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High Powered Variable Speed Fluid Drives are used to provide power to Boiler Feed Pumps and Fans. Each of these Fluid Drives has an input shaft assembly that rotates at the speed of the input power source, which may be an electric motor or a steam turbine-generator that operates at constant speed. The output shaft assembly rotates at a variable speed, typically controllable from 20% to 97.5% of the input shaft speed.

The typical design of a variable speed fluid drive uses an input rotating assembly that is coupled to and rotates with the speed of the driver (motor or turbine-generator). The output rotating assembly is connected to and rotates with the driven load, a pump, fan, or compressor.

The input rotating assembly consists of an input coupling, a shaft supported by the two journal bearings and a thrust bearing, and a fluid drive element that consists of an impeller and a casing that surrounds the impeller and a runner, all rotating together.


The output rotating element consists of a runner, an output shaft supported by two journal bearings and a thrust bearing, and an output coupling hub.

The impeller and runner are structures with vanes and pockets between the vanes shaped much like a half of a grapefruit with the linings that separate the fruit intact and the fruit removed.

The impeller rotates at the input shaft speed and the runner drives the output shaft, or pump (or fan) speed. The two rotating elements are connected hydraulically, but not mechanically. The speed of the output rotating assembly is controlled by hydraulic forces. A scoop tube is used to remove fluid from rotating assembly and to control the level of fluid in the element, thereby controlling the hydraulic forces between the impeller and runner. A scoop tube actuator, is driven by a signal from the plant control system, such as the central DCS (Distributed Control System), for the unit.

The casing that surrounds the impeller and runner is attached to the input rotating assembly so it rotates with the input shaft speed. The casing contains a fluid, usually an oil, being thrown outward by centrifugal force so that it is in the shape of a torus adjacent to the inside surface of the spinning casing. The thicker the torus of liquid that exists, the greater the coupling between the impeller and the runner. Oil, called “circuit oil”, is pumped into the spinning element, or “circuit” and is removed by a scoop tube at the rate that it is pumped in, removing heat from the process. The inside diameter of the torus is adjustable by the radial movement of the scoop tube as it removes circuit oil from the torus, thereby controlling the coupling between impeller and runner.

A fluid drive is a constant torque machine, which means that the torque transmitted by the output shaft to the load (pump or fan) is equal to the torque absorbed from the driver (motor or turbine) by the input shaft. Because there is always a difference in speed between the input shaft and output shaft, the power into the fluid drive is equal to the power out plus the power lost in the fluid drive process within the element. Power-in equals input speed times torque, Power-out equals output speed times torque, and power lost equals torque times the slip speed (input speed - output speed). It is the heat generated by the lost power that requires the circuit oil to be exchanged continuously.

The circuit oil flow pattern within the element consisting of the impeller pockets and runner pockets and the axial gap between them is extremely turbulent. For reference, for a fluid drive impeller of 25 to 27 inches in diameter, the impeller and runner pockets are about 6 inches deep axially and the gap is on the order of 1/4 inch. The pockets are larger or smaller in proportion to the size of the fluid drive element.

The turbulence of the circuit oil is affected by the speed difference, the torque transmitted, and the amount of circuit oil in the rotating element. The speed at which the greatest turbulence occurs is when the output shaft is at approximately 2/3 of the input shaft speed. This is in the range of 2400 rpm for a fluid drive input speed of 3600 rpm. The turbulence at a given speed increases with increased torque at that speed. Generally, when the circuit oil is the most turbulent, the fluid drive is generating the highest amount of heat and vibration in the element.

Four performance charts are presented that are based on the head-capacity plot for a main boiler feed pump for a 375 MW Turbine-Generator. The curves on the charts are made of constant speed lines (generally curving down to the right). The fluid drive power loss chart also has curves of constant efficiency lines (generally curving upward). The grids are the same on all of the charts. Data is presented at the intersection points of the grids.

Fluid Drive Chart 1
Chart 1 presents the shaft power absorbed by the pump (output of the fluid drive)

Fluid Drive Chart 2
Chart 2 presents the shaft power absorbed by the fluid drive (out of the motor or T-G)
at the same intersection points of the same grid.

Fluid Drive Chart 3
Chart 3 presents the power lost to the circuit oil in
the element (power into the fluid drive minus power out of the fluid drive).

Fluid Drive Chart 4
Chart 4 presents the circuit oil flow rate required to maintain a
constant temperature rise of the circuit oil
(entering the fluid drive to leaving the fluid drive element).

For instance, following the constant speed line for 2400 rpm from left to right, it can be seen in these charts that power transmission increases, the torque increases, the losses increase, and the circuit oil flow rate required to maintain a constant temperature rise increases. All of this is associated with increased turbulence of the circuit oil when traversing the constant speed line going from low pump flow to higher pump flow.

A Size 270 Single Circuit Fluid Drive has a 27.0 inch diameter circuit, and it holds approximately 15 gallons, or 105 lbs) of oil when the output shaft speed is in the 2400 - 3000 rpm range. The radial acceleration that the oil experiences can be measured in the number of “g”s by dividing the acceleration by the gravitational acceleration constant (386 in/sec^2). At 3600 rpm, and at a diameter of 27 inches, the “g field” is approximately 4600.

To explain the features of the vibratory forces acting on the impeller and casing as well as on the runner, consider the following:

  • The impeller is not perfectly uniform in shape from pocket to pocket and
    neither is the runner;
  • The impeller and runner have some vibration and therefore, for each one,
    the center of rotation is not the geometric center of the respective part so
    that the radial distance between their centers varies continually between 0
    and 0.005 inches, or more;
  • The circuit oil exiting the impeller and entering the runner near the OD
    has a frequency of the impeller (input);
  • The circuit oil exiting the runner and entering the impeller near the ID
    has the frequency of the runner (output); and
  • The circuit oil exiting the gap between the impeller and runner and
    passing between the input casing on the outside and the back of the runner
    on the inside has a mixed flow influenced by both the runner and the
    impeller positions and hence their speeds.

Under these circumstances, it is very likely that the oil flows can be very turbulent and non-uniform circumferentially. With over 100 lbs, or over 1600 ounces, of circuit oil, for every 1 ounce of oil that is out of balance, a radial force of approximately 300 lbs is exerted radially on the inside of the input casing. If there are 5 ounces of oil that are unbalanced, then the radial unbalance force is on the order of 1500 lbs.

Given that the entire input rotating element weighs only 1500 lbs, and it is the element carrying the circuit oil, the input rotating element can be expected to respond to these forces by vibrating at the frequencies of excitation: Input shaft frequency, output shaft frequency, and a mixed frequency that is usually slightly above the output shaft frequency.

The runner has a vaned structure on the inside and has a smooth surface on the outside. There is a gap between the outside runner surface and the inside of the input casing that contains circuit oil. Both the vaned structure on the inside of the runner and the smooth surface on the outside of the runner are exposed to circuit oil, and the pressures in the oil on both sides are proportional to each other. Consequently, the runner does not experience the radial force levels that the impeller and input casing experience due to unbalanced oil caused by turbulence.

However, for both the impeller and casing assembly as well as for the runner, the following applies: The greater the turbulence in the circuit oil, the greater the unbalance of the oil, the greater the radial forces on the impeller and casing as well as for the runner, and the higher the vibration of the rotating assemblies, particularly the input assembly. It also follows, that the greater the amplitudes of vibration of both the input and the output shafts, the greater the distances between their respective centers, the greater will be the non-uniform flows of circuit oil from each one to the other and the greater will be the unbalanced amount of circuit oil.

The radial forces acting on the impeller and casing and on the runner by the circuit oil are transferred to the respective input and output shaft assemblies. Each rotating assembly is supported by two journal bearings and a thrust bearing. The journal bearings resist the radial forces applied to the respective rotating assembly.

The two journal bearings, known as the input-outboard and the input-inboard bearing, support the input shaft. Two journal bearings, known as the output-inboard and the output-outboard support the output shaft. The input-inboard journal bearing is the bearing that primarily supports the impeller and input casing as well as the spinning circuit oil, and therefore, it is the one that is subjected to the greatest force levels due to the dynamic forces of mechanical unbalance and unbalance of the turbulent circuit oil as described above, as well as the static weight of the inboard end of the input rotating assembly.

In the Size 270 Single Circuit Fluid Drive, the centerlines of the journal bearings are approximately 25 inches apart. The impeller is overhung from the input-inboard bearing by about 12 inches, and the runner is overhung from the output-inboard bearing by approximately 18 inches.

The original pressure dam journal bearings were relatively small bearings, being 7 inches in diameter and about 2.75 inches long. These bearings were generally adequate to resist the mechanical unbalance forces of a well balanced fluid drive and they could resist a minimum level of radial forces due to the unbalanced circuit oil flow when the boiler feed pump was operated in the 2300 to 2700 rpm speed range, such as would occur during start-up - if the feed pump flow was kept at the very minimum permitted flow, or if the start-up pump were used to bring the boiler to full pressure and then the transfer to the main pump was made.

However, most operators would not operate the unit this way. The typical start-up is to get on the main pump as soon as is practical, and while limiting the pressure to the silica limit, say 1200 psi at first, the flow would be increased to the maximum possible by operating the throttle valves to the Valves Wide Open condition. There are two benefits: The greater the flow, the faster the boiler water will clear and the faster the boiler pressure could be raised, and the second is that the power generated would be maximized during the water clean-up process.

Problems arise for the fluid drive when it is operated this way, even for a few hours during a start-up, because the output shaft speed is gradually raised and it traverses the 2300-2700 rpm speed range causing (a) high heat generation with the fluid drive getting quite hot, and (b) there is a considerable amount of turbulent circuit oil flow with consequent high rotating radial forces in the element that are applied to the bearings, particularly, the input-inboard journal bearing. The original small pressure dam bearings could not withstand the forces, and would be damaged causing increased bearing clearances. In turn, this would permit larger distances between the impeller and runner centers which lead to bigger unbalanced circuit oil flows which led to higher forces that the bearings had to resist. Even when the fluid drives were built with great precision, the original bearings, particularly the input-inboard bearing, would fail repeatedly.

Sample Vibration Measured at the Input-Inboard fluid drive journal

The root cause of the continued bearing failures was discovered via a considerable amount of testing and redesign work. Once it was realized that the radial forces resulted from unbalanced circuit oil flow in the element, and that they were torque dependent and an inherent part of the fluid drive process, then the solution became apparent: Build the rotor and journal bearing system so that it could easily resist these unbalance forces without degrading.

In 1984, using its high tech bearing design capabilities, TRI designed and manufactured a heavy duty input shaft and tilting pad journal bearing system that was extremely robust. The upgraded tilting pad journal bearings are the same as those that support an IP turbine rotor weighing some 40,000 lbs.

From the first start-up, this upgraded TRI Heavy Duty journal bearing system proved to be the solution to this vibration problem. Subsequently, the journal bearings for the output shaft were upgraded to tilting pad bearings. Typically, these fluid drives are run for approximately 10 years between planned disassembly for inspection, cleaning, and reconditioning.

The various frequencies that were observed in the vibration traces from the proximity probes that were installed in the original fluid drives, described above, can still be seen in the vibration traces of the upgraded fluid drives. The only difference is that the upgraded journal bearings can resist the forces on a continuous operational basis.