is free and that the readings are repeatable.

Figure 6.16 Dial gage setup at top position. The difference in readings after 180

indicates offset in vertical or horizontal planes

· Shafts are turned manually through one complete revolution, and readings at every quadrant (quarter) are noted.
· The readings recorded at the four locations are written down in the format shown below (fig. 6.17).
· The ‘R’ in Figure 6.17 indicates that these are radial readings, meant for offset corrections.

Figure 6.17Readings in mils

ii- The Facial reading to measure the ANGULARITY;
· The clamp is re-adjusted with the dial gage pointer now set to measure the angularity, as shown in Figure 6.18.
· The pointer (as shown in the figure) is now parallel to the axes of the shafts.
· Just like the offset, the angularity must be measured in horizontal and vertical planes as well.
· The dial gage is rotated through one complete revolution and stopped at every quadrant to make a note of the readings.

Figure 6.18 Dial gage setup at top position. The difference of readings
after 180° indicates angularity in vertical or horizontal planes

· The ‘F’ in Figure 6.19 indicates that these are facial readings, meant for angularity corrections.

Figure 6.19 The ‘F’ indicates facial readings
(note the diameter described by the dial gage)

iii-Steps to fix the alignment
· The next step is to convert these values of (R) and (F) to appropriate shim thickness that should be added or removed to fix the alignment.
To proceed to the next step, additional information about the location of the front and the rear feet from the dial gage pointer is required.

*In Figure 6.20
· The pump is the fixed machine (FM) and the motor is the machine to be shimmed (MTBS).
· This implies that all the corrections will be done by adding and removing shims under the motor feet. The pump will not be disturbed from its position.
· The distance from the pointer of the dial gage to the front foot (FF) of the motor is designated as ‘A’.
· The distance of the rear foot (RF) to the dial gage pointer is designated as ‘B’.

Figure 6.20 Shimming Calculation

· Two sets of calculations are required. One set for the vertical plane and the other for the horizontal plane.

1. Calculations for the vertical plane
*Offset correction
· Let us say the offset readings for the top and bottom positions are 0 and -5 mils, respectively.
· If the dial gage pointer is on the motor (MTBS) and the dial gage is rotating, hence the –ve and +ve signs are as shown in (figure 6.9).
· The negative sign indicates that the motor shaft is higher than the pump shaft.
· It is higher by half the final reading minus the initial readings. Thus:

Hence, shims of 2.5 mils should be removed from the front and rear feet of the MOTOR.

*Angularity correction
· Let us say the angularity readings for the top and bottom readings were 0 and - 2 mils, respectively.
· If the dial gage pointer is touching the rear face of the motor coupling hub see (figure 6.9).
· The negative sign indicates that the coupling has a narrower gap at the bottom than at the top.
The dial measures at (scribes) a circle of 5 in.

The angle

Because the angle is very small, the tan inverse function can be neglected:
Hence, P1=.002 in.

(The formula would reverse if the pointer is touching the front face of the coupling hub, which is normally the case when there is a long spacer between the couplings.)
= 0.0004 radian
=0.4 milli-radians = 0.0004´ (180/p) = (0.023o)

· This angle “θ” is also the angle of inclination of the motor axis w.r.t. the pump axis.

· Line AB is the existing axis inclination of the motor (Figure 6.21).
· It must be lifted by amount x at the FF (front foot) location and by y at the RF (rear foot) location.

Figure 6.21 Calculating X and Y values

· The x and y values are calculated as follows;
x and y are approximated as arcs and the following formula can be used:
S = r × θ

Where;
S = arc length;
r = radius;
θ = included angle in milli-radians.

Final- Vertical

The final results should include corrections for both the offset and the angular corrections.
At point A{Front Foot FF}:
Offset results – remove shims of 2.5 mils
Angularity results – add shims of 3.2 mils
Thus, insert shims of 0.7 mils under the front foot of the motor.

At point B{Rear Foot RF}:
Offset results – remove shims of 2.5 mils
Angularity results – add shims of 7.2 mils
Thus, insert shims of 4.7 mils under the rear feet of the motor.

Calculations for the horizontal plane
The dial gauge is: from behind the motor, left is the initial reading and right is the final reading.
*Offset calculations:
Left reading: +1 mils
Right reading: - 6 mils
Pointer on left;
+1 means the measured point on motor shaft is to the left by “1”
Pointer on the right;
- 6 means the measured point on motor shaft is to the left by “6”
{To imagine this, just draw the dial gage and its direction}
· Because the dial pointer is on the motor shaft, a negative reading indicates that the motor shaft axis is to the left of the pump shaft axis.

Offset = (Final reading – Initial reading) /2

Move points A and B of the motor to the right by 3.5 mils.

*Angular calculations:
Left reading: + 4 mils
Right reading: - 6 mils
+ 4 means that the left point of the motor hub is away to the pump hub by “4”.
- 6 means that the right point of the motor hub is close from the pump hub by “6”.
{To imagine this, just draw the dial gage and its direction}

As the dial pointer touches the rear face of the motor coupling hub, the shaft axis resembles what is shown in Figure 6.22.
In this case:
mils = 0.01 inch.
Thus:

= 0.002 radians (0.114o)
= 2 milli-radians

Hence:
x = 2 ´ 8 = 16 mils – move to the left;
y = 2 ´18 = 36 mils – move to the left.

Figure 6.22

Final- Horizontal

At point A{Front Foot FF}:
Offset results – move 3.5 mils to the right
Angularity results – move 16 mils to the left
Thus, move to the left by 12.5 mils.

At point B {Rear Foot RF}:
Offset results – move 3.5 mils to the right
Angularity results – move 36 mils to the left
Thus, move to the left by 32.5 mils.

The procedure would be;
· The vertical shim corrections should always be done prior to the horizontal shifts.

· Once the vertical shims are adjusted, the bolts should be tightened and a quick test of the vertical plane reading should be made to confirm the accuracy.

· If the accuracy is satisfactory, the bolts can be loosened and the horizontal alignment should be done with jack bolts (if provided).

The limitations of this method are:
· Calculations are necessary, which may be difficult to do in the field.
· It is beneficial to be able to visualize the shaft orientation from the dial gage readings but this requires practice.
· Inexperienced technicians can find this confusing.
· Errors in calculations may occur if there is bracket sag and/or error in the dial gage readings.
· If the shaft of one or both the machines has substantial axial floats, the angular readings can be erroneous.

Laser alignment
Alignment with comparators such as dial gages characterized by;
· a fair degree of precision,
· demand skill,
· demand training and
· Require experience.
Consequently, these methods are;
· tend to provide errors and
· can take a considerable amount of time.

The method of alignment using LASERS (Figure 6.34);
· overcomes the disadvantages listed above and
· it is gradually becoming the preferred method of alignment for most machines.
· Data collection and calculations have become;
o fast and
o accurate
· Some laser systems need less than a quarter turn of the shaft to produce very good shim correction data.
· They have built-in alignment tolerances, and hence there is no need for an expert to judge on the quality of the residual misalignment.
· Laser beams can travel over long distances, and alignment can be done very accurately with relative ease (comfort).
· Laser beams do not bend over great distances and for this reason the sag effect is entirely eliminated.

Figure 6.34 Laser alignment

· The laser alignment system (Figure 6.35) comprises;
Ø an analyzer and
Ø two laser heads.
· The laser heads are attached to the two shafts
· The laser heads must face each other, and each head has a laser emitter and receiver.
· When the shafts are turned, the receivers trace the movement of the laser beams.
· These values are communicated to the analyzer.
· Machinery data and the required distances are initially entered into the analyzer.
· The data from;
o the laser heads and
o the given machinery data
are used to accurately determine the shim corrections for the machine.
· Once the laser head and the reflector are installed, the shafts must be rotated.

Figure 6.35 Laser alignment system comprising of laser head,
reflector and analyzer (Prueftechnik – Optalign Plus system)

One emitter and one receiver system;
· Some entry-level laser alignment systems only have one laser emitter head and a reflecting prism on the other.
· These systems are ideal for general purpose machines.
· They eliminate the dial gages and provide an alignment calculator.
· The methodology with these systems is same as the previous one.
· At every quarter revolution, the analyzer must be activated to acquire the reading.
· After this, the analyzer provides the alignment correction information.

Some advanced LASER systems
Some systems include additional features that make alignment of machines an easy task.
These features are:
· Complex trains comprising of as many as five machines can be handled.
· Communications that eliminate cables between the laser heads and the analyzer.
· Errors due to vibrations from other machines can also be eliminated through averaging.
· Uncoupled and non-rotating machines can also be aligned.
· Less than a quarter rotation may be sufficient to obtain misalignment data.
· It is possible to do live horizontal alignment. This means that there is no need to take a reading and transfer it to the analyzer for calculation. The instant communication of the heads and analyzer accomplishes this automatically.
· One or two soft foot conditions can be identified.
· Once a machine is aligned, its history and data can be stored.
· They provide built-in misalignment tolerances.

Alignment tolerances
In practice, it is almost impossible to obtain;
· a zero offset
· and zero angularity,
and thus machines have to be left with a certain residual misalignment.
This residual misalignment has little or no detrimental effect on the operation of machines.

· The above values are assumed to be pure offset or pure angle.
· In practice, a combination of the two is more common and tolerances should account for this combination.
For example, a machine is running at 3000 rpm and the residual misalignment data is:
· offset: 2.6 mils
· angularity: 0.25 mil/in.
In pure terms, these values would be acceptable.
Nonetheless, let us see if the combination of the two is acceptable. To achieve this, a XY graph is made as shown in Figure 6.36.
If an offset of 2.6 with an angularity of 0.25 mils/in. is plotted, it could be beyond the acceptable range.

Figure 6.36 Alignment tolerances

ROTALIGN PRO
(Laser shaft alignment system)

1. Lock-out the machine.
2. Mount the chain brackets to the shafts.
3. Mount the emitter [waterproof, dust proof] on one bracket.
4. Mount the receiver [waterproof, dust proof] on the other bracket.
5. Turn on the emitter.
6. You have only one laser to adjust.
7. Only one cable is needed to connect the laser head to the hand-held computer.
8. Three short steps and you have your alignment data;
§ Dimension
§ Measure
§ Results
9. DIM
§ Press the DIM key your screen displays to you some data to put in.
§ Then determine the dimension of your machine and put them into the computer.
10. Measure (M)
§ Press measure key then you are ready to begin.
§ Turn the shafts for 1/4 (quarter) turn or less and forget about the clock position
11. Results
§ Press the result key, the screen displays as found to scale the alignment conditions AND the corrections needed.
12. Do physically the corrections required for the machine and repeat your job UNTIL the computer tells you that you have a good alignment.
(This is appeared with the sign on the screen).
13. Some benefits of this laser system are;
§ Determine the soft foot condition and analyze it and give the suggestions.
§ It can provide alignment for non-rotating shafts.
§ It can automatically the thermal expansion and gives the corrections.
§ It is capable of aligning the vertical machines.
§ The software can be updated from the website.

Alignment Handout

amrxp2005@gmail.com (Engineer) — Mon, 13 Apr 2009 02:31:07 PDT

Factors Influencing Alignment procedure
1- Eccentricity (runout)
Check
· This might be done by a dial gage

Power Transmissions

amrxp2005@gmail.com (Engineer) — Sun, 22 Feb 2009 01:48:13 PST

GENERAL CONSIDERATIONS
The first decision in designing an engine installation is selection of the coupling and drive method to connect the engine to the driven equipment..
The coupling and drive selection connections are closely related to the proper selection of engine support and mounting. This ensures a successful trouble-free installation from the standpoint of both the engine and driven equipment, as well as the power transmission components. (Refer to Mounting and Alignment section.)

A rigid precision-type mounting system must be provided for both the engine and driven equipment if a solid or nearly solid driveline is utilized.

Drive components which utilize universal joints, drive shafts or belts, and chain-type drives permit slightly greater alignment deviations.

When selecting the power transmission system, the possible need for a complete torsional analysis must be considered. System incompatibility will result in premature and/or avoidable failures.Refer to Mounting and Alignment section

CLUTCHES
General Description and Selection Considerations
Engine starting capability is normally limited and the direct connection of large mass
driven equipment makes starting difficult or impossible, therefore, a type of clutch or
disconnect device may not only be desirable but necessary.

Exceptions, if properly sized to the engine starting capability, may be centrifugal pumps, fans or propellers, and generators which provide a direct connected load with
a low starting torque requirement. Certain compressors which utilize a starting “unloading device” may also be direct connected.

Piston-type pumps, most compressors, belt- and chain-driven equipment, and all mobile vehicles will require an engine disconnect system.

The engine disconnect feature provides an important safety and service function. It permits rotating the engine for service and adjustment, as well as servicing the driven
equipment without disconnecting the drive-train. It also permits engine warm up before applying load — an accepted requirement for extended engine life. On multiple engine installations driving into a common compound or driven machine, it permits operating at less than full power level if desired, as well as at partial power should one engine be down for routine service or because of failure.

Numerous devices are available for connection or engagement of the engine to the driven machine. The device selection will depend on the desired engagement function; however, several general considerations must be made regardless of the
device selected.

The selected device must have adequate capacity to transmit the maximum engine
torque to the driven equipment. With the exception of “dog-type” clutches, which are
generally not acceptable on material handling equipment, clutches rely on friction
for power transmission.
(Dog-type clutches provide a direct mechanical connection and cannot be engaged
during operation nor do they have any modulating [slipping] capability).

Engine-Mounted Enclosed Clutches
These clutches (power takeoffs) will be covered in greater detail under the following
classifications (clutch rating definitions), as well as the specific selection considerations for the type of clutch and application.

Enclosed clutch selection for either rear or front engine mounting must be made in
accordance with the “Horsepower Absorption Capability”.

The following rating definitions are applicable to clutch arrangements offered by
Caterpillar.

Light-Duty (LD)
A light-duty clutch is used primarily to disconnect and pick up light inertia loads, but
does more work during engagement than “cut-off” duty.

A light-duty clutch should engage within two seconds, start the load less than six times per hour, and never heat the pressure plate outer surface above hand holding temperature.

Example: Disconnect clutch between engine and hydraulic torque converter with engine above low idle when engaging clutch, as in power shovel master clutch, generator, or similar drives.

Normal-Duty (ND)
A normal-duty clutch is used to start inertia loads with frequencies up to 30 engagements per hour. More important is that the clutch can start the heaviest inertia load within three seconds, and that the product of seconds of clutch slip per engagement times number of engagements per hour be under 90.

A normal-duty application may raise the outer clutch surface temperature to under
100°F (37.8°C) rise above ambient air temperature.

Example: Power takeoff starting average inertia loads where starting load is 40% of the running load.

Heavy-Duty (HD)
A heavy-duty clutch is used to start inertia loads with frequencies up to 60 engagements per hour. More important is that the clutch can start the heaviest inertia loads within four seconds, and that the product of seconds of clutch slip per engagement times number of engagements per hour be under 180.

Heavy-duty applications may raise the clutch outer surface temperature to a maximum of 150°F (65.6°C) rise above ambient air temperature.

Example: Power takeoff starting average inertia loads whose starting load is 80% of
the running load. Also, rock crusher applications where the clutch is not used to
“break loose” jammed loads.

Extra Heavy-Duty (EHD)
An extra heavy-duty clutch is used to start inertia loads requiring over four seconds to start the heaviest load, with longest slip period per engagement not exceeding 10seconds. Also, when the product of seconds of clutch slip per engagement times number of engagements per hour exceeds 180, it is beyond extra heavy-duty. Contact your Caterpillar dealer for application approval of extra heavy-duty-type service.
Example: Power takeoff starting inertia loads whose starting load approaches or
exceeds the running load.

Typical Light-Duty (LD)
Clutch Applications
A.Agitators — pure liquids.
B.Cookers —cereal.
C.Elevators, bucket — uniform loads,
all types.
D.Feeders — disc-type.
E.Kettle — brew.
F.Line shafts — light-duty.
G. Machines, general — all types with uniform loads, nonreversing.
H.Pumps — centrifugal.

Typical Normal-Duty (ND)
Clutch Applications
A.Agitators — solid or semisolids.
B.Batchers — textile.
C.Blowers and fans — centrifugal andlobe.
D.Bottling machines.
E.Compressors — all centrifugal
andlobe-type.
F.Elevators, bucket — uniformly loaded or fed.
G.Feeders — apron, belt, screw, or vane.
H.Filling machine — can type.
I.Mixers — continuous.
J.Pumps — three or more cylinders; gear- or rotary-type.
K.Conveyor — uniform load.

Typical Heavy-Duty (HD)
Clutch Applications
A.Cranes and hoist — working clutch.
B.Crushers — ore and stone.
C.Drums — braking.
D.Compressors — lobe rotary plus three or more cylinder reciprocating-type.
E.Haulers — car puller and barge-type.
F.Mills — ball-type.
G. Paper mill machinery — except calenders and driers.
H.Presses — brick and clay.
I.Pumps — one- and two-cylinder reciprocating-type.
J.Mud pumps — one- and two-cylinder reciprocating-type.

Typical Extra Heavy-Duty (EHD)
Clutch Applications
A.Compressors — one- and two-cylinder reciprocating-type.
B.Calenders and driers — paper mill.
C.Mills — hammer-type.
D.Shaker — reciprocating-type.

Once all machine parameters have been established, contact your Caterpillar dealer
for selection assistance.

Automotive-Type Clutches
Also known as diaphram or spring-loaded-type clutches, this category is generally a
light-duty classification; it is normally used in strictly mobile applications, such as on-
highway trucks or higher speed mobile machines, which utilize a multi speed transmission. The automotive-type clutch is normally foot-operated for disengagement or is engaged with the friction being generated by spring force acting on an engine-driven plate.

Although this type of clutch is not a Caterpillar price list attachment, on the
smaller engine families, there is offered a selection of flywheels to accommodate the
more common commercial models offered by a number of manufacturers.

If the machine design requires this type of clutch, the package designer and installer
should work very closely with the clutch manufacturer to ensure proper selection.

CAUTION: THIS TYPE OF CLUTCH, DUE TO ITS INHERENT TORQUE CAPACITY LIMITATIONS, SHOULD NOT BE USED WITH THE LARGER 3500 FAMILY CATERPILLAR ENGINES.

Air Clutches
Basically, engagement friction is maintained by air pressure. This feature is particularly advantageous when remote control of the engagement/disengagement functions is required.

Air clutches utilize an expanding air bladder for the clutch element. (See Figure 3).

Air clutches do not normally have side load capability, so if such capability is required, the output shaft must be supported by two support bearings. These bearings must be mounted on a common base with the engine package. Air pressure to operate the clutch is supplied by an air connection through the drilled passage in the output shaft. Clutch alignment tolerances are reduced as air pressure to the clutch increases.

When selecting an air clutch, the package designer/installer must work closely with the clutch manufacturer.

Centrifugal Clutches
The centrifugal clutch accomplishes the engagement/disengagement functions by centrifugal force which is generated by the engine operating speed. It provides a power engagement/disengagement function controlled strictly by the engine governor speed control (throttle).

Centrifugal clutches offer smooth automatic engagement of load without complicated
controls. Typically, a diesel engine with a full load operating speed of 1800 rpm will
be fitted with a centrifugal clutch which effects engagement at a speed of about
1000 engine rpm. Once engaged, most clutches of this type will remain engaged
even if the engine speed is pulled down due to load — as low as the engagement
speed (i.e., 1000 rpm) or lower (e.g., disengagement at 800 rpm). If the load is
such that engine stall speed is approached, the clutch will disengage.

As with the air-type clutches, they have limited or no side load capability and
for other than in-line drive loads, a separately supported output shaft with two support bearings must be provided and must be mounted on a common base with the engine package.

When selecting a centrifugal clutch, the package designer/installer must work closely with the clutch manufacturer.

TRANSMISSIONS
Over the years rapid technological advances have enabled numerous commercial
manufacturers to offer a broad range of transmissions with nearly unlimited features and options.

For this discussion transmissions will be divided into three broad classifications all
of which transmit power through sets of mechanical gears, either spur or helical
types, or planetary designs. Where multi-speed capability is provided, it is accomplished either mechanically or automatically (hydraulically, pneumatically, etc.)

Due to the large number of transmissions commercially available and the fact that
Caterpillar does not offer transmissions (with the exception of marine transmissions —single speed — forward/reverse functions(,the transmission discussion will be restricted to general operating principles and considerations.

When selecting a transmission, the package designer must work closely with the
transmission manufacturer.

CAUTION: REGARDLESS OF THE TYPE OR BRAND OF TRANSMISSION SELECTED,THE DESIGNER MUST ENSURE THAT IT HAS THE CORRECT HORSE-POWER, TORQUE, AND SPEED CAPABILITY TO MATCH THE DIESEL ENGINE PERFORMANCE CHARACTERISTICS.

Mechanical Transmission
The mechanical transmission provides the lowest cost method of providing multiple
output speeds when the driven equipment input speed range or torque requirements
exceed the operating capability of the diesel engine. Mechanical transmissions are usually equipped with some type of clutch assembly to facilitate not only engine starting but also to change gear ratios.

This type of transmission is applicable to both semi mobile and mobile installations
where the momentary loss of power to the driven equipment when gear changes are
effected does not pose operating problems.

Generally, the mechanical transmission is employed when the gear speed change
requirements are not a constant require- ment and the speed shifts do not have to
be executed rapidly.

Today’s modern mechanical transmission, when properly matched to the engine-driven equipment, will provide reliable trouble-free service. Frequent gear changes,
however, will accelerate clutch wear and maintenance costs.

Installation is simplified since mechanical transmissions do not normally require oil
cooling systems as do the automatic type.

Automatic, Semiautomatic, and Preselector-Type Transmissions
As the names imply, these transmission types effect the gear changes either completely automatically or as predetermined by the machine operator.
Engine power engagement/disengagement clutching is normally fully automatic and
does not require the machine operator to physically move a clutch pedal or lever. For
disengagement the operator need only move the selector lever to a neutral position.

As with the mechanical transmission, the automatic type must be carefully matched
to the engine operating horsepower, torque, and speed characteristics. However, with the automatic types, additional match consideration may be required since they normally utilize a torque converter, hydraulic coupling, or other type of non mechanical engagement device for the power engagement/disengagement function.

This is nearly always accomplished hydraulically. The automatic-type transmissions provide operator ease of machine operation, as well as a nearly constant power flow to the driven equipment during gear changes.

A number of commercial manufacturers offer a wide range of automatic-type transmission. The package designer/installer must work closely with the transmission
supplier to ensure the transmission properly matches the machine application and provides the desired operating features.

Some automatic transmission designs utilize a lockup feature. This device, in effect,
turns the transmission into a direct mechanical drive to eliminate the inherent inefficiencies of the hydraulic clutching device.

Generally, the higher cost of an automatic transmission can be justified with a machine requiring high productivity and frequent load cycle changes.

When using automatic-type transmissions, other installation considerations are required since most types require a system to cool the transmission oil. Caterpillar offers jacket water connections to supply cooling water to customer or transmission manufacturer-supplied heat exchangers.

Also offered are complete heat exchanger packages, but care must be exercised to
ensure that the Caterpillar system is capable of handling the transmission heat rejection. The cooling system capacity of the systems offered by Caterpillar can be obtained from your Caterpillar dealer and is in the Owner’s Maintenance Manual.

Speed Increasers/Reducers
These power transmission devices resemble a mechanical transmission in that power is normally transmitted through a mechanical gear set of spur or helical gears. They are used when the engine speed range is not compatible with the driven equipment input speed requirements and when the installation is best suited to an in-line drive arrangement rather than the offset belt of chain drive systems.

Speed increasers/reducers generally utilize a mechanical cutoff clutch for engine starting and are usually of a single-speed, non reversing design, although exceptions
to the above do exist. They seldom exceed two speed ratios.

Speed increasers/reducers are available for either direct engine mounting or for remote mounting. The remote-mounted type should be on a rigid common base with the engine for ease of alignment.
The package designer/installer must work closely with the commercial gear supplier to ensure proper selection and installation.

Compounds
Although infrequently found in material handling/agriculture applications, specific de-signs may require an engine compound.

Basically, a compound is an enclosed gear or chain device which permits several engines to provide input power with the power output coming from one or more shafts.

Compounds providing a single engine input and multiple outputs is most common. An example would be a hydrostatic machine where a single engine provides power to
multiple hydraulic pumps when separate pumps are used for the various functional drives of the machine.

Multiple engine compounds can be used in applications where less than the installed horsepower capability is occasionally called upon for part load operation of the driven machine.

When part load operation is adequate, the excess capability can be removed by
declutching engines, reducing overall operating costs and maintenance.

The package designer/installer must work
closely with the compound manufacturer
to ensure proper selection and installation.

Stub Shafts
Where the application permits, a stub shaft will provide a low cost, simple method of
direct power transmission.

Stub shaft drives must not be used when the starting load of the driven equipment is
sufficient to impair engine starting unless a declutching or unloading device is utilized.
Stub shafts also have limited side load capability.

Hydraulic drive
Hydraulic drive devices generally fall into two major classifications: fluid or hydraulic couplings and torque converters.

The theory involved is similar in all types of hydraulic drives although the internal
design may vary. Basically, the engine output is absorbed by a turbine-type pump.

The oil or fluid in the pump housing is accelerated outward, and the engine power is
transmitted to the outer edge of the pump as kinetic energy in the form of high velocity fluid. This energy is then transferred back towards the center of the output
shaft. This is where the differences occur between a hydraulic or fluid coupling and a
torque converter.

Fluid (Hydraulic) Couplings
In the fluid couplings, the high velocity fluid is directed into a matching turbine located very close to the turbine-type pump which is engine driven. The matching turbine absorbs the energy as the fluid is directed back toward the center of the coupling and the energy is delivered to the output shaft.

The output torque will always equal the input torque less internal friction losses which will be observed as a lower output speed (rpm)than the input speed (engine rpm).

The primary advantage of a hydraulic coupling is the total lack of a mechanical connection between the driving engine and the driven equipment.
This isolates or greatly reduces the transfer of mechanical shocks, vibration, and undesirable torsional effects between the driven load and the engine.

A hydraulic coupling will prevent engine stall under load; however, the engine can be pulled down in speed by varying degrees depending on the hydraulic coupling fluid
cooling capacity. It also permits starting high inertia-driven loads without the use of a cut-off clutch.

The main disadvantages of a hydraulic coupling are the reduced efficiency over a mechanically coupled drive and its inability to generate a torque multiplication as is
possible with a torque converter.

Normally, hydraulic couplings are best suited to applications which are constant speed applications where the slip capability is desirable to compensate for shock loads, overloads, high inertia load startups, and assist in torsional vibration reduction.
Torque Converters
As with hydraulic couplings, torque converters differ considerably in internal construction and refinement but can generally be placed in two classifications: single-stage and multistage. These differences will be expanded later in this section.

The torque converter differs from the hydraulic coupling in that one or more third
members, called stators or turbine reactors, are utilized in addition to the input pump and the output turbine. These stators or reactor members are imposed in the fluid flow path in such a manner as to produce a multiplication of the input torque to the output shaft at reduced output speeds (rpm).

The maximum torque is transmitted to the output shaft (driven equipment) at stall condition (output shaft is not rotating) when it will equal from 1.6 to more than 6.0 times the converter input torque (engine output torquevalue). When operating at full
speed, with the imposed load at a level which permits the output speed to be close to the engine speed, the torque converter acts in principle like a hydraulic coupling.

The necessity of matching a torque converter to the engine cannot be overemphasized. An improperly sized converter, one with the wrong blading or one which operates in a highly inefficient speed range, will prove unsatisfactory. An improperly matched torque converter can result in engine over- load, high inefficiency, high fuel consumption, poor engine response, and other undesirable results.

The torque converter manufacturer generally has computer programs which, when
coupled to the performance characteristics of the engine, can ensure a correct “match” for any installation/application. Most converter manufacturers have performance data on the Caterpillar Diesel Engine models or data can be obtained from your Caterpillar dealer. This data is covered in the Caterpillar Technical Information File (TIF). Performance data for nonstandard ratings is also available from your Caterpillar dealer.

Additionally, cooling of the torque converter fluid is required. Torque converter cooling must be provided for the equivalent of at least 30% of the total engine heat rejection when using a pre combustion chamber-type engine. When using a direct injection-type engine, torque converter cooling must be provided for the equivalent of at least 50% of the total engine heat rejection.
Caterpillar offers, as price list attachments, either jacket water connections for heat
exchanger-type coolers or, on the 3200,3300 and 3400 Series Engines, complete heat exchanger cooling packages. It is imperative that the cooling package be of adequate capacity. The capacity of

Most commercially available converters are also offered with attachment cooling
packages.

If the engine cooling system is used to cool the torque converter, adequate reserve
radiator capacity must be provided.

Single-Stage Torque Converters
This type of converter is normally selected for light-duty applications. It has a decreasing torque absorption curve as the output speed approaches stall condition and will not pull down the engine input speed (lug the engine).

Multistage Torque Converters
Most applications will utilize a multistage converter. They provide a broader usable
range and higher torque multiplication value than single-stage converters.

Torque converter manufacturers provide excellent manuals and assistance in the
selection of the correct converter for a specific application. Consequently, rather than elaborating on selection guidelines in this publication, it is suggested that the package designer/installer counsel with the converter manufacturer for expert advice.
In addition to offering the same benefits as a hydraulic drive, the torque converter also offers a torque multiplication benefit as well as, if properly matched, higher power transmission efficiency. The multistage converter is particularly preferred for variable output speed applications.

As standard price list attachments, Caterpillar offers flywheels to couple to most commercial torque converters and hydraulic drives.

Special Considerations
With the selection of any of the above methods of power transmission, several
general areas must also be given special consideration to ensure a successful
installation.

Side Loading
Excessive side loading is one of the most commonly encountered problems in the
transmission of engine power.

It is impossible to overemphasize the need for accurate evaluation of side load imposition on all types of power transmission devices.

For Caterpillar-supplied attachment power takeoffs, the Caterpillar Industrial Engine
Price List LEKI8162 provides complete instructions and capacity data for side load
evaluation.
For power transmission devices supplied by others, the manufacturer must be consulted for a capability analysis of his equipment.

Overhung Power Transmission Equipment
Power transmission equipment, which is directly mounted to the engine flywheel housing, must be evaluated to ensure that the overhung weight is within the tolerable limits of the engine. If not, adequate additional support must be provided to avoid
damage.

CAUTION: CERTAIN APPLICATIONS, SUCH AS AGRICULTURE MACHINES, DRILLS, OFF-HIGHWAY TRUCK, ETC., REQUIRE CONSIDERATION OF THE EFFECTS OF THE DYNAMIC BENDING
MOMENT IMPOSED DURING NORMAL MACHINE MOVEMENT OR ABRUPT
STARTING AND STOPPING.

The dynamic load limits and the maximum bending moment that can be tolerated by
the flywheel housing can be obtained from your Caterpillar dealer.

For determination of the bending moment of overhung power transmission equipment
installations, see Figure 13.

To compensate for power transmission systems which create a high bending
moment due to overhung load, a third mount is required. Proper design of the
support is essential. Forces and deflections of all components of the mounting
system must be resolved. If the third mount is in the form of a spring, with a vertical rate considerably lower than vertical rate of the rear engine support, the effect
of the mount is in a proper direction to reduce bending forces on the flywheel
housing due to downward gravity forces, but the overall effect may be minor at high
gravity force levels. The use of supports with a vertical rate higher than the engine
rear mount is not recommended since frame bending deflections can subject the
engine power transmission equipment structure to high forces. Another precaution is to design the support so that it provides as little resistance as possible to engine roll.

This also helps to isolate the engine/transmission structure from mounting frame or base deflection.

Wet Flywheel Housings
Certain types of power transmission equipment require a “wet” flywheel housing.
Wet housing equipment requires that the flywheel housing be able to accommodate a degree of flooding by the fluid medium of the power transmission equipment. The standard Caterpillar Diesel Engine does not:

A.Contain sufficient provisions for seal- ing in the area of the rear crankshaft
seal to prevent the transfer of the power transmission fluid into the engine lubricating oil reservoir (pan).

B.Have the capability of evacuating the transmission fluid from the flywheel
housing back to the transmission reservoir to prevent engine crankshaft seal flooding.
COUPLINGS
Unless a belt, chain, or universal joint-type drive is taken directly from the output shaft of the engine-driven power transmission device, the use of some type of mechanical coupling device is recommended.

The coupling must be installed between the power transmission output shaft and
the input drive shaft of the driven machine.

On close-coupled driven equipment, the use of a coupling can be avoided if two basic
criteria are met:
A.Is the torsional compatibility of the driven machine compatible with the engine to the point that lack of a coupling will not cause either engine or driven machine problems?

B.Is the package base sufficiently rigid to avoid any distortion during operation?
Does it contain sufficient alignment control features to successfully retain alignment during operation to preclude the need for the misalignment tolerance capability of a coupling?

Seldom can both of these questions be answered affirmatively.
A large number of commercial coupling designs, are available to the package
designer/installer.

CAUTION: THE COUPLING MUST BE TORSIONALLY COMPATIBLE.

Commercial couplings make use of resilient materials ranging from rubber and tough
fabrics to springs and air-filled tubes and drums in order to absorb minor mechanical misalignment and relative movement between engine and load. It is important to
have the best possible alignment and put a minimum load and reliance on the flexible
coupling. Air clutches are not flexible couplings and imposing misalignment on them
will cause damage.

Four distinct characteristics must be taken into account in the selection of a suitable
coupling:
A. Misalignment Capability
The coupling must be capable of compensating for any misalignment between the engine and equipment to prevent damage to the machine and/or diesel engine crankshaft and bearings.

If single bearing equipment is used, the coupling must be torsionally and radially rigid to transmit the load and support the weight of the driven equipment input shaft.
It must be flexible to compensate for angular misalignment
due to:
1-Thermal growth differences between the diesel engine and driven equipment.
2-Dimensional tolerances between the two units and dynamic conditions, such as torque reaction.
3-Momentary misalignment due to shock or other transient conditions.

B. Stiffness
The coupling must be of proper torsional stiffness to prevent critical orders of torsional vibration from occurring within the operating speed range. For single-bearing driven equipment, a complete torsional analysis is necessary to ensure compatibility. For two-bearing driven equipment, a simpler type of analysis is adequate. A complete torsional vibration analysis can be obtained from your Caterpillar Engine supplier, as can mass-elastic data on the diesel engine to permit
customer analysis.

C. Serviceability
When selecting a coupling, ease of installation and service is an important consideration. If spacers can be used to permit removal and installation of the coupling without disturbing the diesel engine driven machine alignment, time can be saved if service or replacement of the coupling is ever required.

When selecting a coupling, ensure that the design can withstand reasonable misalignment without materially decreasing the service life of the flexible elements.

When coupling design demands extremely close alignment, one of the major purposes for using a coupling is defeated.

D. Coupling Selection
In any installation, the coupling should be the weakest part of the entire power train; the first part to fail.

If failure does occur, the chance of damage to the diesel engine and driven machine is minimized. Safety measures must be considered to prevent major equipment damage should coupling failure occur. The use of a standard, commercially available coupling offers the benefit of parts avail- ability and reduced downtime in case of failure.

AUXILIARY DRIVES
Many applications have a requirement for auxiliary drive capability to power charging
alternators, air compressors, hydraulic steering pumps, etc.
Caterpillar offers, as price list attachments, various auxiliary drive options for all engine models. These attachments provide either mechanical gear or belt drive capability.

Gear Drives
These drives are suitable for direct mounting of air compressors and hydraulic
pumps for power assist steering, etc.

Belt Drives
Several options exist for belt driving various auxiliary attachments. Both of the following methods are available from Caterpillar:

A. Crankshaft Pulleys
Additional stack-on pulleys can be added to the front of the crankshaft.
The number of additional grooves which can be added depends on other belt-driven equipment such as cooling fans and charging alternators and the amount of total side load which will be imposed on the front of the crankshaft.

B. Gear Drive Pulleys
The gear drive auxiliary positions may be equipped with output pulleys.

Engine Alignment

amrxp2005@gmail.com (Engineer) — Sat, 21 Feb 2009 05:02:34 PST

Principles
To provide the necessary alignment between the diesel engine and all mechanically dri-ven components, an understanding of the types of misalignment and the methods of measurement is required.

Many crankshaft and bearing failures are the result of improper alignment of drive
systems at the time of initial engine instal-lation. Misalignment always results in
some type of vibration or stress loading.

CAUTION: BEFORE MAKING ANY ATTEMPTS TO MEASURE RUN OUT OR
ALIGNMENT, IT IS IMPORTANT THAT ALL SURFACES TO BE MEASURED OR MATED BE COMPLETELY CLEAN AND FREE FROM GREASE, PAINT, OXIDA-TION, OR RUST AND DIRT — ALL OF WHICH CAN CAUSE INACCURATE MEA-SUREMENTS.

Common mistakes include failure to detect “run out” of rotating assemblies and paral-lel or angular misalignment of the engineand driven machine.
The run out of a hub or flywheel can be measured by turning the part in question
while measuring from any stationary point to the surface being checked. This can be
done with a dial indicator. Note: Measure to the pilot surface being used, not to an
adjacent surface, because surfaces not used for pilots normally are not machined
as closely.

This check should be made first on the face of the wheel or hub, as illustrated in
Figure 1. Whenever making a face check, make sure the shaft end play does not
change as you rotate it. The crankshaft must be moved within the diesel engine to
remove all end play and that position must be maintained throughout the alignment
procedures.

Checking Face Run Out
While turning the wheel 360°, note any change in the dial indicator reading. Anychange is caused by face run out. Face run out may be caused by foreign
material between a crankshaft flange and flywheel, uneven torquing or from machining variations.

“Cocking” of the wheel being measured may cause indications of outside diameter
run out in addition to face run out. For this reason the face run out is checked first.
After the face run out has been eliminated, outside diameter run out can be checked.

This must also be done with a dial indicator.(See Figure 2.(

Checking Outside Diameter Run Out
While turning the hub through 360°of rotation, check for any change in indicator reading. The indicator is held stationary and, if the reading changes, the outside
diameter is off center.

After the flywheel or driving hub has been checked for run out, the same procedure
should be followed on the driven side of the coupling.

After the run out of both the driving and driven sides of the coupling have been found within limits, the engine and load alignment can be checked. There are two kinds of misalignment: parallel and angular (bore and face). (See Figure 3).

Checking Parallel Alignment
Parallel misalignment can be detected by attaching a dial indicator, as shown in Figure 4, and observing the dial indicator readings at several points around the out-
side diameter of the flywheel as the wheel holding the indicator is turned.

As a rule of thumb, the load shaft should indicate to be higher than the engine shaft
because:
A-Engine bearings have more clearance than most bearings on driven equipment.

B-The flywheel or front drive rotates in a“drooped” position below the center-line of rotation.

C-The vertical thermal growth of the engine is usually more than that of the driven equipment. Engine main bearing clearance should be considered when adjusting for parallel alignment.

Note: Both parts can be rotated together if desired. This would eliminate any out-of- roundness of the parts from showing up in the dial indicator reading. In most cases rubber driving elements must be removed or disconnected on one end during alignment since they can give false parallel readings.

Checking Angular Alignment
Angular misalignment can be determined by measuring between the two parts to be
joined. The measurement can be easily made with a feeler gauge, and it should be the same at four points around the hubs Figure 5.

If the coupling is installed, a dial indicator from one face to the other will indicate any angular misalignment. In either case, the readings will be influenced by how far from the center of rotation the measurement is made.

Note: the face and bore alignment affect each other. Thus, the face alignment should be rechecked after the bore alignment and vice versa.

After determining that the engine and load are in alignment, the crankshaft end play
should be checked to see that bolting and coupling together does not cause end thrust.

Torque Reaction
The tendency of the engine to twist in the opposite direction of shaft rotation and the
tendency of the driven machine to turn in the direction of shaft rotation is torque reaction. It naturally increases with load and may cause a torque vibration. This type of vibration will not be noticeable at idle but will be felt with load. This usually is caused by a change in alignment due to insufficient base strength allowing excessive base deflection under torque reaction load. This has the effect of introducing a side to side centerline offset which disappears when the engine is idled (unloaded)
or stopped.

Belt and Chain Drives
Belt and chain drives may also cause the engine or driven machine to shift or change
position when a heavy load is applied.

Belts and chains may also cause PTO shaft or crankshaft deflection, which can cause bearing failures and shaft bending failures. The driving sprocket or pulley must always be mounted as close to the supporting bearing as possible. Side load limits must not be exceeded. Sometimes, due to heavy side load, it is necessary to provide additional support for the driving pulley or sprocket. This can be done by providing a separate shaft which is supported by a pillow block bearing on each side of the pulley or sprocket. This shaft can then be driven by the engine or clutch through an appropriate coupling.
The size of the driving and driven sprockets or pulleys is also important. A larger pulley or sprocket will give a higher chain or belt speed. This allows more horsepower to be transmitted with less chain or belt tension.

If it is suspected that the engine or the driven machine is shifting under load, it can
be checked by measuring from a fixed point with a dial indicator while loading and unloading the engine. Torque reactive vibrations or torque reactive misalignment
will always occur under load.

Couplings
A coupling must be torsionally compatible with engine and driven load so that torsional vibration amplitudes are kept within acceptable limits. A mathematical study
called a torsional vibration analysis should be done on any combination of engine drive-line-load for which successful experience doesn’t already exist. A coupling with the wrong torsional stiffness can cause serious damage to engine or driven equipment.

All couplings have certain operating ranges of misalignment, and the manufacturers
should be contacted for this information.

Some drives, such as U-joint couplings, have different operating angle limits for different speeds.

As a general rule, the angle should be the same on each end of the shaft. (See Figure6.) The yokes must be properly aligned and sliding spline connections should move freely. If there is no angle at all, the bearings will brinell due to lack of movement.

ALIGNMENT INSTRUCTIONS
General Considerations
Alignment methods will vary depending on the coupling method selected. On Caterpillar Diesel Engines either a flexible-type or rigid-type coupling is acceptable, depending on the specific installation characteristics and the results of the Torsional Analysis.

CAUTION: IT IS IMPORTANT THAT THEPACKAGE ALIGNMENT BE CARRIED OUT AND COMPLETED WITHIN THE PERMISSIBLE TOLERANCES OF THE DRIVEN EQUIPMENT MANUFACTURER.

Alignment Instructions — Single-Bearing Driven Equipment

A. Flexible-Type Couplings — Flywheel
Housing-Mounted Driven Equipment

1-Droop
Mount a dial indicator on the engine flywheel housing. Mark the engine flywheel housing. Mark the flywheel at points A, B, C, and D in 90°increments as shown in Figure 7. The indicator tip must contact the pilot diameter of the flywheel assembly.

With the dial indicator in position (A), set the reading to zero. Place a pry bar under the flywheel assembly at position (C) and, by prying against a floor mounted support, raise the flywheel until it is stopped by the main bearings. (Caution: Do not pry against the flywheel housing.) Record the reading of the dial indicator. This is the amount of droop in the crankshaft, which results from engine bearing clearances and natural droop as a result of the overhung weight of the flywheel. The flywheel should be raised several times to get a “feel” for the bearing clearance to prevent excessive lift which means reverse bending of the crankshaft.

2-Flywheel Concentricity
Remove the pry bar and check to ensure that the dial indicator has returned to zero. If not, reset. Rotate the crankshaft, in the normal direction only, and record the Total Indicator Reading (TIR) when the flywheel positions (A), (B), (C), and (D) are at the top. (Refer to Page 58 for proper tolerances).

3-Crankshaft End Play
Ensure the crankshaft-flywheel assembly is completely to the rear- most position of the engine assembly. Reset the dial indicator to zero.

Relocate the pry bar and move crankshaft-flywheel assembly forward in the engine assembly. The dial indicator reading in this position is the crankshaft end play.

4-Flywheel Face Run Out
Set the tip of the indicator on the face of the flywheel Figure 8. Position the crankshaft to the front of its end play and zero the indicator. Shift the crankshaft to the rear of its end play, and record the TIR. With the crankshaft to the rear of its end play, zero the indicator.
Rotate the crankshaft and record the TIR when the flywheel positions (A), (B), (C), and (D) are at the top. Be sure to remove the crankshaft end play before recording these readings. Remove the flywheel housing access cover and place a pry bar between the rear face of the flywheel housing and the front face of the flywheel assembly. Move the crankshaft flywheel assembly to the rear of the engine to remove all end play.

5-Flywheel Housing Concentricity
Mount the dial indicator on the flywheel assembly with the tip located on the pilot bore of the flywheel housing and set the reading to zero.

Rotate the crankshaft in the direction of normal engine rotation and record the indicator readings at positions (A),(B), (C), and (D).

Subtract the droop dimension (Step 1) from the reading indicated at position (C) and subtract one-half the droop dimension from the reading indicated at positions (B)and (D) on the flywheel housing to determine the true concentricity.

6-Engine Mounting Face Depth
With the crankshaft-flywheel assembly moved to the frontmost position, place a straight edge across the mounting face of the flywheel housing, from position (A) to (C). With a scale measure the distance from the rear face of the flywheel housing to
the coupling mounting face of the flywheel as shown in Figure 9.

Repeat the same measurement with the straight edge located on positions (B) and (D).
Steps 1 through 6 establish the engine tolerances. The following Steps, 7 and
,8determine the driven equipment tolerances or refer to manufacturers specifications.

7-Support the driven equipment
input shaft until it is centered (all droop is removed).

8- Driven Equipment Mounting FaceDepth
With the driven equipment mounting and driving flange or face centered, as described in Step 7, and the flexible coupling attached to the input shaft, the face depth can be measured. Place a straight edge across the surface of the front face of the coupling which mates to the flywheel assembly. With a scale measure the distance from the coupling mounting face to the mounting face of the driven equipment housing as shown in Figure 10.

This dimension must equal the engine mounting face depth Step 6 less one-half of the crankshaft end play as described in Step 4. If not, it must be corrected by changing the adapting parts, or by shimming if the required correction is small. Shimming is usually the less desirable approach.

With the engine and driven equipment tolerances known, proceed to mount the driven equipment to the engine.

9-Support the driven machine on a hoist and bring it into position with the engine.

10-Align the driven equipment housing mounting flange with the flywheel housing, using locating dowels if required. Install connecting bolts with sufficient torque to compress the lock washers, but not to final torque.

11-Install the bolts which secure the coupling to the flywheel and torque as recommended.

12-Check crankshaft end play to ensure that the proper relationship exists between the engine mounting face depth Step 6 and the driven equip- ment mounting face depth Step 8.

Place a pry bar between the flywheel assembly and the flywheel housing.
The crankshaft should move both for ward and backward within the engine
and, in both positions, remain fixed when pressure on the pry bar is relaxed. Any tendency of the crankshaft to move when pry bar pressure is released indicates that the driven equipment and coupling assembly are imposing a horizontal force on the crankshaft, which will result in thrust bearing failure. If this condition exists, readjust the thickness of shims used between the driven equipment input shaft and the coupling as described in Step 8.

13-Determine quantity and thickness of shims required between the driven equipment mounting pads and the base assembly; locate the shim packs and install driven equipment mounting bolts to the base assembly.

NOTE: Always use metal shims. Tighten the bolts to one-half the torque recommendation.

14-Loosen the bolts holding the driven equipment housing to the flywheel housing until the lock washers move freely. Using a feeler gauge, check the clearance between the two housings to determine if the driven equipment is properly shimmed.

Measurement should be made in four 90°increments in the vertical and horizontal planes. If the feeler gauge indicates any area where the clearance varies by more than 0.005 in (0.13mm),readjust the driven equipment housing position by changing the shims.
There must be clearance at all points when making this check.

15-With the proper number of shims installed to align the driven equipment housing parallel to the flywheel housing, tighten the bolts securing the driven equipment housing to the flywheel housing sufficiently to compress the lock washers.

16-Torque the bolts holding the driven equipment frame to the base assembly to one-half the recommended value.

17-Repeat Step 14. If the feeler gauge measurements indicate that misalign-
ment is still present, repeat operation described in Steps 14 through 17 until proper alignment is obtained.

18-Retorque all coupling and mounting bolts to the specified torque value.

B. Flexible-Type Couplings — Remote-Mounted Driven Equipment

1-Droop
Mount a dial indicator on the engine flywheel housing. Mark the flywheel at points A, B, C, and D in 90°increments as shown in Figure 36. The indicator tip must contact the pilot diameter of the flywheel assembly.

With the dial indicator in position (A),set the reading to zero. Place a pry bar under the flywheel assembly at position (C) and, by prying against a floor mounted support, raise the flywheel until it is stopped by the main bearings. (Caution: Do not pry against the flywheel housing.) Record the reading of the dial indicator. This is the amount of droop in the crankshaft which results from engine bearing clearances and natural droop as a result of the overhung weight of the flywheel.

The flywheel should be raised several times to get a “feel” for the bearing clearance to prevent excessive lift which means reverse bending of the crankshaft.

2-Flywheel Concentricity
Remove the pry bar and check to ensure that the dial indicator has re- turned to zero; if it is not, reset. Rotate the crankshaft, in the normal direction only, and record the TIR when the flywheel positions (A), (B), (C),and (D) are at the top.

3-Crankshaft End Play
Ensure the crankshaft-flywheel assembly is completely to the rearmost position of the engine assembly. Reset the dial indicator to zero. Relocate the pry bar and move crankshaft-flywheel assembly forward in the engine assembly. The dial indicator reading in this position is the crankshaft end play.

4-Flywheel Face Run out
Set the tip of the indicator on the face of the flywheel Figure 36. Position the crankshaft to the front of its end play and zero the indicator. Shift the crankshaft to the rear of its end play and record the TIR. With the crankshaft at the rear of its end play, zero the indicator. Rotate the crankshaft and record the TIR when the flywheel positions (A), (B), (C), and (D) are at the top. Remove all end play before recording each reading. Remove the flywheel housing access cover. Then place a pry bar between the rear face of the flywheel housing and the front of the flywheel assembly.

Move the crankshaft-flywheel assembly to the rear of the engine, removing all end play.

5- Mounting
The engine and the driven equipment should be mounted so that any necessary shimming is applied to the driven equipment. The centerline of the engine crankshaft should be lower than the centerline of the driven equipment by approximately 0.0065 in (0.165mm) to allow for thermal expansion of the engine. The value 0.0065 in(0.165mm)allowed for thermal expansion is for the engine only. If it is anticipated that thermal expansion will also affect the driven equipment centerline to mounting plane distance, that value must be subtracted from the engine thermal expansion value in order to establish the total engine centerline to driven equipment centerline distance. When measuring this value, the TIR will be 0.013in plus the droop as estab lished in Step 1.

Shim packs under all equipment should be 0.200 in (5 mm) minimum thickness to provide for later correc- tions which might require the removal of shims.

6-Coupling
Attach the driven member of the coupling to the flywheel and tighten all bolts to the specified torque value.

Gear-type couplings, double sets of plate-type rubber block drives, and Cat viscous-damped couplings are the only ones that can be installed prior to making the alignment check. Most couplings are stiff enough to affect the bore alignment and give a false reading.

7-Angular Alignment
Mount a dial indicator to read between the driven equipment input flange and the flywheel face and measure angular misalignment. Adjust position of driven equipment until TIR is within 0.008 in.

8-Linear Relationship
Mount dial indicator to the driven equipment side of the flexible coupling and indicate on the outside diameter of the flywheel side of the coupling. Zero the indicator at 12 o’clock and rotate the engine in its normal direction of rotation and check the total indicator reading at every 90°. Subtract the full“droop” from the bottom reading to give the corrected alignment reading.

The value of the top-to-bottom reading should be 0.008 in (0.20 mm) or less
under operating temperature conditions, with the engine indicating low.

Adjust all shims under the feet of the driven equipment the same amount
to obtain this limit.

The final value of the top-to-bottom alignment should include a factor for
vertical thermal growth.

Subtract one-half the “droop” from the 3 o’clock and 9 o’clock reading. This
should be 0.008 in (0.20 mm) or less.

Shift the driven equipment on the mounts until this limit is obtained.

Note: the sum of the side “raw” reading should equal the bottom reading within
0.002 in (0.051 mm). Otherwise the mounting of the dial indicator is too weak to support the indicator weight.

9-The combined difference or readings
at points B and D should not exceed a total of 0.008 in (0.20 mm). (SeeFigure 12).

10-Crankshaft End Play
The crankshaft end play must be rechecked to ensure that the driven equipment is not positioned in a manner which imposes a preload on the crankshaft thrust washers. (Refer to Step 4.) Place a pry bar between the flywheel assembly and the flywheel housing. The crankshaft should move both forward and backward within the engine and, in both positions, remain fixed when pressure on the pry bar is relaxed. Any tendency of the crankshaft to move when pry bar pressure is released indicates that the driven equipment assembly must be moved rearward on the base assembly or, if
used, the number of shims between the input flange and the flexible coupling must be reduced.

Tolerances and Torque Values
Permissible alignment tolerances and torque values for Caterpillar standard mounting hardware are available from your Caterpillar.

CAUTION: DURING OPERATION, SHOULD A CHANGE IN THE VIBRATION
OR SOUND LEVEL OCCUR, ALIGNMENT SHOULD BE RECONFIRMED. THIS IS PARTICULARLY TRUE FOR SEMIMOBILE INSTALLATIONS AND ON ANY FIXED INSTALLATIONS WHICH ARE SUBJECT TO INFREQUENT RELOCATION. ALIGNMENT SHOULD ALSO BE CHECKED ON A PERIODIC BASIS OR AT TIME OF MOVEMENT IF INSTAL- LATION IS ON A SUBBASE OR SKID- TYPE BASE.

Angular Contact Thrust Ball Bearing

amrxp2005@gmail.com (Engineer) — Sun, 25 Jan 2009 05:17:15 PST

The angular contact thrust ball bearings shown here were originally designed to support the rotary tables of drilling rigs but are also suitable for other applications where high load carrying capacity, high axial stiffness and low friction torque are important. In contrast to conventional thrust ball bearings, angular contact thrust ball bearings can accommodate radial loads in addition to axial loads and are able to operate at high speeds.

Single direction angular contact thrust ball bearings
Single direction angular contact ball bearings (fig 1) are able to take up axial loads acting in one direction. They are of separable design, i.e. the washers and ball and cage assembly can be mounted individually. In the main application area for these bearings – rotary tables – two bearings are always adjusted against each other. The second bearing has to carry the weight of the drill when it is stationary and has a higher load load carrying capacity than the first bearing.

fig 1

Double direction angular contact thrust ball bearings
In bearings of the double direction design (fig 2), the upper bearing with the higher load carrying capacity and the smaller bearing which locates in the opposite direction are combined togther to form a unit. These bearings have low height and can accommodate axial loads acting in both directions as well as moment loads. A single bearing can therefore be used to support rotating machine components in relation to stationary components. These double direction bearings are also of separable design.

fig 2

Other angular contact thrust ball bearings
In addition to the angular contact thrust ball bearings described here, SKF also manufactures single direction bearings – screw support bearings (fig 3)

– and double direction precision bearings for the work spindles of machine tools (fig 4

and fig 5).

These machine tool bearings are avialable with a contact angle of 60° for high stiffness, or where high speed capability is required, with a contact angle of 40° or 30°. Information about these precision bearings will be supplied on request.

Misalignment
Angular contact thrust ball bearings cannot tolerate any misalignment of the shaft with respect to the housing or any angular misalignment of the support surfaces on the shaft and in the housing.

Cages
Angular contact thrust ball bearings are produced either with a full complement of balls (without cage) or are fitted with machined brass cages (fig 6) which are centred on the shaft washer or the housing washer

fig 6

Minimum load
When angular contact thrust ball bearings operate at high speeds, the centrifugal forces and gyratory moments acting on the balls may cause them to slide on the raceways. To prevent the damaging consequences of this, for example smearing, the bearing must always be subjected to a given axial load.
The requisite minimum axial load to be applied to single direction angular contact thrust ball bearings can be obtained from

where
Fam =minimum axial load, kN
A= minimum load factor (see product tables)
n = rotational speed, r/min
When starting up at low temperatures or when the lubricant is highly viscous, even greater loads may be required. The weights of the components supported by the bearing, together with the external forces, generally exceed the requisite minimum load. The double direction angular contact ball bearings are preloaded by adjusting the two shaft washers against each other.

Equivalent dynamic bearing load
For single direction angular contact thrust ball bearings
P = Fa when Fa/Fr ≤ e
P = Fa + XFr when Fa/Fr > e
The values of factors e and X will be found in the product tables.

Equivalent static bearing load
For single direction angular contact thrust ball bearings
P0 = Fa + X0Fr
The values of factor X0 will be found in the product tables.

Needle Roller Bearing

amrxp2005@gmail.com (Engineer) — Sun, 25 Jan 2009 02:27:46 PST

Needle roller bearings are roller bearings with cylindrical rollers which are thin and long in relation to their diameter. ISO uses the definition that the roller length is 2,5 times the roller diameter or more. They are referred to as needle rollers. In spite of their low cross section the bearings have a high load carrying capacity and are thus extremely suitable for bearing arrangements where radial space is limited.
The SKF range of needle roller bearings is extensive and, in addition to customized designs, it comprises the following types:
– Needle roller and cage assemblies

– Drawn cup needle roller bearings
with open ends

and with closed end

– Needle roller bearings
with and without flanges

with

and without inner ring

– Sealed needle roller bearings

with

and without inner ring
– Alignment needle roller bearings

with

and without inner ring

– Combined needle roller bearings
*Needle roller/ball bearings
*Needle roller/thrust ball bearings
*Needle roller/cylindrical roller thrust bearings

Combined needle roller bearings consist of a radial needle roller bearing combined with a thrust bearing and are consequently able to take up both radial and axial loads. Combined needle roller bearings provide the means to produce locating bearing arrangements in a minimum of radial space. They are particularly useful where the axial loads are too heavy, speeds too high, or lubrication inadequate for simple thrust washers to be suitable, and other types of locating bearing take up too much room.
Combined needle roller bearings are available in the following designs:
– needle roller/angular contact ball bearings

– needle roller/thrust ball bearings

– needle roller/cylindrical roller thrust bearings

–Needle roller thrust bearings Needle roller thrust bearings can support heavy axial loads, are insensitive to shock loads and provide stiff bearing arrangements which require a minimum of axial space. They are single direction bearings and can only accommodate axial loads acting in one direction. Particularly compact bearing arrangements can be made, taking up no more space than a conventional thrust washer, if the faces of adjacent machine components can serve as raceways for a needle roller and cage thrust assembly. For applications where adjacent components cannot serve as raceways, the assemblies can also be combined with washers of various designs.
Because of all the possible combinations, all bearing components must be ordered separately.
–Inner rings for needle roller bearings

Taper Roller Bearing

amrxp2005@gmail.com (Engineer) — Sun, 25 Jan 2009 02:03:37 PST

Taper roller bearings are produced in many designs and sizes to match their many uses. These can be grouped as follows:
–single row taper roller bearings (fig 1)

–paired single row taper roller bearings (fig 2)

–double row taper roller bearings (fig 3),

and
–four-row taper roller bearings (fig 4).

These are described in four separate sections under the appropriate headings.
There are also sealed, greased and preadjusted units based on taper roller bearings, such as
– hub bearing units for passenger cars (fig 5)

– hub bearing units for trucks (fig 6),

and
– tapered bearing units (fig 7) for railbound vehicles.

Design featuresTaper roller bearings have tapered inner and outer ring raceways between which tapered rollers are arranged. The projection lines of all the tapered surfaces meet at a common point on the bearing axis. Their design makes taper roller bearings particularly suitable for the accommodation of combined (radial and axial) loads. The axial load carrying capacity of the bearings is largely determined by the contact angle α (fig 8);

the larger α, the higher the axial load carrying capacity (fig 9)

An indication of the angle size is given by the calculation factor e; the larger the value of e, the larger the contact angle and the greater the suitability of the bearing for carrying axial loads.
Taper roller bearings are generally separable, i.e. the cone, consisting of the inner ring with roller and cage assembly, can be mounted separately from the cup (outer ring).
SKF taper roller bearings have the logarithmic contact profile that provides for optimum stress distribution over the roller/raceway contacts. The special design of the sliding surfaces of the guide flange and large roller ends considerably promotes lubricant film formation in the roller end/flange contacts. The resulting benefits include increased operational reliability and reduced sensitivity to misalignment.

Demineralizers

amrxp2005@gmail.com (Engineer) — Mon, 29 Dec 2008 03:48:45 PST

Purpose of Demineralizers
Dissolved impurities in power plant fluid systems generate corrosion problems and decrease efficiency due to fouled heat transfer surfaces. Demineralization of the water is one of the most practical and common methods available to remove these dissolved impurities.

In the plant, demineralizers (also called ion-exchangers) are used to hold ion exchange resins and transport water through them. Ion exchangers are generally classified into two groups: singlebed ion exchangers and mixed-bed ion exchangers.

Demineralizers
A demineralizer is basically a cylindrical tank with connections at the top for water inlet and resin addition, and connections at the bottom for the water outlet. The resin can usually be changed through a connection at the bottom of the tank. The resin beads are kept in the demineralizer by upper and lower retention elements, which are strainers with a mesh size smaller then the resin beads. The water to be purified enters the top at a set flow rate and flows down through the resin beads, where the flow path causes a physical filter effect as well as a chemical ion exchange.

Single-Bed Demineralizers
A single-bed demineralizer contains either cation or anion resin beads. In most cases, there are two, single-bed ion exchangers in series; the first is a cation bed and the second is an anion bed. Impurities in plant water are replaced with hydrogen ions in the cation bed and hydroxyl ions in the anion bed. The hydrogen ions and the hydroxyl ions then combine to form pure water.
The Chemistry Handbook, Module 4, Principles of Water Treatment, addresses the chemistry of demineralizers in more detail.

Figure 13 illustrates a single-bed demineralizer. When in use, water flows in through the inlet to a distributor at the top of the tank. The water flows down through the resin bed and exits out through the outlet. A support screen at the bottom prevents the resin from being forced out of the demineralizer tank.

Figure 13 Single-Bed Demineralizer

Single-Bed Regeneration
The regeneration of a single-bed ion exchanger is a three-step process. The first step is a backwash, in which water is pumped into the bottom of the ion exchanger and up through the resin. This fluffs the resin and washes out any entrained particles. The backwash water goes out through the normal inlet distributor piping at the top of the tank, but the valves are set to direct the stream to a drain so that the backwashed particles can be pumped to a container for waste disposal.

The second step is the actual regeneration step, which uses an acid solution for cation units and caustic solution for anion units. The concentrated acid or caustic is diluted to approximately 10% with water by opening the dilution water valve, and is then introduced through a distribution system immediately above the resin bed. The regenerating solution flows through the resin and out the bottom of the tank to the waste drain.

The final step is a rinsing process, which removes any excess regenerating solution. Water is pumped into the top of the tank, flows down through the resin bed and out at the bottom drain.

To return the ion exchanger to service, the drain valve is closed, the outlet valve is opened, and the ion exchanger is ready for service.

Single-bed demineralizers are usually regenerated "in place." The resins are not pumped out to another location for regeneration. The regeneration process is the same for cation beds and for anion beds; only the regenerating solution is different. It is important to realize that if the ion exchanger has been exposed to radioactive materials, the backwash, regeneration, and rinse solutions may be highly radioactive and must be treated as a radioactive waste.

Mixed-Bed Demineralizer
A mixed-bed demineralizer is a demineralizer in which the cation and anion resin beads are mixed together. In effect, it is equivalent to a number of two-step demineralizers in series. In a mixed-bed demineralizer, more impurities are replaced by hydrogen and hydroxyl ions, and the water that is produced is extremely pure. The conductivity of this water can often be less than 0.06 micromhos per centimeter.

Mixed-Bed Regeneration
The mixed-bed demineralizer shown in Figure 14 is designed to be regenerated in place, but the process is more complicated than the regeneration of a single-bed ion exchanger. The steps in the regeneration are shown in Figure 14.

Figure 14a shows the mixed-bed ion exchanger in the operating, or on-line mode. Water enters through a distribution header at the top and exits through the line at the bottom of the vessel. Regeneration causes the effluent water to increase in electrical conductivity.

The first regeneration step is backwash, as shown in Figure 14b. As in a single-bed unit,backwash water enters the vessel at the bottom and exits through the top to a drain. In addition to washing out entrained particles, the backwash water in a mixed-bed unit must also separate the resins into cation and anion beds. The anion resin has a lower specific gravity than the cation resin; therefore, as the water flows through the bed, the lighter anion resin beads float upward to the top. Thus, the mixed-bed becomes a split bed. The separation line between the anion bed at the top and the cation bed at the bottom is called the resin interface. Some resins can be separated only when they are in the depleted state; other resins separate in either the depleted form or the regenerated form.

The actual regeneration step is shown in Figure 14c. Dilution water is mixed with caustic solution and introduced at the top of the vessel, just above the anion bed. At the same time, dilution water is mixed with acid and introduced at the bottom of the vessel, below the cation bed. The flow rate of the caustic solution down to the resin interface is the same as the flow rate of the acid solution up to the resin interface. Both solutions are removed at the interface and dumped to a drain.

Figure 14 Regeneration of a Mixed-Bed Demineralizer

During the regeneration step, it is important to maintain the cation and anion resins at their proper volume. If this is not done, the resin interface will not occur at the proper place in the vessel, and some resin will be exposed to the wrong regenerating solution. It is also important to realize that if the ion exchanger has been involved with radioactive materials, both the backwash and the regenerating solutions may be highly radioactive and must be treated as liquid radioactive waste.

The next step is the slow rinse step, shown in Figure 14d, in which the flow of dilution water is continued, but the caustic and acid supplies are cut off. During this two-direction rinse, the last of the regenerating solutions are flushed out of the two beds and into the interface drain.

Rinsing from two directions at equal flow rates keeps the caustic solution from flowing down into the cation resin and depleting it.

In the vent and partial drain step, illustrated in Figure 14e, the drain valve is opened, and some of the water is drained out of the vessel so that there will be space for the air that is needed to re-mix the resins. In the air mix step, (Figure 14f) air is usually supplied by a blower, which forces air in through the line entering the bottom of the ion exchanger. The air mixes the resin beads and then leaves through the vent in the top of the vessel. When the resin is mixed, it is dropped into position by slowly draining the water out of the interface drain while the air mix continues.

In the final rinse step, shown in Figure 14g, the air is turned off and the vessel is refilled with water that is pumped in through the top. The resin is rinsed by running water through the vessel from top to bottom and out the drain, until a low conductivity reading indicates that the ion exchanger is ready to return to service.

External Regeneration
Some mixed-bed demineralizers are designed to be regenerated externally, with the resins being removed from the vessel, regenerated, and then replaced. With this type of demineralizer, the first step is to sluice the mixed bed with water (sometimes assisted by air pressure) to a cation tank in a regeneration facility. The resins are backwashed in this tank to remove suspended solids and to separate the resins. The anion resins are then sluiced to an anion tank. The two batches of separated resins are regenerated by the same techniques used for single-bed ion exchangers. They are then sluiced into a holding tank where air is used to remix them. The mixed, regenerated, resins are then sluiced back to the demineralizer.

External regeneration is typically used for groups of condensate demineralizers. Having one central regeneration facility reduces the complexity and cost of installing several demineralizers.

External regeneration also allows keeping a spare bed of resins in a holding tank. Then, when a demineralizer needs to be regenerated, it is out of service only for the time required to sluice out the depleted bed and sluice a fresh bed in from the holding tank. A central regeneration facility may also include an ultrasonic cleaner that can remove the tightly adherent coating of dirt or iron oxide that often forms on resin beads. This ultrasonic cleaning reduces the need for chemical regeneration.

Demineralizers Summary

Demineralization of water is one of the most practical and common methods used to remove dissolved contaminates. Dissolved impurities in power plant fluid systems can generate corrosion problems and decrease efficiency due to fouled heat transfer surfaces. Demineralizers (also called ion-exchangers) are used to hold ion exchange resins and transport water through them. Ion exchangers are generally classified
into two groups: single-bed ion exchangers and mixed-bed ion exchangers.

A demineralizer is basically a cylindrical tank with connections at the top for water inlet and resin addition, and connections at the bottom for the water outlet. The resin can usually be changed out through a connection at the bottom of the tank. The resin beads are kept in the demineralizer by upper and lower retention elements, which are strainers with a mesh size smaller then the resin beads.

The water to be purified enters the top at a set flow rate, flows down through the resin beads where the flow path causes a physical filter effect as well as a chemical ion exchange. The chemistry of the resin exchange is explained in detail in the Chemistry Fundamentals Handbook.

There are two types of demineralizers, single-bed and mixed-bed.
Single-bed demineralizers have resin of either cation or anion exchange sites. Mixed-bed demineralizers contain both anion and cation resin.

All demineralizers will eventually be exhausted from use. To regenerate the resin and increase the demineralizer's efficiency, the demineralizers are regenerated. The regeneration process is slightly different for a mixed-bed demineralizer compared to the single-bed demineralizer. Both methods were explained in this chapter.

Boilers

amrxp2005@gmail.com (Engineer) — Mon, 29 Dec 2008 03:21:48 PST

Introduction
The primary function of a boiler is to produce steam at a given pressure and temperature. To accomplish this, the boiler serves as a furnace where air is mixed with fuel in a controlled combustion process to release large quantities of heat. The pressure-tight construction of a boiler provides a means to absorb the heat from the combustion and transfer this heat to raise water to a temperature such that the steam produced is of sufficient temperature and quality (moisture content) for steam loads.

Boilers
Two distinct heat sources used for boilers are electric probes and burned fuel (oil, coal, etc.)This chapter will use fuel boilers to illustrate the typical design of boilers. Refer to Figure 9 during the following discussion.

The boiler has an enclosed space where the fuel combustion takes place, usually referred to as the furnace or combustion chamber. Air is supplied to combine with the fuel, resulting in combustion. The heat of combustion is absorbed by the water in the risers or circulating tubes. The density difference between hot and cold water is the driving force to circulate the water back to the steam drum. Eventually the water will absorb sufficient heat to produce steam.

Steam leaves the steam drum via a baffle, which causes any water droplets being carried by the steam to drop out and drain back to the steam drum. If superheated steam is required, the steam may then travel through a superheater. The hot combustion gasses from the furnace will heat the steam through the superheater's thin tube walls. The steam then goes to the steam supply system and the various steam loads.

Some boilers have economizers to improve cycle efficiency by preheating inlet feedwater to the boiler. The economizer uses heat from the boiler exhaust gasses to raise the temperature of the inlet feedwater.

Figure 9 Typical Fuel Boiler

Fuel Boiler Components
Figure 9 illustrates a typical fuel boiler. Some of the components are explained below.

Steam drum - The steam drum separates the steam from the heated water. The
water droplets fall to the bottom of the tank to be cycled again, and the steam leaves the drum and enters the steam system. Feedwater enters at the bottom of the drum to start the heating cycle.

Downcomers - Downcomers are the pipes in which the water from the steam drum travels in order to reach the bottom of the boiler where the water can enter the distribution headers.

Distribution headers - The distribution headers are large pipe headers that carry the water from the downcomers to the risers.

Risers - The piping or tubes that form the combustion chamber enclosure are called risers. Water and steam run through these to be heated. The term risers refers to the fact that the water flow direction is from the bottom to the top of the boiler. From the
risers, the water and steam enter the steam drum and the cycle starts again.

Combustion chamber - Located at the bottom of a boiler, the combustion chamber is where the air and fuel mix and burn. It is lined with the risers.

Boilers Summary

*Boilers are vessels that allow water in contained piping to be heated to steam by a heat source internal to the vessel. The water is heated to the boiling point. The resulting steam separates, and the water is heated again. Some boilers use the heat from combustion off-gasses to further heat the steam (superheat) and/or to preheat the feedwater.

*The following components were discussed:
The steam drum is where the steam is separated from the heated water.

Downcomers are the pipes in which the water from the steam drum travels to reach the bottom of the boiler.

Distribution headers are large pipe headers that carry the water from th downcomers to the risers.

Risers are the piping or tubes that form the combustion chamber enclosure. Water and steam run through the risers to be heated.

The combustion chamber is located at the bottom of the boiler and is where the air and fuel mix and burn.

CLASSIFICATION OF ROLLING-ELEMENT BEARINGS

amrxp2005@gmail.com (Engineer) — Mon, 29 Dec 2008 00:37:50 PST

CLASSIFICATION OF ROLLING-ELEMENT BEARINGS

Ball bearings can operate at higher speed in comparison to roller bearings because they have lower friction. In particular, the balls have less viscous resistance when rolling through oil or grease. However, ball bearings have lower load capacity compared with roller bearings because of the high contact pressure of point contact. There are about 50 types of ball bearings listed in manufacturer catalogues. Each one has been designed for specific applications and has its unique characteristics. The following is a description of the most common types.

12.2.1 Ball Bearings

12.2.1.1 Deep-Groove Ball Bearing
The deep-groove ball bearing (Fig. 12-2) is the most common type, since it can be used for relatively high radial loads. Deep-groove radial ball bearings are the most widely used bearings in industry, and their market share is about 80% of industrial rolling-element bearings. Owing to the deep groove in the raceways, they can support considerable thrust loads (in the axial direction of the shaft) in

addition to radial loads. A deep-groove bearing can support a thrust load of about
70% of its radial load. The radial and axial load capacity increases with the bearing size and number of balls.
For maximum load capacity, a filling-notch type of bearing can be used that has a larger number of balls than the standard bearing. In this design, there is a notch on one shoulder of the race. The circular notch makes it possible to insert more balls into the deep groove between the two races. The maximum number of balls can be inserted if the outer ring is split. However, in that case, external means must be provided to hold and tighten the two ring halves together.

12.2.1.2 Self-Aligning Ball Bearings
It is very important to compensate for angular machining and assembly errors between the centerlines of the bearing and the shaft. The elastic deflection of the shaft is an additional cause of misalignment. In the case of a regular deep-groove ball bearing, the misalignment causes a bending moment in the bearing and additional severe contact stresses between the balls and races. However, in the self-aligning bearing (Fig. 12-3), the spherical shape of the outer race allows an additional angular degree of freedom (similar to that of a universal joint) that prevents the transfer of any bending moment to the bearing and prevents any additional contact stresses.

Self-aligning ball bearings have two rows of balls, and the outer ring has a common spherical raceway that allows for the self-aligning characteristic. The

inner ring is designed with two restraining ribs (also known as lips), one at each
side of the roller element, for accurately locating the rolling elements’ path on the
inner raceway. But the outside ring has no ribs, in order to allow for selfalignment.
A wide spherical outer race allows for a higher degree of selfalignment.

Self-aligning ball bearings are widely used in applications where misalignment
is expected due to the bending of the shaft, errors in the manufacture of the shaft, or mounting errors. The design engineer must keep in mind that there are always tolerances due to manufacturing errors. Self-aligning bearings can be applied for radial loads combined with moderate thrust loads. The feature that self-aligning bearings do not exert any bending moment on the shaft is particularly important in applications that require high precision (low radial run-out) at high speeds, because shaft bending causes imbalance and vibrations.

The concept of self-alignment is useful in all types of bearings, including sleeve
bearings.

12.2.1.3 Double-Row Deep-Groove Ball Bearing
This bearing type (Fig. 12-4) is used for relatively high radial loads. It is more sensitive to misalignment errors than the single row and should be used only for
applications where minimal misalignment is expected. Otherwise, a self-alignment
bearing should be selected.

The design of double-row ball bearings is similar to that of single-row ball
bearings. Since double-row ball bearings are wider and have two rows, they can

carry higher radial loads. Unlike the deep-groove bearing, designs of split rings (for the maximum number of balls) are not used, and each ring is made from one piece. However, double-row bearings include groups with larger diameters and a larger number of balls to further improve the load capacity.

12.2.1.4 Angular Contact Ball Bearing
This bearing type (Fig. 12-5) is used to support radial and thrust loads. Contact angles of up to 40 (from the radial direction) are available from some bearing manufacturers, but 15 and 25 are the more standard contact angles. The contact
angle determines the ratio of the thrust to radial load.

Angular contact bearings are widely used for adjustable arrangements, where they are mounted in pairs against each other and preloaded. In this way, clearances in the bearings are eliminated or even preload is introduced in the rolling contacts. This is often done to stiffen the bearings for a rigid support of the shaft. This is important for reducing the amplitude of shaft vibrations under oscillating forces. This type of design has significant advantages whenever precision is required (e.g., in machine tools), and it reduces vibrations due to imbalance. This is particularly important in high-speed applications. An adjustable arrangement is also possible in tapered bearings; however, angular contact ball bearings have lower friction than do tapered bearings. However, the friction of angular contact ball bearings is somewhat higher than that of radial ball bearings. Angular contact ball bearings are the preferred choice in many important applications, such as high-speed turbines, including jet engines.

Single-row angular contact ball bearings can carry considerable radial loads
combined with thrust loads in one direction. Prefabricated mountings of two or more single-row angular contact ball bearings are widely used for two-directional thrust loads. Two bearings in series can be used for heavy unidirectional thrust loads, where two single-row angular contact ball bearings share the thrust load.

Precise axial internal clearance and high-quality surface finish are required to
secure load sharing of the two bearings in series. The bearing arrangement of two or more angular contact bearings facing the same direction is referred to as tandem arrangement. The bearings are mounted adjacent to each other to increase the thrust load carrying capacity.

12.2.2 Roller Bearings
Roller bearings have a theoretical line contact between the unloaded cylindrical rollers and races. This is in comparison to ball bearings, which have only a theoretical point contact with the raceways. Under load, there is elastic deformation, and line contact results in a larger contact area than that of a point contact in ball bearings. Therefore, roller bearings can support higher radial loads. At the same time, the friction force and friction-energy losses are higher for a line contact; therefore, roller bearings are usually not used for high-speed applications.

Roller bearings can be classified into four categories: cylindrical roller
bearings, tapered roller bearings, needle roller bearings and spherical roller
bearings.

12.2.2.1 Cylindrical Roller Bearings
The cylindrical roller bearing (Fig. 12-6) is used in applications where high radial load is present without any thrust load. Various types of cylindrical roller bearings are manufactured and applied in machinery. In certain applications where diameter space is limited, these bearings are mounted directly on the shaft, which serves as the inner race. For direct mounting, the shaft must be hardened to high Rockwell hardness, similar to that of the bearing race. For direct mounting, the radial load must be high in order to prevent slipping between the rollers and the shaft during the start-up. It is important to keep in mind that cylindrical roller bearings cannot support considerable thrust loads. Thus, for applications where both radial and thrust loading are present, it is preferable to use ball bearings.

12.2.2.2 Tapered Roller Bearing
The tapered roller bearing is used in applications where a high thrust load is present that can be combined with a radial load. The bearing is shown in Fig. 12-7. The races of inner and outer rings have a conical shape, and the rolling elements between them have a conical shape as well. In order to have a rolling motion, the contact lines formed by each of the various tapered roller elements

and the two races must intersect at a common point on the bearing axis. This intersection point is referred to as an apex point. The apex point is closer to the bearing when the cone angle is steeper. A steeper cone angle can support a higher
thrust load relative to a radial load.

The inner ring is referred to as cone, while the outer ring is referred to as cup. The cone is designed with two retaining ribs (also known as lips) to confine

the tapered rollers as shown in Fig. 12-7. The ribs also align the rollers between the races. In addition, the larger rib has an important role in supporting the axial load. A cage holds the cone and rollers together as one unit, but the cup (outer ring) can be pulled apart.

A single-row tapered roller bearing can support a thrust load in only one
direction. Two tapered roller bearings are usually mounted in opposition, to allow for thrust support in both directions (in a similar way to opposing angular contact ball bearings). Moreover, double or four-row tapered roller bearings are applied in certain applications to support a high bidirectional thrust load as well as radial load.

The reaction force on the cup acts in the direction normal to the line of contact of the rolling elements with the cup race (normal to the cup surface). This force can be divided into axial and radial load components. The intersection of the resultant reaction force (which is normal to the cup angle) with the bearing centerline is referred to as the effective center. The location of the effective center is useful in bearing load calculations.

For example, when a radial load is applied on the bearing, this produces both radial and thrust reactions. The thrust force component, which acts in the direction of the shaft centerline, can separate the cone from the cup by sliding the shaft in the axial direction through the cone or by the cup’s sliding axially in its seat. To prevent such undesired axial motion, a single-row tapered bearing should be mounted with another tapered bearing in the opposite direction. This arrangement is also very important for adjusting the clearance.

One major advantage of the tapered roller bearing is that it can be applied in
adjustable arrangement where two tapered roller bearings are mounted in opposite directions (in a similar way to the adjustable arrangement of the angular contact ball bearing that was discussed earlier). This arrangement allows one to eliminate undesired clearance and to provide a preload (interference or negative clearance). Bearing preload increases the bearing stiffness, resulting in reduced vibrations as well as a lower level of run-out errors in precision machining.
However, the disadvantage of bearing preloading is additional contact stresses
and higher friction. Preload results in lowering the speed limit because the higher friction causes overheating at high speeds.

The adjustment of bearing clearance can be done during assembly and even during steady operation of the machine. The advantage of adjustment during operation is the precise elimination of the clearance after the thermal expansion of
the shaft.

12.2.2.3 Multirow Tapered Roller Bearings
The multirow tapered roller bearing (Fig. 12-8) is manufactured with a predetermined
adjustment that enables assembly into a machine without any further adjustment. The multirow arrangement includes spacers and is referred to as a

spacer assembly. The spacer is matched with a specific bearing assembly during
manufacturing. It is important to note that components of these assemblies are not
interchangeable. Other types, without spacers, are manufactured with predetermined
internal adjustment, and their components are also not interchangeable.

12.2.2.4 Needle Roller Bearing
These bearings (Fig. 12-9) are similar to cylindrical roller bearings, in the sense that they support high radial load. This type of bearing has a needle like appearance because of its higher length-to-diameter ratio.
The objective of a needle roller bearing is to save space. This is advantageous in applications where bearing space is limited. Furthermore, in certain applications needle roller bearings can also be mounted directly on the

shaft. For a direct mounting, the shaft must be properly hardened to a similar
hardness of a bearing ring.

Two types of needle roller bearings are available. The first type, referred to as full complement, does not include a cage; the second type has a cage to separate the needle rollers in order to prevent them from sliding against each other. The full-complement bearing has more rollers and can support higher radial load. The second type has a lower number of rollers because it has a cage to separate the needle rollers to prevent them from rubbing against each other. The speed of a full-complement bearing is limited because it has higher friction between the rollers. A full-complement needle bearing may comprise a maximum number of needle rollers placed between a hardened shaft and a housing bore. An outer ring may not be required in certain situations, resulting in further saving of space.

12.2.2.5 Self-Aligning Spherical Roller Bearing
This bearing has barrel-shaped rollers (Fig. 12-10). It is designed for applications that involve misalignments due to shaft bending under heavy loads and due to manufacturing tolerances or assembly errors (in a similar way to the self-aligning
ball bearing). The advantage of the spherical roller bearing is its higher load capacity in comparison to that of a self-aligning ball bearing, but it has higher frictional losses.

Spherical roller bearings are available as single-row, double-row, and thrust
types. The single-row thrust spherical roller bearing is designed to support only thrust load, and it is not recommended where radial loads are present. Double-row

spherical roller bearings are commonly used when radial as well as thrust loads
are present.

The double-row spherical roller bearing has the highest load capacity of all rolling bearings. This is due to the relatively large radius of contact of the rolling element. It can resist impact and other dynamic forces. It is used in heavy-duty
applications such as ship shafts, rolling mills, and stone crushers.

Types of Control Valves (Part 2)

amrxp2005@gmail.com (Engineer) — Sat, 06 Dec 2008 02:22:47 PST

Shuttle and fast exhaust valves
A shuttle valve, also known as a double check valve, allows pressure in a line to be obtained from alternative sources. It is primarily a pneumatic device and is rarely found in hydraulic circuits.

Construction is very simple and consists of a ball inside a cylinder, as shown in Figure 4.25a. If pressure is applied to port X, the ball is blown to the fight blocking port Y and linking ports X and A.

Similarly, pressure to port Y alone connects ports Y and A and blocks port X. The symbol of a shuttle valve is given in Figure 4.25b.

A typical application is given in Figure 4.25c, where a spring return cylinder is operated from either of two manual stations.

Isolation between the two stations is provided by the shuttle valve. Note a simple T-connection cannot be used as each valve has its A port vented to the exhaust port.

A fast exhaust valve (Figure 4.26) is used to vent cylinders quickly. It is primarily used with spring return (single-acting) pneumatic cylinders. The device shown in Figure 4.26a consists of a movable disc which allows port A to be connected to

pressure port P or large exhaust port R. It acts like, and has the same symbol as, a shuttle valve. A typical application is shown in Figure 4.26b.

Fast exhaust valves are usually mounted local to, or directly onto, cylinders and speed up response by avoiding any delay from return pipes and control valves. They also permit simpler control valves to be used.

Sequence valves
The sequence valve is a close relative of the pressure relief valve and is used where a set of operations are to be controlled in a pressure related sequence. Figure 4.27 shows a typical example where a workpiece is pushed into position by cylinder 1 and clamped by cylinder 2.

Sequence valve V 2 is connected to the extend line of cylinder 1. When this cylinder is moving the workpiece, the line pressure is low, but rises once the workpiece hits the end stop. The sequence valve opens once its inlet pressure rises above a preset level.

Cylinder 2 then operates to clamp the workpiece. A check valve across V 2 allows both cylinders to retract together.

Time delay valves
Pneumatic time delay valves (Figure 4.28) are used to delay operations where time-based sequences are required. Figure 4.28a shows construction of a typical valve. This is similar in construction to a 3/2 way pilot-operated valve, but the space above the main valve is comparatively large and pilot air is only allowed in via a flow reducing needle valve. There is thus a time delay between application of pilot pressure to port Z and the valve operation, as shown by the timing diagram in Figure 4.28b. The time delay is adjusted by the needle valve setting.

The built-in check valve causes the reservoir space above the valve to vent quickly when pressure at Z is removed to give no delay off.

The valve shown in Figure 4.28 is a normally-closed delay-on valve. Many other time delay valves (delay-off, delay on/off, normally- open) can be obtained. All use the basic principle of the air reservoir and needle valve.

The symbol of a normally-dosed time delay valve is shown in Figure 4.28c.

Proportional Valves
The solenoid valves described so far act, to some extent, like an electrical switch, i.e. they can be On or Off. In many applications it is required to remotely control speed, pressure or force via an electrical signal. This function is provided by proportional valves.

A typical two position solenoid is only required to move the spool between 0 and 100% stroke against the restoring force of a spring. To ensure predictable movement between the end positions the solenoid must also increase its force as the spool moves to ensure the solenoid force is larger than the increasing opposing
spring force at all positions.

A proportional valve has a different design requirement. The spool position can be set anywhere between 0% and 100% stroke by varying the solenoid current. To give a predictable response the solenoid must produce a force which is dependent solely on the

current and not on the spool position, i.e. the force for a given current must be constant over the full stroke range. Furthermore, the force must be proportional to the current.

Figure 4.29 shows a typical response. The force from the solenoid is opposed by the force from a restoring spring, and the spool will move to a position where the two forces are equal. With a current of 0.75 A, for example, the spool will move to 75% of its stroke.

The spool movement in a proportional valve is small; a few mm stroke is typical. The valves are therefore very vulnerable to stiction, and this is reduced by using a 'wet' design which immerses the solenoid and its core in hydraulic fluid.

A proportional valve should produce a fluid flow which is proportional to the spool displacement. The spools therefore use four triangular metering notches in the spool lands as shown on Figure 4.30. As the spool is moved to the right, port A will progressively link to the tank and port B to the pressure line.

The symbol for this valve is also shown. Proportional valves are drawn with parallel lines on the connection sides of the valve block on circuit diagrams.

Figure 4.30 gives equal flow rates to both A and B ports.Cylinders have different areas on the full bore and annulus sides

(see Figure 5.4). To achieve equal speeds in both directions, the notches on the lands must have different areas. With a 2:1 cylinder ratio, half the number of notches are used on one side.

Figure 4.31 shows the construction and symbol for a restricted centre position valve. Here the extended notches provide a restricted (typically 3%) flow to tank from the A and B ports when the valve is in the centre position.

Types of Control Valves(Part 1)

amrxp2005@gmail.com (Engineer) — Sat, 06 Dec 2008 01:57:23 PST

Types of control valve

There are essentially three types of control valve; poppet valves, spool valves and rotary valves.

Poppet valves
In a poppet valve, simple discs, cones or balls are used in conjunction with simple valve seats to control flow. Figure 4.9 shows the construction and symbol of a simple 2/2 normally-closed valve, where depression of the pushbutton lifts the ball off its seat and

allows fluid to flow from port P to port A. When the button is released, spring and fluid pressure force the ball up again closing the valve.

Figure 4.10 shows the construction and symbol of a disc seal 3/2 poppet. With the pushbutton released, ports A and R are linked via the hollow pushbutton stem. If the pushbutton is pressed, port R is first sealed, then the valve disc pushed down to open the valve and connect ports P and A. As before, spring and fluid pressure from
port P closes the valve.

The valve construction and symbol shown in Figure 4.11 is a poppet changeover 4/2 valve using two stems and disc valves. With the pushbutton released, ports A and R are linked via the hollow left-hand stem and ports P and B linked via the normally-open right hand disc valve. When the pushbutton is pressed, the link between ports A and R is first closed, then the link between P and B closed.

The link between A and P is next opened, and finally the link between B and R opened. When the pushbutton is released, air and spring pressure puts the valve back to its original state.

Poppet valves are simple, cheap and robust, but it is generally simpler to manufacture valves more complicated than those shown in Figure 4.11 by using spool valves. Further, a major disadvantage of poppet valves is the force needed to operate them. In the poppet valve of Figure 4.10, for example, the force required on the pushbutton
to operate the valve is P x a newtons. Large capacity valves need large valve areas, leading to large operating force. The high pressure in hydraulic systems thus tends to prevent use of simple

poppet valves and they are, therefore, mainly found in low pressure pneumatic systems.

Spool valves
Spool (or slide) valves are constructed with a spool moving horizontally within the valve body, as shown for the 4/2 valve in Figure 4.12. Raised areas called 'lands' block or open ports to give the required operation.

The operation of a spool valve is generally balanced. In the valve construction in Figure 4.12b, for example, pressure is applied to opposing faces D and E and low tank pressure to faces F and G.

There is no net force on the spool from system pressure, allowing the spool to be easily moved.

Figure 4.13 is a changeover 4/2 spool valve. Comparison of the valves shown in Figures 4.12 and 4.13 shows they have the same body construction, the only difference being the size and position of lands on the spool. This is a major cost-saving advantage of spool valves; different operations can be achieved with a common body and different spools. This obviously reduces manufacturing costs.

Figure 4.14 shows various forms of three position changeover valves; note, again, these use one body with different functions achieved by different land patterns.
Spool valves are operated by shifting the spool. This can be achieved by button, lever or striker, or remotely with a solenoid.

Self-centring can easily be provided if springs are mounted at the end of the spool shaft.

Solenoid-operated valves commonly work at 24 V DC or 110 V AC. Each has its own advantages and disadvantages. A DC power supply has to be provided for 24 V DC solenoids, which, in large systems, is substantial and costly. Operating current of a 24 V solenoid is higher than a 110 V solenoid's. Care must be taken with plant cabling to avoid voltage drops on return legs if a common single line return is used.

Current through a DC solenoid is set by the winding resistance. Current in an AC solenoid, on the other hand, is set by the inductance of the windings, and this is usually designed to give a high inrush current followed by low holding current. This is achieved by using the core of the solenoid (linked to the spool) to raise the coil inductance when the spool has moved. One side effect of this is that a jammed spool results in a permanent high current which can damage the coil or the device driving it.

Each and every AC solenoid should be protected by an individual fuse. DC solenoids do not suffer from this characteristic. A burned out DC solenoid coil is
almost unknown.

Whatever form of solenoid is used it is very useful when fault finding to have local electrical indication built into the solenoid plug top. This allows a fault to be quickly identified as either an electrical or hydraulic problem. Fault finding is discussed further in Chapter 8.

A solenoid can exert a pull or push of about 5 to 10 kg. This is adequate for most pneumatic spool valves, but is too low for direct operation of large capacity hydraulic valves. Here pilot operation must be used, a topic discussed later.

Rotary valves
Rotary valves consist of a rotating spool which aligns with holes in the valve casing to give the required operation. Figure 4.15 shows the construction and symbol of a typical valve with centre off action.

Rotary valves are compact, simple and have low operating forces. They are, however, low pressure devices and are consequently mainly used for hand operation in pneumatic systems.

Pilot-operated valves
With large capacity pneumatic valves (particularly poppet valves) and most hydraulic valves, the operating force required to move the valve can be large. If the required force is too large for a solenoid or manual operation, a two-stage process called pilot operation is used.

The principle is shown in Figure 4.16. Valve 1 is the main operating valve used to move a ram. The operating force required to move the valve, however, is too large for direct operation by a solenoid, so a second smaller valve 2, known as the pilot valve, has been added to allow the main valve to be operated

by system pressure. Pilot pressure lines are normally shown dotted in circuit diagrams, and pilot ports on main valves are denoted Z, Y, X and so on.

In Figure 4 16, pilot port Z is depressurised with the solenoid deenergised, and the ram is retracted. When the solenoid is energised valve 2 changes over, pressurising Z; causing valve 1 to energize and the ram to extend.

Although pilot operation can be achieved with separate valves it is more usual to use a pilot/main valve assembly manufactured as a complete ready made unit. Figure 4.17 shows the operation of a pilot-operated 3/2 pneumatic valve. The solenoid operates
the small pilot valve directly. Because this valve has a small area, a low operating force is required. The pilot valve applies line pressure to the top of the control valve causing it to move down, closing the exhaust port. When it contacts the main valve disc there are two forces acting on the valve stem. The pilot valve applies a downwards force of P x D, where P is the line pressure and D is the area of the control valve. Line pressure also applies an upwards force P x E to the stem, where E is the area of the main valve.

The area of the control valve, D, is greater than area of the main valve E, so the downwards force is the larger and the valve opens.

When the solenoid de-energises, the space above the control valve is vented. Line an spring pressure on the main valve causes the valve stem to rise again, venting port A.

A hydraulic 4/2 pilot-operated spool valve is shown in Figure4.18. The ends of the pilot spool in most hydraulic pilot-operated valves are visible from outside the valve. This is useful from a maintenance viewpoint as it allows the operation of a valve to be
checked. In extreme cases the valve can be checked by pushing the pilot spool directly with a suitably sized rod (welding rod is ideal !).

Care must be taken to check solenoid states on dual solenoid valves before attempting manual operation. Overriding an energised AC solenoid creates a large current which may damage the coil, (or blow the fuse if the solenoid has correctly installed protection).

Check valves
Check valves only allow flow in one direction and, as such, are similar in operation to electronic diodes. The simplest constructionis the ball and seat arrangement of the valve in Figure 4.19a, commonly used in pneumatic systems. The right angle construction in Figure 4.19b is better suited to the higher pressures of a hydraulic

system. Free flow direction is normally marked with an arrow on the valve casing.

A check valve is represented by the graphic symbols in Figure 4.20. The symbol in Figure 4.20a is rather complex and the simpler symbol in Figure 4.20b is more commonly used.

Figure 4.21 illustrates several common applications of check valves. Figure 4.21a shows a combination pump, used where an application requires large volume and low pressure, or low volume and high pressure. A typical case is a clamp required to engage quickly (high volume and low pressure) then grip (minimal volume but high pressure). Pump 1 is the high volume and low pressure pump, and pump 2 the high pressure pump. In high volume mode both pumps deliver to the system, pump 1 delivering through the check valve V 3. When high pressure is required, line pressure at X rises operating unloading valve V 1 via pilot port Z taking pump 1 off load. Pump 2 delivers the required pressure set by relief valve V 2, with the check valve preventing fluid leaking back to pump 1 and V1.

Figure 4.21b shows a hydraulic circuit with a pressure storage device called an accumulator (described in a later chapter). Here a check valve allows the pump to unload via the pressure regulating valve, while still maintaining system pressure from the accumulator.

A spring-operated check valve requires a small pressure to open (called the cracking pressure) and acts to some extent like a low pressure relief valve. This characteristic can be used to advantage.

In Figure 4.21c pilot pressure is derived before a check valve, and in Figure 4.21 d a check valve is used to protect a blocked filter by diverting flow around the filter when pressure rises. A check valve is also included in the tank return to prevent fluid being sucked out of the tank when the pump is turned off.

Pilot-operated check valves
The cylinder in the system in Figure 4.22 should, theoretically, hold position when the control valve is in its centre, off, position. In practice, the cylinder will tend to creep because of leakage in the control valve.

Check valves have excellent sealage in the closed position, but a simple check valve cannot be used in the system in Figure 4.22 because flow is required in both directions. A pilot-operated check is similar to a basic check valve but can be held open permanently by application of an external pilot pressure signal.

There are two basic forms of pilot-operated check valves, shown in Figure 4.23. They operate in a similar manner to basic check valves, but with pilot pressure directly opening the valves. In the 4C valve shown in Figure 4.23a, inlet pressure assists the pilot. The

symbol of a pilot-operated check valve is shown in Figure 4.23c. The cylinder application of Figure 4.22 is redrawn with pilot operated check valves in Figure 4.23d. The pilot lines are connected to the pressure line feeding the other side of the cylinder. For any cylinder movement, one check valve is held open by flow (operating
as a normal check valve) and the other is held open by pilot pressure. For no required movement, both check valves are closed and the cylinder is locked in position.

Restriction check valves
The speed of a hydraulic or pneumatic actuator can be controlled by adjusting the rate at which a fluid is admitted to, or allowed out from, a device. This topic is discussed in more detail in Chapter 5 but a speed control is often required to be direction-sensitive and this requires the inclusion of a check valve.

A restriction check valve (often called a throttle relief valve in pneumatics) allows full flow in one direction and a reduced flow in the other direction. Figure 4.24a shows a simple hydraulic valve and Figure 4.24b a pneumatic valve. In both, a needle valve sets restricted flow to the required valve. The symbol of a restriction
check valve is shown in Figure 4.24c.

Figure 4.24d shows a typical application in which the cylinder extends at full speed until a limit switch makes, then extend further at low speed. Retraction is at full speed.

A restriction check valve V 2 is fitted in one leg of the cylinder. With the cylinder retracted, limit-operated valve V 3 is open allowing free flow of fluid from the cylinder as it extends. When the striker plate on the cylinder ram hits the limit, valve V 3 closes and flow out of the cylinder is now restricted by the needle valve setting
of valve V 2. In the reverse direction, the check valve on valve V 2 opens giving full speed of retraction.

Introduction to Control Valves (symbols)

amrxp2005@gmail.com (Engineer) — Sat, 06 Dec 2008 01:17:49 PST

Pneumatic and hydraulic systems require control valves to direct and regulate the flow of fluid from compressor or pump to the various load devices. Although there are significant practical differences between pneumatic and hydraulic devices (mainly arising from differences in operating pressures and types of seals needed for gas or liquid) the operating principles and descriptions are very similar.
Although valves are used for many purposes, there are essentially only two types of valve. An infinite position valve can take up any position between open and closed and, consequently, can be used to modulate flow or pressure. Relief valves described in earlier chapters are simple infinite position valves.
Most control valves, however, are only used to allow or block flow of fluid. Such valves are called finite position valves. An analogy between the two types of valve is the comparison between an electric light dimmer and a simple on/off switch.
Connections to a valve are termed 'ports'. A simple on/off valve therefore has two
ports. Most control valves, however, have four ports shown in hydraulic and pneumatic forms in Figure 4.1.
In both the load is connected to ports labelled A, B and the pressure supply (from pump or compressor) to port E In the hydraulic valve, fluid is returned to the tank from port T. In the pneumatic valve return air is vented from port R.
Figure 4.2 shows internal operation of valves. To extend the ram, ports P and B are connected to deliver fluid and ports A and T connected to return fluid. To retract the ram, ports P and A are connected to deliver fluid and ports B and T to return fluid.

Another consideration is the number of control positions. Figure 4.3 shows two possible control schemes. In Figure 4.3a, the ram is controlled by a lever with two positions; extend or retract. This valve has two control positions (and the ram simply drives to one end or other of its stroke). The valve in Figure 4.3b has three positions;
extend, off, retract. Not surprisingly the valve in Figure 4.3a is called a two position valve, while that in Figure 4.3b is a three position valve.

Finite position valves are commonly described as a port/position valve where port is the number of ports and position is the number of positions. Figure 4.3a therefore illustrates a 4/2 valve, and Figure 4.3b shows a 4/3 valve. A simple block/allow valve is a 2/2 valve.

The numbers of ports and positions does not, however, completely describe the valve. We must also describe its action. Figure 4.4 shows one possible action for the 4/3 valve of Figure 4.3b.
Extend and retract connections are similar, but in the off position ports P and T are connected-unloading the pump back to the tank without need of a separate loading valve, while leaving the ram locked in position. (This approach could, of course, only be used where the pump supplies one load). Other possible arrangements may block all four ports in the off position (to maintain pressure), or connect ports A, B and T (to leave the ram free in the off position).
A complete valve description thus needs number of ports, number of positions and the control action.

Graphic symbols
Simple valve symbols have been used so far to describe control actions. From the discussions in the previous section it can be seen that control actions can easily become too complex for representation by sketches showing how a valve is constructed.

A set of graphic symbols has therefore evolved (similar, in principle, to symbols used on electrical circuit diagrams). These show component function without showing the physical construction of each device. A 3/2 spool valve and a 3/2 rotary valve with the same function have the same symbol; despite their totally different constructions.

Symbols are described in various national documents; DIN24300, BS2917, ISO1219 and the new ISO5599, CETOP RP3 plus the original American JIC and ANSI symbols. Differences between these are minor.

A valve is represented by a square for each of its switching positions. Figure 4.5a thus shows the symbol of a two position valve, and Figure 4.5b a three position valve. Valve positions can be represented by letters a, b, c and so on, with 0 being used for a central neutral position.

Ports of a valve are shown on the outside of boxes in normal unoperated or initial position. Four ports have been added to the two position valve symbol shown in Figure 4.5c. Designations given to ports are normally:

Port Designation
Working lines A, B, C and so on

Pressure (power) supply P

Exhaust/Return R, S, T and so on

Control (Pilot) Lines (T normally used for hydraulic systems, R and S for
pneumatic systems) Z, Y, X and so on

ISO 5599 proposes to replace these letters with numbers, a retrograde step in the author's opinion.

Arrow-headed lines represent direction of flow. In Figure 4.6a, for example fluid is delivered from port P to port A and returned from port B to port T when the valve is in its normal state a. In state b, flow is reversed. This valve symbol corresponds to the valve represented in Figures 4.2 and 4.3a.

Shut off positions are represented by "r, as shown by the central position of the valve in Figure 4.6b, and internal flow paths can be represented as shown in Figure 4.6c. This latter valve, incidentally, vents the load in the off position.

In pneumatic systems, lines commonly vent to atmosphere directly at the valve, as shown by port R in Figure 4.6d.

Figure 4.7a shows symbols for the various ways in which valves can be operated. Figure 4.7b thus represents a 4/2 valve operated by a pushbutton. With the pushbutton depressed the ram extends. With the pushbutton released, the spring pushes the valve to position a and the ram retracts.

Actuation symbols can be combined. Figure 4.7c represents asolenoid-operated 4/3 valve, with spring return to centre.

Infinite position valve symbols are shown in Figure 4.8. A basic valve is represented by a single square as shown in Figure 4.8a, with the valve being shown in a normal, or non-operated, position.

Control is shown by normal actuation symbols" in Figure 4.8b, for example, the spring pushes the valve right decreasing flow, and pilot pressure pushes the valve left increasing flow. This represents a pressure relief valve which would be connected into a hydraulic system as shown in Figure 4.8c.

Cooling Towers

amrxp2005@gmail.com (Engineer) — Sat, 06 Dec 2008 05:12:07 PST

Purpose
Before the development of cooling towers, rivers, lakes, and cooling ponds were required to supply cooling. Through the development of the mechanical draft cooling tower, as little as one square foot of area is needed for every 1000 square feet required for a cooling pond or lake.

Cooling towers minimize the thermal pollution of the natural water heat sinks and allow the reuse of circulating water. An example of the manner in which a cooling tower can fit into a system is shown in Figure 10.

The cooling of the water in a cooling tower is accomplished by the direct contact of water and air. This cooling effect is provided primarily by an exchange of latent heat of vaporization resulting from evaporation of a small amount of water and by a transfer of sensible heat, which raises the temperature of the air. The heat transferred from the water to the air is dissipated to the atmosphere.

*Induced Draft Cooling Towers
Induced draft cooling towers, illustrated in Figure 11, are constructed such that the incoming circulating water is dispersed throughout the cooling tower via a spray header. The spray is directed down over baffles that are designed to maximize the contact between water and air. The air is drawn through the baffled area by large circulating fans and causes the evaporation and the cooling of the water.

The nomenclature for induced draft cooling towers, including some items not illustrated in Figure 11 is summarized below.

Casing - The casing encloses the walls of the cooling tower, exclusive of fan deck and louvers.

Collecting basin - The collecting basin is a receptacle beneath the cooling tower for collecting the water cooled by the cooling tower. It can be made of concrete, wood, metal, or an alternative material.
Certain necessary accessories are required such as sump, strainers, overflow, drain, and a makeup system.

Drift eliminators - The drift eliminators are parallel blades of PVC, wood, metal,
or an alternative material arranged on the air discharge side of the fill to remove entrained water droplets from the leaving air stream.

Driver - The driver is a device that supplies power to turn the fan. It is usually an electric motor, but turbines and internal combustion engines are occasionally used.

Drive shaft - The drive shaft is a device, including couplings, which transmits power from the driver to the speed reducer.

Fan - The fan is a device used to induce air flow through the cooling tower.

Fan deck - The fan deck is a horizontal surface enclosing the top of the cooling tower above the plenum that serves as a working platform for inspection and maintenance.

Fan stack - The fan stack is a cylinder enclosing the fan, usually with an eased inlet and an expanding discharge for increased fan efficiency.

Fill - The fill is PVC, wood, metal, or an alternative material that provides extended water surface exposure for evaporative heat transfer.

Intake louvers - The intake louvers are an arrangement of horizontal blades at the air inlets that prevent escape of falling water while allowing the entry of air.

Makeup valve - The makeup valve is a valve that introduces fresh water into the collection basin to maintain the desired collecting basin water level.

Overflow-The overflow is a drain that prevents the collecting basin from overflowing.

Partition - The partition is a baffle within a multicell cooling tower that is used to prevent air and/or water flow between adjacent cells.

Plenum - The plenum is the internal cooling tower area between the drift eliminators and the fans.

Speed reducer - The speed reducer is a right-angle gear box that transmits power to the fan while reducing the driver speed to that required for optimal fan performance.

Sump - The sump is a depressed portion of the collecting basin from which cold water is drawn to be returned to the connected system. The sump usually contains strainer screens, antivortex devices, and a drain or cleanout connection.

Distribution system - The distribution system is that portion of a cooling tower that distributes water over the fill area. It usually consists of one or more flanged inlets, flow control valves, internal headers, distribution basins, spray branches, metering orifices, and other related components.

*Forced Draft Cooling Towers
Forced draft cooling towers are very similar to induced draft cooling towers. The primary difference is that the air is blown in at the bottom of the tower and exits at the top. Forced draft cooling towers are the forerunner to induced draft cooling towers. Water distribution problems and recirculation difficulties discourage the use of forced draft cooling towers.

*Natural Convection Cooling Towers
Natural convection cooling towers, illustrated in Figure 12, use the principle of convective flow to provide air circulation. As the air inside the tower is heated, it rises through the tower. This process draws more air in, creating a natural air flow to provide cooling of the water. The basin at the bottom of the tower is open to the atmosphere. The cooler, more dense air outside the tower will flow in at the bottom and contribute to the air circulation within the tower. The air circulation will be self perpetuating due to the density difference between the warmer air inside and the cooler air outside.

The incoming water is sprayed around the circumference of the tower and cascades to the bottom. The natural convection cooling towers are much larger than the forced draft cooling towers and cost much more to construct. Because of space considerations and cost, natural convection cooling towers are built less frequently than other types.

Summary

The cooling tower removes heat from water used in cooling systems within the plant. The heat is released to the air rather than to a lake or stream. This allows facilities to locate in areas with less water available because the cooled water can be recycled. It also aids environmental efforts by not contributing to thermal pollution.

Induced draft cooling towers use fans to create a draft that pulls air through the cooling tower fill. Because the water to be cooled is distributed such that it cascades over the baffles, the air blows through the water, cooling it.

Forced draft cooling towers blow air in at the bottom of the tower. The air exits at the top of the tower. Water distribution and recirculation difficulties limit their use.

Natural convection cooling towers function on the basic principle that hot air rises. As the air inside the tower is heated, it rises through the tower. This process draws more air in, creating a natural air flow to provide cooling of the water.

Centrifugal Compressor

amrxp2005@gmail.com (Engineer) — Fri, 05 Dec 2008 17:28:54 PST

Introduction
Air compressors of various designs are used widely throughout DOE facilities in numerous applications. Compressed air has numerous uses throughout a facility including the operation of equipment and portable tools. Three types of designs include reciprocating, rotary, and centrifugal air compressors.

Centrifugal Compressors
The centrifugal compressor, originally
built to handle only large volumes of low
pressure gas and air (maximum of 40
psig), has been developed to enable it to
move large volumes of gas with discharge
pressures up to 3,500 psig. However,
centrifugal compressors are now most
frequently used for medium volume and
medium pressure air delivery. One
advantage of a centrifugal pump is the
smooth discharge of the compressed air.

The centrifugal force utilized by the
centrifugal compressor is the same force
utilized by the centrifugal pump. The air
particles enter the eye of the impeller,
designated D in Figure 6. As the impeller rotates, air is thrown against the casing of the compressor. The air
becomes compressed as more and more air is thrown out to the casing by the impeller blades.

The air is pushed along the path designated A, B, and C in Figure 6. The pressure of the air is increased as it is pushed along this path. Note in Figure 6 that the impeller blades curve forward, which is opposite to the backward curve used in typical centrifugal liquid pumps.

Centrifugal compressors can use a variety of blade orientation including both forward and backward curves as well as other designs.

There may be several stages to a centrifugal air compressor, as in the centrifugal pump, and the result would be the same; a higher pressure would be produced. The air compressor is used to create compressed or high pressure air for a variety of uses.

Some of its uses are pneumatic control devices, pneumatic sensors, pneumatic valve operators, pneumatic motors, and starting air for diesel engines.

Rotary Compressors

amrxp2005@gmail.com (Engineer) — Fri, 05 Dec 2008 17:23:13 PST

*The rotary lobe-type, illustrated in Figure 4,
features two mating lobe-type rotors mounted in a case. The lobes are gear driven at close clearance, but without metal-to-metal contact. The suction to the unit is located where the cavity made by the lobes is largest. As the lobes rotate, the cavity size is reduced, causing
compression of the vapor within. The compression continues until the discharge port is reached, at which point the vapor exits the compressor at a higher pressure.

*The rotary liquid seal ring-type,
illustrated in Figure 5, features a forward inclined, open impeller, in an oblong cavity filled with liquid. As the impeller
rotates, the centrifugal force causes the seal liquid to collect at the outer edge of
the oblong cavity. Due to the oblong configuration of the compressor case, large longitudinal cells are created and reduced to smaller ones. The suction port is positioned where the longitudinal cells are the largest, and for the discharge port, where they are smallest, thus causing the vapor within the cell to compress as the rotor rotates. The rotary liquid seal compressor is frequently used in specialized applications for the compression of extremely corrosive and exothermic gasses and is commonly used in commercial nuclear plants as a means of establishing initial condenser vacuum.

Reciprocating Compressors

amrxp2005@gmail.com (Engineer) — Fri, 05 Dec 2008 17:14:59 PST

c. A self-contained lubricating system for bearings, gears, and cylinder walls,
including a reservoir or sump for the lubricating oil, and a pump,
or other means of delivering oil to the various parts. On some compressors a
separate force-fed lubricator is installed to supply oil to the compressor
cylinders.

d. A regulation or control system designed to maintain the pressure in the
discharge line and air receiver (storage tank) within a predetermined range of
pressure.

e. An unloading system, which operates in conjunction with the regulator, to
reduce or eliminate the load put on the prime mover when starting the unit.

A section of a typical reciprocating single-stage, single-acting compressor cylinder is shown in Figure 2. Inlet and discharge valves are located in the clearance space and connected through ports in the cylinder head to the inlet and discharge connections.

During the suction stroke the compressor piston starts its downward stroke and the air under pressure in the clearance space rapidly expands until the pressure falls below that on the opposite side of the inlet valve (Figures 2B and 2C). This difference in pressure causes the inlet valve to open into the cylinder until the piston reaches the bottom of its stroke (Figure 2C).

During the compression stroke the piston starts upward, compression begins, and at point D has reached the same pressure as the compressor intake. The spring-loaded inlet valve then closes.

As the piston continues upward, air is compressed until the pressure in the cylinder becomes great enough to open the discharge valve against the pressure of the valve springs and the pressure of the discharge line (Figure 2E). From this point, to the end of the stroke (Figures 2E and 2A), the air compressed within the cylinder is discharged at practically constant pressure.

Pumps Characteristic CURVES

amrxp2005@gmail.com (Engineer) — Wed, 03 Dec 2008 07:34:54 PST

*Centrifugal Pump Characteristic Curves
For a given centrifugal pump operating at a constant speed, the flow rate through the pump is dependent upon the differential pressure or head developed by the pump.

The lower the pump head, the higher the flow rate. A vendor manual for a specific pump usually contains a curve of pump flow rate versus pump head called a pump characteristic curve. After a pump is installed in a system, it is usually tested to ensure that the flow rate and head of the pump are within the required specifications. A typical centrifugal pump characteristic curve is shown in Figure 11.

There are several terms associated with the pump characteristic curve that must be defined.

Shutoff head is the maximum head that can be developed by a centrifugal pump operating at a
set speed.

Pump runout is the maximum flow that can be developed by a centrifugal pump
without damaging the pump. Centrifugal pumps must be designed and operated to be protected from the conditions of pump runout or operating at shutoff head.

*Positive Displacement Pump Characteristic Curves
Positive displacement pumps deliver a definite volume of liquid for each cycle of pump operation. Therefore, the only factor that effects flow rate in an ideal positive displacement pump is the speed at which it operates. The
flow resistance of the system in which the pump is operating will not effect the flow rate through the pump.

Figure 21 shows the characteristic curve for a positive
displacement pump.

The dashed line in Figure 21 shows actual positive
displacement pump performance. This line reflects the
fact that as the discharge pressure of the pump increases,
some amount of liquid will leak from the discharge of the
pump back to the pump suction, reducing the effective
flow rate of the pump. The rate at which liquid leaks
from the pump discharge to its suction is called slippage.

Positive displacement pumps

amrxp2005@gmail.com (Engineer) — Wed, 03 Dec 2008 07:16:53 PST

Introduction
A positive displacement pump is one in which a definite volume of liquid is delivered for each cycle of pump operation. This volume is constant regardless of the resistance to flow offered by the system the pump is in, provided the capacity of the power unit driving the pump or pump component strength limits are not exceeded. The positive displacement pump delivers liquid in separate volumes with no delivery in between, although a pump having several chambers may have an overlapping delivery among individual chambers, which minimizes this effect. The positive displacement pump differs from centrifugal pumps, which deliver a continuous flow for any given pump speed and discharge resistance.

Positive displacement pumps can be grouped into three basic categories based on their design and operation. The three groups are reciprocating pumps, rotary pumps, and diaphragm pumps.

Principle of Operation
All positive displacement pumps operate on the same basic principle. This principle can be most easily demonstrated by considering a reciprocating positive displacement pump consisting of a single reciprocating piston in a cylinder with a single suction port and a single discharge port as shown in Figure 12. Check valves in the suction and discharge ports allow flow in only one direction.

Figure 12 Reciprocating Positive Displacement Pump Operation

During the suction stroke, the piston moves to the left, causing the check valve in the suction line between the reservoir and the pump cylinder to open and admit water from the reservoir.
During the discharge stroke, the piston moves to the right, seating the check valve in the suction line and opening the check valve in the discharge line. The volume of liquid moved by the pump in one cycle (one suction stroke and one discharge stroke) is equal to the change in the liquid volume of the cylinder as the piston moves from its farthest left position to its farthest right position.

Reciprocating Pumps
Reciprocating positive displacement pumps are generally categorized in four ways: direct-acting or indirect-acting; simplex or duplex; single-acting or double-acting; and power pump

Direct-Acting and Indirect-Acting Pumps
Some reciprocating pumps are powered by prime movers that also have reciprocating
motion, such as a reciprocating pump powered by a reciprocating steam piston. The piston rod of the steam piston may be directly connected to the liquid piston of the pump or it may be indirectly connected with a beam or linkage. Direct-acting pumps have a plunger on the liquid (pump) end that is directly driven by the pump rod (also the piston rod or extension thereof) and carries the piston of the power end. Indirect-acting pumps are driven by means of a beam or linkage connected to and actuated by the power piston rod of a separate reciprocating engine.

Simplex and Duplex Pumps
A simplex pump, sometimes referred to as a single pump, is a pump having a single liquid (pump) cylinder. A duplex pump is the equivalent of two simplex pumps placed side by side on the same foundation.

The driving of the pistons of a duplex pump is arranged in such a manner that when one piston is on its upstroke the other piston is on its downstroke, and vice versa. This arrangement doubles the capacity of the duplex pump compared to a simplex pump of comparable design.

Single-Acting and Double-Acting Pumps
A single-acting pump is one that takes a suction, filling the pump cylinder on the stroke in only one direction, called the suction stroke, and then forces the liquid out of the cylinder on the return stroke, called the discharge stroke. A double-acting pump is one that, as it fills one end of the liquid cylinder, is discharging liquid from the other end of the cylinder.

On the return stroke, the end of the cylinder just emptied is filled, and the end just filled is emptied. One possible arrangement for single-acting and double-acting pumps is shown in Figure 13.

Power Pumps
Power pumps convert rotary motion to low speed reciprocating motion by reduction
gearing, a crankshaft, connecting rods and crossheads. Plungers or pistons are driven by the crosshead drives. Rod and piston construction, similar to duplex double-acting steam pumps, is used by the liquid ends of the low pressure, higher capacity units. The higher pressure units are normally single-acting plungers, and usually employ three (triplex) plungers. Three or more plungers substantially reduce flow pulsations relative to simplex and even duplex pumps.

Figure 13 Single-Acting and Double-Acting Pumps

Power pumps typically have high efficiency and are capable of developing very high pressures.
They can be driven by either electric motors or turbines. They are relatively expensive pumps and can rarely be justified on the basis of efficiency over centrifugal pumps. However, they are frequently justified over steam reciprocating pumps where continuous duty service is needed due to the high steam requirements of direct-acting steam pumps.

In general, the effective flow rate of reciprocating pumps decreases as the viscosity of the fluid being pumped increases because the speed of the pump must be reduced. In contrast to centrifugal pumps, the differential pressure generated by reciprocating pumps is independent of fluid density. It is dependent entirely on the amount of force exerted on the piston. For more information on viscosity, density, and positive displacement pump theory, refer to the handbook on Thermodynamics, Heat Transfer, and Fluid Flow.

Rotary Pumps
Rotary pumps operate on the principle that a rotating vane, screw, or gear traps the liquid in the suction side of the pump casing and forces it to the discharge side of the casing. These pumps are essentially self-priming due to their capability of removing air from suction lines and producing a high suction lift. In pumps designed for systems requiring high suction lift and self priming features, it is essential that all clearances between rotating parts, and between rotating and stationary parts, be kept to a minimum in order to reduce slippage. Slippage is leakage of fluid from the discharge of the pump back to its suction.

Due to the close clearances in rotary pumps, it is necessary to operate these pumps at relatively low speed in order to secure reliable operation and maintain pump capacity over an extended period of time. Otherwise, the erosive action due to the high velocities of the liquid passing through the narrow clearance spaces would soon cause excessive wear and increased clearances, resulting in slippage.

There are many types of positive displacement rotary pumps, and they are normally grouped into three basic categories that include gear pumps, screw pumps, and moving vane pumps.

Simple Gear Pump
There are several variations of gear pumps. The simple gear pump shown in Figure 14 consists of two spur gears meshing together and revolving in opposite directions within a casing. Only a few thousandths of an inch clearance exists between the case and the gear faces and teeth extremities. Any liquid that fills the space bounded by two successive gear teeth and the case must follow along with the teeth as they revolve. When the gear teeth mesh with the teeth of the other gear, the space between the teeth is reduced, and the entrapped liquid is forced out the pump discharge pipe.

As the gears revolve and the teeth disengage, the space again opens on the suction side of the pump, trapping new quantities of liquid and carrying it around the pump case to the discharge. As liquid is carried away from the suction side, a lower pressure is created, which draws liquid in through the suction line.

Fig 24 Simple Gear Pump

With the large number of teeth usually employed on the gears, the discharge is relatively smooth and continuous, with small quantities of liquid being delivered to the discharge line in rapid succession. If designed with fewer teeth, the space between the teeth is greater and the capacity increases for a given speed; however, the tendency toward a pulsating discharge increases. In all simple gear pumps, power is applied to the shaft of one of the gears, which transmits power to the driven gear through their meshing teeth.

There are no valves in the gear pump to cause friction losses as in the reciprocating pump.

The high impeller velocities, with resultant friction losses, are not required as in the
centrifugal pump. Therefore, the gear pump is well suited for handling viscous fluids such as fuel and lubricating oils.

Other Gear Pumps
There are two types of gears used in gear pumps in addition to the simple spur gear. One type is the helical gear. A helix is the curve produced when a straight line moves up or down the surface of a cylinder. The other type is the herringbone gear. A herringbone gear is composed of two helixes spiraling in different directions from the center of the gear. Spur, helical, and herringbone gears are shown in Figure 15.

The helical gear pump has advantages over the simple spur gear. In a spur gear, the entire length of the gear tooth engages at the same time. In a helical gear, the point of engagement moves along the length of the gear tooth as the gear rotates. This makes the helical gear operate with a steadier discharge pressure and fewer pulsations than a spur gear pump.

The herringbone gear pump is also a modification of the simple gear pump. Its principal difference in operation from the simple spur gear pump is that the pointed center section of the space between two teeth begins discharging before the divergent outer ends of the preceding space complete discharging. This
overlapping tends to provide a steadier discharge pressure. The power transmission from the driving to the driven gear is also smoother and
quieter.

Lobe Type Pump
The lobe type pump shown in Figure 16 is another variation of the simple gear pump. It is considered as a simple gear pump having only two or three teeth per rotor; otherwise, its operation or the
explanation of the function of its parts is no different. Some designs of lobe pumps are fitted with replaceable gibs, that is, thin plates carried in grooves at the extremity of each lobe where they
make contact with the casing. The gib promotes tightness and absorbs radial wear.

Screw-Type Positive Displacement Rotary Pump
There are many variations in the design of the screw type positive displacement, rotary pump. The primary differences consist of the number of intermeshing screws involved, the pitch of the screws, and the general direction of fluid flow. Two common designs are the two-screw, low-pitch, double-flow pump and the three-screw, high-pitch, double-flow pump.

*Two-Screw, Low-Pitch, Screw Pump
The two-screw, low-pitch, screw pump consists of two screws that mesh with close
clearances, mounted on two parallel shafts. One screw has a right-handed thread, and
the other screw has a left-handed thread. One shaft is the driving shaft and drives the
other shaft through a set of herringbone timing gears. The gears serve to maintain
clearances between the screws as they turn and to promote quiet operation. The
screws rotate in closely fitting duplex cylinders that have overlapping bores. All
clearances are small, but there is no actual contact between the two screws or between
the screws and the cylinder walls.

The complete assembly and the usual flow Figure 18 Three-Screw, High-Pitch, Screw Pump path are shown in Figure 17. Liquid is trapped at the outer end of each pair of screws. As the first space between the screw threads rotates away from the opposite screw, a one-turn, spiral-shaped quantity of liquid is enclosed when the end of the screw again meshes with the opposite screw. As the
screw continues to rotate, the entrapped spiral turns of liquid slide along the cylinder toward the center discharge space while the next slug is being entrapped.
Each screw functions similarly, and each pair of screws discharges an equal quantity of liquid in opposed streams toward the center, thus eliminating hydraulic thrust. The removal of liquid from the suction end by the screws produces a
reduction in pressure, which draws liquid through the suction line.

*Three-Screw, High-Pitch, Screw Pump
The three-screw, high-pitch, screw pump, shown in Figure 18, has many of the same elements as the two-screw, low-pitch, screw pump, and their operations are similar.
Three screws, oppositely threaded on each end, are employed. They rotate in a triple cylinder, the two outer bores of which overlap the center bore. The pitch of the screws is much higher than in the low pitch screw pump; therefore, the center screw, or
power rotor, is used to drive the two outer idler rotors directly without external timing gears.

Pedestal bearings at the base support the weight of the rotors and maintain their axial position. The liquid being pumped enters the suction opening, flows through passages around the rotor housing, and through the screws from each end, in opposed
streams, toward the center discharge. This eliminates unbalanced hydraulic thrust.

The screw pump is used for pumping viscous fluids, usually lubricating, hydraulic, or fuel oil.

Rotary Moving Vane Pump
The rotary moving vane pump shown in Figure 19 is another type of positive displacement pump used. The pump consists of a cylindrically bored housing with a suction inlet on one side and a discharge outlet on the other. A cylindrically shaped rotor with a diameter smaller than the cylinder is driven about an axis placed above the centerline of the cylinder.

The clearance between rotor and cylinder is small at the top but increases at the bottom.

The rotor carries vanes that move in and out as it rotates to maintain sealed spaces between the rotor and the cylinder wall. The vanes trap liquid or gas on the suction side and carry it to the discharge side, where contraction of the space expels it through the discharge line.The vanes may swing on pivots, or they may slide in slots in the rotor.

Diaphragm Pumps
Diaphragm pumps are also classified as positive displacement pumps because the diaphragm acts as a limited displacement piston. The pump will function when a diaphragm is forced into reciprocating motion by mechanical linkage, compressed air, or fluid from a pulsating, external source. The pump construction eliminates any contact between the liquid being pumped and the source of energy. This eliminates the possibility of leakage, which is important when handling toxic or very expensive liquids. Disadvantages include limited head and capacity range, and the necessity of check valves in the suction and discharge nozzles. An example of a diaphragm
pump is shown in Figure 20.

Classification of Centrifugal Pumps

amrxp2005@gmail.com (Engineer) — Wed, 03 Dec 2008 05:16:45 PST

Centrifugal Pump Classification by Flow

Centrifugal pumps can be classified based on the manner in which fluid flows through the pump. The manner in which fluid flows through the pump is determined by the design of the pump casing and the impeller. The three types of flow through a centrifugal pump are radial flow, axial flow, and mixed flow.

Radial Flow Pumps

In a radial flow pump, the liquid enters at the center of the impeller and is directed out along the impeller blades in a direction at right angles to the pump shaft. The impeller of a typical radial flow pump and the flow through a radial flow pump are shown in Figure 6.

Fig 6 Radial Flow Centrifugal Pump

Axial Flow Pumps

In an axial flow pump, the impeller pushes the liquid in a direction parallel to the pump shaft. Axial flow pumps are sometimes called propeller pumps because they operate essentially the same as the propeller of a boat. The impeller of a typical axial flow pump and the flow through a radial flow pump are shown in Figure 7.

Fig 7 Axial Flow Centrifugal Pump

Mixed Flow Pumps

Mixed flow pumps borrow characteristics from both radial flow and axial flow pumps.
As liquid flows through the impeller of a mixed flow pump, the impeller blades push the liquid out away from the pump shaft and to the pump suction at an angle greater than 90o. The impeller of a typical mixed flow pump and the flow through a mixed flow pump are shown in Figure 8.

Fig 8 Mixed Flow Centrifugal Pump

Multi-Stage Centrifugal Pumps

A centrifugal pump with a single impeller that can develop a differential pressure of more than 150 psid between the suction and the discharge is difficult and costly to design and construct. A more economical approach to developing high pressures with a single centrifugal pump is to include multiple impellers on a common shaft within the same pump casing. Internal channels in the pump casing route the discharge of one impeller to the suction of another impeller.

Figure 9 shows a diagram of the arrangement of the impellers of a four-stage pump. The water enters the pump from the top left and passes through each of the four impellers in series, going from left to right. The water goes from the volute surrounding the discharge of one impeller to the suction of the next impeller.

A pump stage is defined as that portion of a centrifugal pump consisting of one impeller and its associated components. Most centrifugal pumps are single-stage pumps, containing only one impeller. A pump containing seven impellers within a single casing would be referred to as a seven-stage pump or, or generally, as a multi-stage pump.

Fig 9 Multi Stage Centrifugal Pump

Governor

amrxp2005@gmail.com (Engineer) — Wed, 03 Dec 2008 01:03:58 PST

Governor

Diesel engine speed is controlled solely by the amount of fuel injected into the engine by the injectors. Because a diesel engine is not self-speed-limiting, it requires not only a means of changing engine speed (throttle control) but also a means of maintaining the desired speed. The governor provides the engine with the feedback mechanism to change speed as needed and to maintain a speed once reached.

A governor is essentially a speed-sensitive device, designed to maintain a constant engine speed regardless of load variation. Since all governors used on diesel engines control engine speed through the regulation of the quantity of fuel delivered to the cylinders, these governors may be classified as speed-regulating governors. As with the engines themselves there are many types and variations of governors. In this module, only the common mechanical-hydraulic type governor will be reviewed.
The major function of the governor is determined by the application of the engine. In an engine that is required to come up and run at only a single speed regardless of load, the governor is called a constant-speed type governor. If the engine is manually controlled, or controlled by an outside device with engine speed being controlled over a range, the governor is called a variable speed type governor. If the engine governor is designed to keep the engine speed above a minimum and below a maximum, then the governor is a speed-limiting type. The last category of governor is the load limiting type. This type of governor limits fuel to ensure that the engine is not loaded above a specified limit. Note that many governors act to perform several of these functions simultaneously.

Operation of a Governor
The following is an explanation of the operation of a constant speed, hydraulically compensated governor using the Woodward brand governor as an example. The principles involved are common in any mechanical and hydraulic governor.

The Woodward speed governor operates the diesel engine fuel racks to ensure a constant engine speed is maintained at any load. The governor is a mechanical-hydraulic type governor and receives its supply of oil from the engine lubricating system. This means that a loss of lube oil pressure will cut off the supply of oil to the governor and cause the governor to shut down the engine. This provides the engine with a built-in shutdown device to protect the engine in the event of loss of lubricating oil pressure.

Simplified Operation of the Governor
The governor controls the fuel rack position through a combined action of the hydraulic piston and a set of mechanical flyweights, which are driven by the engine blower shaft.

Figure 28 provides an illustration of a functional diagram of a mechanical-hydraulic
governor. The position of the flyweights is determined by the speed of the engine. As
the engine speeds up or down, the weights move in or out. The movement of the
flyweights, due to a change in engine speed, moves a small piston (pilot valve) in the
governor's hydraulic system. This motion adjusts flow of hydraulic fluid to a large
hydraulic piston (servo-motor piston). The large hydraulic piston is linked to the fuel
rack and its motion resets the fuel rack for increased/decreased fuel.

Fig 28 simplified Mechanical-Hydraulic Governor

Detailed Operation of the Governor
With the engine operating, oil from the engine lubrication system is supplied to the
governor pump gears, as illustrated in Figure 29. The pump gears raise the oil pressure to a value determined by the spring relief valve. The oil pressure is maintained in the annular space between the undercut portion of the pilot valve plunger and the bore in the pilot valve bushing. For any given speed setting, the spring speeder exerts a force that is opposed by the centrifugal force of the revolving flyweights. When the two forces are equal, the control land on the pilot valve plunger covers the lower ports in the pilot valve bushing.

Fig 29 Cutway of Woodward Governor

Under these conditions, equal oil pressures are maintained on both sides of the buffer piston and tension on the two buffer springs is equal. Also, the oil pressure is equal on both sides of the receiving compensating land of the pilot valve plunger due to oil passing through the compensating needle valve. Thus, the hydraulic system is in balance, and the engine speed remains constant.

When the engine load increases, the engine starts to slow down in speed. The reduction in engine speed will be sensed by the governor flyweights. The flyweights are forced inward (by the spring), thus lowering the pilot valve plunger (again, due to the downward spring force). Oil under pressure will be admitted under the servo-motor piston (topside of the buffer piston) causing it to rise. This upward motion of the servo-motor piston will be transmitted through the terminal lever to the fuel racks, thus increasing the amount o f fuel injected into the engine. The oil that forces the servo-motor piston upward also forces the buffer piston upward because the oil pressure on each side of the piston is unequal.

This upward motion of the piston compresses the upper buffer spring and relieves the pressure on the lower buffer spring.

The oil cavities above and below the buffer piston are common to the receiving
compensating land on the pilot valve plunger. Because the higher pressure is below the compensating land, the pilot valve plunger is forced upward, recentering the flyweights and causing the control land of the pilot valve to close off the regulating port. Thus, the upward movement of the servo-motor piston stops when it has moved far enough to make the necessary fuel correction.

Oil passing through the compensating needle valve slowly equalizes the pressures above and below the buffer piston, thus allowing the buffer piston to return to the center position, which in turn equalizes the pressure above and below the receiving
compensating land. The pilot valve plunger then moves to its central position and the
engine speed returns to its original setting because there is no longer any excessive
outward force on the flyweights.

The action of the flyweights and the hydraulic feedback mechanism produces stable
engine operation by permitting the governor to move instantaneously in response to the load change and to make the necessary fuel adjustment to maintain the initial engine speed.

Fuel Injectors

amrxp2005@gmail.com (Engineer) — Wed, 03 Dec 2008 00:49:05 PST

Fuel Injectors

Each cylinder has a fuel injector designed to meter and inject fuel into the cylinder at the proper instant. To accomplish this function, the injectors are actuated by the engine's camshaft. The camshaft provides the timing and pumping action used by the injector to inject the fuel. The injectors meter the amount of fuel injected into the cylinder on each stroke. The amount of fuel to be injected by each injector is set by a mechanical linkage called the fuel rack. The fuel rack position is controlled by the engine's governor. The governor determines the amount of fuel required to maintain the desired engine speed and adjusts the amount to be injected by adjusting
the position of the fuel rack.

Each injector operates in the following manner. As illustrated in Figure 26, fuel under pressure enters the injector through the injector's filter cap and filter element. From the filter element the fuel travels down into the supply chamber (that area between the plunger bushing and the spill deflector). The plunger operates up and down in the bushing, the bore of which is open to the fuel supply in the supply chamber by two funnel-shaped ports in the plunger bushing.

Figure 26 Fuel Injector Cutway

The motion of the injector rocker arm (not shown) is transmitted to the plunger by the injector follower which bears against the follower spring. As the plunger moves downward under pressure of the injector rocker arm, a portion of the fuel trapped under the plunger is displaced into the supply chamber through the lower port until the port is closed off by the lower end of the plunger. The fuel trapped below the plunger is then forced up through the central bore of the plunger and back out the upper port until the upper port is closed off by the downward motion of the plunger.

With the upper and lower ports both closed off, the remaining fuel under the plunger is subjected to an increase in pressure by the downward motion of the plunger.
When sufficient pressure has built up, the injector valve is lifted off its seat and the fuel is forced through small orifices in the spray tip and atomized into the combustion chamber. A check valve, mounted in the spray tip, prevents air in the combustion chamber from flowing back into the fuel injector. The plunger is then returned back to its original position by the injector follower spring.

On the return upward movement of the plunger, the high pressure cylinder within the bushing is again filled with fresh fuel oil through the ports. The constant circulation of fresh, cool fuel through the injector renews the fuel supply in the chamber and helps cool the injector. The fuel flow also effectively removes all traces of air that might otherwise accumulate in the system.

The fuel injector outlet opening, through which the excess fuel returns to the fuel return manifold and then back to the fuel tank, is adjacent to the inlet opening and contains a filter element exactly the same as the one on the fuel inlet side. In addition to the reciprocating motion of the plunger, the plunger can be rotated during operation around its axis by the gear which meshes with the fuel rack. For metering the fuel, an upper helix and a lower helix are machined in the lower part of the plunger. The relation of the helices to the two ports in the injector bushing changes with the rotation of the plunger.

Changing the position of the helices, by rotating the plunger, retards or advances the closing of the ports and the beginning and ending of the injection period. At the same time, it increases or decreases the amount of fuel injected into the cylinder. Figure 27 illustrates the various plunger positions from NO LOAD to FULL LOAD. With the control rack pulled all the way (no injection), the upper port is not closed by the helix until after the lower port is uncovered.

Consequently, with the rack in this position, all of the fuel is forced back into the supply chamber and no injection of fuel takes place. With the control rack pushed all the way in (full injection), the upper port is closed shortly after the lower port has been covered, thus producing a maximum effective stroke and maximum fuel injection. From this no-injection position to the full-injection position (full rack movement), the contour of the upper helix advances the closing of the ports and the beginning of injection.

Fig 27 Fuel Injector Plunger

Two Stroke Diesel Engine

amrxp2005@gmail.com (Engineer) — Wed, 03 Dec 2008 00:33:36 PST

The Two-Stroke Cycle
Like the four-stroke engine, the two-stroke engine must go through the same four events: intake, compression, power, and exhaust. But a two-stroke engine requires only two strokes of the piston to complete one full cycle. Therefore, it requires only one rotation of the crankshaft to complete a cycle. This means several events must occur during each stroke for all four events to be completed in two strokes, as opposed to the four-stroke engine where each stroke basically contains one event.

In a two-stroke engine the camshaft is geared so that it rotates at the same speed as the crankshaft (1:1). The following section will describe a two-stroke, supercharged, diesel engine having intake ports and exhaust valves with a 3.5-inch bore and 4-inch stroke with a 16:1 compression ratio, as it passes through one complete cycle. We will start on the exhaust stroke. All the timing marks given are generic and will vary from engine to engine.

Exhaust and Intake
At 82° ATDC, with the piston near the end of its power stroke, the exhaust cam begins to lift the exhaust valves follower. The valve lash is taken up, and 9° later (91° ATDC), the rocker arm forces the exhaust valve off its seat. The exhaust gasses start to escape into the exhaust manifold, as shown in Figure 21. Cylinder pressure starts to decrease.

After the piston travels three-quarters of its (down) stroke, or 132° ATDC of crankshaft rotation, the piston starts to uncover the inlet ports. As the exhaust valve is still open, the uncovering of the inlet ports lets the compressed fresh air enter the cylinder and helps cool the cylinder and scavenge the cylinder of the remaining exhaust gasses (Figure 22).Commonly, intake and exhaust occur over approximately 96° of crankshaft rotation.

At 43° ABDC, the camshaft starts to close the exhaust valve. At 53° ABDC (117° BTDC), the camshaft has rotated sufficiently to allow the spring pressure to close the
exhaust valve. Also, as the piston travels past 48°ABDC (5° after the exhaust valve starts closing), the intake ports are closed off by the piston.

Compression
After the exhaust valve is on its seat (53° ATDC), the temperature and pressure begin to rise in nearly the same fashion as in the four-stroke engine. Figure 23 illustrates the compression in a 2-stroke engine. At 23° BTDC the injector cam begins to lift the
injector follower and pushrod. Fuel injection continues until 6° BTDC (17 total degrees of injection), as illustrated in Figure 24.

Power
The power stroke starts after the piston passes TDC. Figure 25 illustrates the power stroke which continues until the piston reaches 91° ATDC, at which point the
exhaust valves start to open and a new cycle begins.

Four Stroke Diesel Engine

amrxp2005@gmail.com (Engineer) — Mon, 01 Dec 2008 10:18:42 PST

Four Stroke Diesel Engine
In a four-stroke engine the camshaft is geared so that it rotates at half the speed of the crankshaft (1:2). This means that the crankshaft must make two complete revolutions before the camshaft will complete one revolution. The following section will describe a four-stroke, normally aspirated, diesel engine having both intake and exhaust valves with a 3.5-inch bore and 4-inch stroke with a 16:1 compression ratio, as it passes through one complete cycle. We will start on the intake stroke. All the timing marks given are generic and will vary from engine to engine. Refer to Figures 10, 16, and 17 during the following discussion.

Intake
As the piston moves upward and approaches 28° before top dead center (BTDC), as measured by crankshaft rotation, the camshaft lobe starts to lift the cam follower.

This causes the pushrod to move upward and pivots the rocker arm on the rocker arm shaft. As the valve lash is taken up, the rocker arm pushes the intake valve
downward and the valve starts to open. The intake stroke now starts while the exhaust valve is still open.

The flow of the exhaust gasses will have created a low pressure condition within the cylinder and will help pull in the fresh air charge as shown in Figure 16.

The piston continues its upward travel through top dead center (TDC) while fresh air
enters and exhaust gasses leave. At about 12° after top dead center (ATDC), the
camshaft exhaust lobe rotates so that the exhaust valve will start to close. The valve is fully closed at 23° ATDC. This is accomplished through the valve spring, which was
compressed when the valve was opened, forcing the rocker arm and cam follower back against the cam lobe as it rotates. The time frame during which both the intake and exhaust valves are open is called valve overlap (51° of overlap in this example) and is necessary to allow the fresh air to help scavenge (remove) the spent exhaust gasses and cool the cylinder. In most engines, 30 to 50 times cylinder volume is scavenged through the cylinder during overlap. This excess cool air also provides the necessary cooling effect on the engine parts.

As the piston passes TDC and begins to travel down the cylinder bore, the movement of the piston creates a suction and continues to draw fresh air into the cylinder.

Compression
At 35° after bottom dead center (ABDC), the intake Figure 17 Compression valve starts to close. At 43° ABDC (or 137° BTDC), the intake valve is on its seat and is fully closed. At this point the air charge is at normal pressure (14.7 psia) and ambient air temperature (~80°F), as illustrated in Figure 17.
At about 70° BTDC, the piston has traveled about 2.125 inches, or about half of its stroke, thus reducing the volume in the cylinder by half. The temperature has now
doubled to ~160°F and pressure is ~34 psia.

At about 43° BTDC the piston has traveled upward 3.062 inches of its stroke and the volume is once again halved. Consequently, the temperature again doubles to about
320°F and pressure is ~85 psia. When the piston has traveled to 3.530 inches of its stroke the volume is again halved and temperature reaches ~640°F and pressure 277 psia. When the piston has traveled to 3.757 inches of its stroke, or the volume is again halved, the temperature climbs to 1280°F and pressure reaches 742 psia. With a piston area of 9.616 in2 the pressure in the cylinder is exerting a force of approximately 7135 lb. or 3-1/2 tons of force.

The above numbers are ideal and provide a good example of what is occurring in an
engine during compression. In an actual engine, pressures reach only about 690 psia. This is due primarily to the heat loss to the surrounding engine parts.

Fuel Injection
Fuel in a liquid state is injected into the cylinder at a precise time and rate to ensure that the combustion pressure is forced on the piston neither too early nor too late, as shown in Figure 18. The fuel enters the cylinder where the heated compressed air is present; however, it will only burn when it is in a vaporized state (attained through the addition of heat to cause vaporization) and intimately mixed with a supply of oxygen.

The first minute droplets of fuel enter the combustion chamber and are quickly vaporized. The vaporization of the fuel causes the air surrounding the fuel to cool and it requires time for the air to reheat sufficiently to ignite the vaporized fuel. But once ignition has started, the additional heat from combustion helps to further vaporize the new fuel entering the chamber, as long as oxygen is present. Fuel injection starts at 28° BTDC and ends at 3° ATDC; therefore, fuel is injected for a duration of 31°.

Power
Both valves are closed, and the fresh air charge has been compressed. The fuel has been injected and is starting to burn. After the piston passes TDC, heat is rapidly released by the ignition of the fuel, causing a rise in cylinder pressure. Combustion
temperatures are around 2336°F. This rise in pressure forces the piston downward and increases the force on the crankshaft for the power stroke as illustrated in Figure 19. The energy generated by the combustion process is not all harnessed. In a two stroke diesel engine, only about 38% of the generated power is harnessed to do work, about 30% is wasted in the form of heat rejected to the cooling system, and about 32% in the form of heat is rejected out the exhaust.

In comparison, the four-stroke diesel engine has a thermal distribution of 42% converted to useful work, 28% heat rejected to the cooling system, and 30% heat rejected out the exhaust.

Exhaust
As the piston approaches 48° BBDC, the cam of the exhaust lobe starts to force the follower upward, causing the exhaust valve to lift off its seat. As shown in Figure 20, the exhaust gasses start to flow out the exhaust valve due to cylinder pressure and into the exhaust manifold. After passing BDC, the piston moves upward and accelerates to its maximum speed at 63° BTDC. From this point on the piston is decelerating. As the piston speed slows down, the velocity of the gasses flowing out
of the cylinder creates a pressure slightly lower than atmospheric pressure. At 28° BTDC, the intake valve opens and the cycle starts again.